Advances in heat and mass transfer in the membrane-based dehumidifiers and liquid desiccant air dehumidification systems

International Journal of Heat and Mass Transfer 139 (2019) 881–906

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International Journal of Heat and Mass Transfer journal homepage: www.elsevier.com/locate/ijhmt

Review

Advances in heat and mass transfer in the membrane-based dehumidifiers and liquid desiccant air dehumidification systems Si-Min Huang, Wu-Zhi Yuan, Minlin Yang ⇑ Guangdong Provincial Key Laboratory of Distributed Energy Systems, Dongguan University of Technology, Dongguan 523808, China

a r t i c l e

i n f o

Article history: Received 3 January 2019 Received in revised form 19 April 2019 Accepted 21 May 2019 Available online 31 May 2019 Keywords: Parallel-plate membrane contactor Hollow fiber membrane contactor Heat and mass transfer Dehumidifier Membrane-based liquid desiccant air dehumidification

a b s t r a c t The membrane-based liquid desiccant air dehumidification method had been reviewed based on the momentum, thermal, and mass transports, the performance analysis, and the thermodynamics. The parallel-plate membrane dehumidifiers (PMC) and hollow fiber membrane dehumidifiers (HFMC) with the cross-flow, quasi-counter flow, and counter flow configurations were summarized. The adiabatic and internally-cooled membrane dehumidifiers were described. The single-stage and the multistage membrane-based liquid desiccant air dehumidification (M-LDAD) systems to enhance their performances were analyzed. The internally-cooled quasi-counter flow PMC of the side in and side out type and hexagonal type were the promising alternatives. The counter flow and cross-flow HFMC were suitable for the small and relatively large air flow rates, respectively. The inclined flow type should be avoided to prevent from the performance deterioration. The internally-cooled HFMC with the cooling coils inside the solution channels between the neighboring short sub-contactors might have better performances with at the expenses of increasing equipment and operating costs. Effects of the membrane deformations on the momentum, thermal, and mass transports both in the PMC and HFMC should be studied. The randomly distributions of the hollow fiber tubes should be avoided. The multistage M-LDAD systems based on the internally-cooled PMC and HFMC complementarily driven by the waste energy and renewable energy sources were the better choices in the practical application whose thermodynamic and dynamics should be investigated. Ó 2019 Elsevier Ltd. All rights reserved.

Contents 1. 2.

3.

Introduction . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Parallel-plate membrane dehumidifiers . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2.1. Cross-flow parallel-plate membrane dehumidifiers . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2.1.1. Fluid flow, heat and mass transports in the cross-flow membrane contactors . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2.1.2. Performance studies of the cross-flow membrane contactors . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2.2. Quasi-counter flow parallel-plate membrane dehumidifiers . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2.2.1. Fluid flow, heat, and mass transfer in the quasi-counter flow membrane contactors . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2.2.2. Performance studies of the quasi-counter flow membrane contactors . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2.3. Counter flow parallel-plate membrane dehumidifiers . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Hollow fiber membrane dehumidifiers. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 3.1. Counter flow hollow fiber membrane dehumidifiers . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 3.1.1. Heat and mass transfer inside membranes. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 3.1.2. Heat and mass transfer inside the tubes. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 3.1.3. Heat and mass transfer outside the tubes . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 3.2. Cross-flow hollow fiber membrane dehumidifiers . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 3.2.1. Free surface model . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 3.2.2. Periodic element model . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

⇑ Corresponding author. E-mail address: [email protected] (M. Yang). https://doi.org/10.1016/j.ijheatmasstransfer.2019.05.069 0017-9310/Ó 2019 Elsevier Ltd. All rights reserved.

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Nomenclature b/a B0 C COP CMI d EER f fRe H Dh j m* N NTU Nu r ro/rf Re Sh SL ST T xin x0

aspect ratios in the PMC viscous flow membrane morphology parameter (m2) concentration (%) coefficient of performance capacity matching index diameter (m) energy efficiency ratio friction factor product of friction factor and Reynolds number duct height (m) deformation height (m) Colburn j factor for heat transfer ratio of solution to air mass flow rate tube number number of heat transfer units Nusselt number radius (m) radius ratio Reynolds number Sherwood number longitudinal pitch (m) transverse pitch (m) temperature (K) entrance length (m) channel length (m)

Acronyms AMLDD adiabatic membrane liquid desiccant dehumidifiers CAC conventional air conditioning DOAS dedicated outdoor air system DOAS-CC dedicated outdoor air-chilled ceiling system ECOS evaporative cooled sportive dehumidification system EHFMTB elliptical hollow fiber membrane tube bank ERV energy recovery ventilator HFMC hollow fiber membrane contactors HPMC hexagonal parallel-plate membrane channels HFMTB hollow fiber membrane tube bank HVAC heating, ventilating, and air-conditioning IMLDD internally-cooled membrane-based liquid desiccant dehumidifiers

4.

5.

ISPMC

internally-cooled side in and side out parallel-plate membrane channels LAMEE liquid-to-air membrane energy exchanger LDAD liquid desiccant air dehumidification M-LDAC membrane-based liquid desiccant air-conditioning M-LDAD membrane-based liquid desiccant air dehumidification PMC parallel-plate membrane channels or contactors RAMEE run-round membrane energy exchangers SPMC side in and side out parallel-plate membrane channels or contactors STS solar thermal system Greek letters u packing fraction e effectiveness Subscripts a air c cool cooling C fully developed values under conjugate transport phenomena boundary condition cool cooling effectiveness deh dehumidification effectiveness f free surface fe feed air H values under uniform flux (heat/mass flux) boundary condition i inner in inlet le left m average o outer ri right s solution sw sweeping air tube cooling tubes T values under uniform value (temperature/concentration) boundary condition w water falling film, water vapor

3.3. Inclined flow hollow fiber membrane dehumidifiers . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Membrane-based liquid desiccant air dehumidification systems . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 4.1. Single-stage systems . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 4.1.1. Operational systems driven by pumps . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 4.1.2. Natural systems driven by solution convection . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 4.2. Multistage system . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Conclusions and perspectives . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Declaration of Competing Interest . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Acknowledgements . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

1. Introduction In 2013, the building energy consumption of China was 1.05– 1.128 billion ton of standard coal, which covered 28–30% of the energy consumption of the whole society [1]. With the economic development, the building energy consumption of China would account for as large as 30–35% of the energy consumption of the whole society in the year of 2030 [2]. It was obvious that the building energy conservation was an important component to insure the

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energy security of China, which was related to our country’s energy security and sustainable development. Air-conditioning energy consumption occupied 30–50% of the building energy consumption [3]. Obviously, it accounted for as large as 15% of the total energy consumption. The main reason of the power shortage was using air-conditioners in summer, especially in Guangdong province where the electricity consumption of the whole province was as large as 523.5 billion kWh in 2014 ranked the first in China. The electricity consumption of the air-conditioners exceeded

S.-M. Huang et al. / International Journal of Heat and Mass Transfer 139 (2019) 881–906

100 billion kWh [2]. The increasing electricity consumption led to global warming and ecological damage. Therefore the energy conservations of the air-conditioners turned to be the topmost priority among the construction energy consumption. Further, it already became an important component of the national energy strategy, which was related with the achievement of energy conservation and emission reduction in the construction field of our country. The temperature and humidity independent control airconditioning system was one of the key technologies to realize the energy conservation, which could reduce about 20–30% of the energy consumption compared with the conventional airconditioning system [3]. Our country was located in Eurasia facing the wide Pacific Ocean. The huge thermal differences between land and sea made the monsoon type climate be obvious in most zones of China (Guangdong province, Hong Kong, etc.), which were in particularly influenced by the warm and wet air current of ocean in spring and summer. Therefore the weather was hot and humid, which increased the humidity control load. In the energy consumption of the air-conditioning system, as large as 20–40% of energy consumption would be used to deal with the humidity load [3]. It was obvious that the humidity control system was one of the key parts in the temperature and humidity independent control air-conditioning system, which was an important link to achieve the energy conservation. Therefore it was rather significant for the building energy conservation to develop novel energy saving dehumidification technologies. Air dehumidification methods mainly included cooling dehumidification, rotary dehumidification, liquid desiccant air dehumidification (LDAD), etc. [3]. At present, though the cooling dehumidification was the uppermost method because of its maturity and reliable operation of the equipment, the heat could not be recovered. Further, the effectiveness was relatively low and the operation consumption was high. The rotary dehumidification was widely used in solid desiccant adsorption dehumidification technology because of its strong dehumidification ability. Further, the heat of the exhausted air could be recovered. However the regeneration temperature of the solid desiccant generally required more than 140 °C commonly driven by the electrical energy or steam energy, which was necessary to consume a great amount of high-grade energy. Further, the efficiency of the energy utilization was relatively low. The LDAD had several obvious advantages of high effectiveness, no liquid water condensation, easy separation between the dehumidification air and regeneration air, easy utilization of low-grade energy or renewable energy, etc. [3]. Therefore this technology developed rapidly after the first LDAD system proposed by Lof [4] with triethylene glycol as the absorbent, and driven by solar energy applying packed beds as the dehumidifier and regenerator [5–8]. However the serious defect of the liquid desiccant (slat solution) crossover problem encountered in the traditional packed bed dehumidifier and regenerator largely limited its applications in a great extent. It was because the salt solution had strong corrosivity, which could largely damage human’s health and indoor decorations. In order to overcome this problem, the membrane contactors were used as the dehumidifier to realize indirectly contacted LDAD [9]. The humid air and the solution streams were separated by membranes from each other, which could only allow the water vapor to penetrate, while other gases or liquid were prevented from penetrating [9–11]. Therefore the droplet entrainment shortage could be prevented. Further, Ge, et al. [12] conducted the heat and mass transfer performance comparisons between a direct-contact pecked bed and a liquid-to-air membrane energy exchanger (LAMEE) for LDAD. The LAMEE had 13 and 20% higher latent and total effectiveness, respectively, than the packed bed at the same air pressure drop. The packed bed had similar sensible effectiveness and up to 16% higher latent effectiveness than the LAMEE with the same heat and mass transfer area.

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Smaller air channel (
3 mm) could be designed to compensate the heat and mass transfer resistances of membrane and achieved similar or even better performance than the packed bed. Therefore the membrane-based liquid desiccant air dehumidification (MLDAD) might be a promising and environmentally-friendly air dehumidification technology. In addition, super hydrophobic modification of the air side membrane surface might be anther way. In the last two decades, the M-LDAD technology had a great development because of a vital advantage of preventing solution droplets by separating the solution and the air streams. However the heat and water vapor could be exchanged across the membranes effectively. Therefore the humid air could be dehumidified and cooled. The membranes might be parallel-plate type or hollow fiber type, which could be applied to form the parallel-plate membrane contactors (PMC) or hollow fiber membrane contactors (HFMC), respectively. The structures of the former and the latter ones were respectively presented in Fig. 1(a) and (b). As seen from Fig. 1(a), the flow channels were formed between neighboring membranes by keeping the equal spacing. The solution and the air streams flowed in the rectangular channels in a cross-flow, a counter-flow or a quasi-counter flow configuration. As depicted from Fig. 1(b), compared to the PMC, the HFMC were more difficult to construct because of the sealing conditions of the two ends. However the HFMC had some coherent advantages of larger packing densities, higher effectiveness, and smaller air side pressure drop [10]. The concepts were similar to the tube-and-shell heat exchangers where the air and the solution streams flowed on the respective sides of the hollow fibers. In the real applications, the processing air flowed in the shell side, while the solution flowed in the tube side. The PMC and HFMC could be used as dehumidifiers or regenerators for LDAD. However the single membrane contactors couldn’t be used in the practical air dehumidification processes. Some other devices such as air blowers, pumps, pipelines, heat exchangers, liquid desiccant containers, etc. should be connected with the membrane contactors to form a whole liquid dehumidification system. Different type systems were designed for the better performances. Compared with the traditional ones, this novel technology has lower energy consumption, higher dehumidification effectiveness, smaller volume, etc. Further, the membrane-based liquid desiccant air dehumidification (M-LDAD) systems could be complementarily driven by solar energy, industrial waste heat, electrical energy, etc. shown in Fig. 2, which had been designed in an invention patent [13]. The industrial waste heat from flue gases or water and renewable energy sources such as solar energy could be used for the regeneration of the diluted solution. The electrical heat could be used to make the M-LDAD system be continuously operated when the other heat sources were absent. The M-LDAD technology had been proposed about 20 years. However it developed rapidly in recent 10 years, which were mainly focused on two aspects: (1) Parallel-plate membrane and hollow fiber membrane contactors with counter flow, cross-flow, or quasi-counter flow configurations had been designed and applied for replacing the packed beds. Further, adiabatic and internally-cooled membrane contactors had been investigated. Fluid flow, conjugate heat and mass transports in the various membrane contactors had been studied; (2) Single-stage and multistage M-LDAD systems had been deigned to improve the system performances. They have been reviewed based on the aspects of feasibility, membrane materials, energy and economic benefits, environmental impacts, etc [14–16]. However this review would give an in-depth analysis of these studies considering the heat and mass transports of the membrane contactors and thermodynamic properties of the whole M-LDAD systems, which were rather important for the design and optimization of the membrane contactors and the M-LDAD systems. Further, some unsolved impor-

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Air out

2b 2b Membranes 2a Solution in Solution out

Duct sealing Air in

(a) Air in

Solution out

Solution in

Air out

Hollow fiber membranes

(b) Fig. 1. Structure of the membrane contactors [1]. (a) The parallel-plate membrane contactor (PMC); (b) The hollow fiber membrane contactor (HFMC).

Solar pond Waste heat Regenerator Electric heater Seawater desalination Solution container Consistency controller

Heat exchanger

Total heat recovery

Indoor environment

Membrane dehumidifier

Evaporative cooler

Fig. 2. The membrane-based liquid desiccant air dehumidification (M-LDAD) systems complementarily driven by solar energy, industrial waste heat, electrical energy, etc. [8].

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tant scientific and technical problems would be proposed. The development prospects of this novel technology would be expected and predicted. 2. Parallel-plate membrane dehumidifiers As mentioned, the membrane contactors could be employed as the dehumidifiers, which were the key modules in the M-LDAD systems. The air and the solution streams in the membrane contactors could be in a cross-flow, a counter flow, or a quasi-counter flow configuration. 2.1. Cross-flow parallel-plate membrane dehumidifiers 2.1.1. Fluid flow, heat and mass transports in the cross-flow membrane contactors 2.1.1.1. Adiabatic types. The cross-flow PMC were commonly used as the dehumidifiers due to their conveniences in duct sealings. Heat and mass transfer inside the membrane contactors were of importance for the contactor design and performance analysis. As we all know, the transport problems had been solved in classical textbooks. A constant temperature or a constant heat flux boundary condition was considered on the membrane surface for simplicity. However, in real situation, the problems became much more complex. The transports in the membrane contactors were conjugated between the air and the solution streams [10]. Therefore the boundary conditions were neither constant temperature nor constant heat flux condition on the membrane surface because of the large mass transfer Biot number (0.1) [10]. On contrary, the membrane surface conditions were naturally generated by the couplings between the neighboring fluids through the membranes, which were a conjugate heat and mass transfer problem [17]. The Nusselt numbers under ideal constant value or constant flux boundary conditions could not accurately reflect the heat and mass transports inside the membrane contactors. Based on the specificities of the heat and mass transports inside the membrane contactors, Zhang [18] obtained the real membrane surface conditions of the cross-flow PMC used for air-to-air energy recovery by the iterative solution of the coupled fluid flow, heat, and mass transfer equations between the neighboring air fluids through the membrane in the unit cell, which contained a membrane and two adjacent air channels. The fluids were assumed laminar flow and Newtonian with constant thermal properties. The fully developed local Nusselt number NuC was smaller than NuH. The fully developed Sherwood number ShC was nearly equal to the NuC. It should be noted that these results were calculated based on the assumption of the fluids being hydrodynamically developed, while thermally and concentrationally developing. However up to 70% of the channel length was in the entrance zone. The Nusselt and Sherwood numbers in the entrance zone were much larger than those in the developed zone. Therefore the assumption of the fully developed entrance could cause large underestimations. To address this problem, Zhang et al. [19] solved the conjugate transport problems in a cross-flow parallel-plate membrane contactor for the air-to-air energy recovery with a commercial CFD code in all entrance zones. The mass transfer problem was solved by transforming it to a heat transfer one. The Nusselt and Sherwood numbers were then obtained, which were more accurate for describing the fundamental transport phenomena in the membrane contactor. Deshko et al. [20] conducted the numerical investigation of the heat and mass transports, frost formation, and condensation conditions in the cross-flow membrane contactor. Local Nusselt numbers were calculated for parallel-plate channels without or with fins. Based on the solution approach of the conjugate heat and mass transports in the air-to-air energy recovery

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process, Huang et al. [21] obtained the Nusselt and Sherwood numbers under the conjugate transport boundary conditions in the cross-flow PMC for LDAD. The air side Sherwood numbers were a little larger than the Nusselt numbers. For the solution side, the heat transfer was fully developed near the entrance, but mass transfer was still in developing even at the outlet. The solution side Sherwood numbers were much larger than those for the air side. The solution side Nusselt numbers were about 15% higher than those for the air side. In the real applications, the channel lengths were limited due to the confinements in noises and pressure drops. The entrance lengths accounted for a relatively large proportion of the whole channel lengths. The performances would be underestimated under hydraulically fully developed flow, thermally, or concentrationally fully developed flow assumptions. Therefore Huang et al. [22] proposed a more accurate method to solve this problem. The governing equations were directly solved by considering the effects of the flow, heat, and mass transfer developing entrances. The (fRe), Nusselt and Sherwood numbers in the channels under various aspect ratios (b/a) were listed in Table 1. The NuC,a and NuC,s calculated considering all the entrance effects were about 10% larger than those given in Ref. [21]. The mean Nusselt numbers (Num,a and Num,s) were 20% higher than those listed in Ref. [21]. The Sherwood numbers also had the same variation trends, meaning that the thermally and concentrationally developing entrances both had relatively large influences on the transports in the channels. 2.1.1.2. Internally-cooled types. The membrane contactors could be the adiabatic and internally-cooled types according to whether there was a cooling element inside. In the adiabatic membranebased liquid desiccant dehumidifiers (AMLDD), the solution was heated due to the absorption and mixing heats. The performances of the dehumidifiers would be deteriorated because of the temperature increase (up to 3 °C) caused by the phase change energy [23]. In order to overcome this problem, the internally-cooled membrane-based liquid desiccant dehumidifiers (IMLDD) had been designed and used for LDAD [17] shown in Fig. 3. The solution and the feed air streams were in a cross-flow configuration. The water was showered into the cooling channels and falls vertically along the plates. The sweeping air flowed over the water falling films in a co-current flow configuration. Sensible heat was exchanged between the solution and the water, while the water was evaporated by sweeping air. The conjugate transport model between the solution, the feed air, the water falling film, and the sweeping air streams were numerically solved. The (fRe), Nusselt and Sherwood numbers for channels under various aspect ratios (H/dfe) were obtained and listed in Table 2. As seen, the NuC,fe, ShC,fe, and NuC,s for the IMLDD were a little (about 2–3%) smaller than those for the AMLDD. 2.1.1.3. Effects of the membrane deformations. The above-mentioned studies were focused on the contactors the parallel-plate membranes, which were the common types. However they might be deformed due to fluid pressure and weak mechanical strengths of the membranes. Huang [24] studied the influences of the aspect ratios (b/a), deformation heights (Dh), and Reynolds numbers (Re) on the (fRe)m and Num. When b/a were equal to or less than 25, the Num increased with an increase in the Dh for the air channel. However the Num increased firstly, and then decreased with an increase in the Dh with the b/a of 30–40. It could be easily found that the membrane deformation had a large effect on the performance of the membrane contactors. It should be noted that only the fluid flow and heat transfer under a uniform temperature condition with five given deformation heights were studied. Actually, the coupling of the fluids and solids (membranes) should be considered to obtain the real deformation heights in the practical applications.

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Table 1 Fully developed and mean (fRe), Nusselt and Sherwood numbers in the channels under various aspect ratios (b/a) [22]. b/a Sources

(fRe)C [22]

(fRe)m [22]

NuH [10]

NuT [10]

NuC,a [21]

NuC,a [22]

Num,a [22]

NuC,s [21]

NuC,s [22]

Num,s [22]

ShC,a [21]

ShC,a [22]

Shm,a [22]

Shm,s [22]

1.0 1.43 2.0 3.0 4.0 8.0 50.0 100.0 1

56.62 58.50 61.81 68.42 72.55 81.76 92.38 93.89 –

66.12 69.44 70.31 78.92 83.05 93.26 104.88 107.39 –

3.61 3.73 4.12 4.79 5.33 6.49 7.84 8.09 8.23

2.98 3.08 3.39 3.96 4.44 5.60 7.15 7.36 7.54

3.12 3.23 3.48 4.15 4.61 5.79 7.54 7.70 –

3.37 3.48 3.73 4.39 4.86 6.04 7.79 7.95 –

3.87 3.99 4.25 4.95 5.34 6.58 8.32 8.49 –

3.41 3.64 4.05 4.74 5.35 6.41 7.91 8.08 –

3.64 3.78 4.19 4.88 5.49 6.55 8.02 8.22 –

3.95 4.18 4.59 5.28 5.89 6.95 8.45 8.62 –

3.30 3.41 3.65 4.28 5.11 5.83 7.74 7.98 –

3.52 3.61 3.85 4.48 5.31 6.03 7.94 8.18 –

3.93 4.01 4.25 4.88 5.71 6.43 8.34 8.58 –

8.54 8.75 9.15 10.55 12.29 13.86 18.08 18.68 –

Water in

Sweep air in

Plastic plate

Solution stream

Membrane

Feed air in Feed air in

Solution in

Dehumidified air out Sweep air in

Water falling film

Channel sealing Solution out

(b)

Exhausted air out (a)

Fig. 3. Structure of a cross-flow internally-cooled membrane-based liquid desiccant dehumidifier (IMLDD) [17]. (a) Space diagram; (b) Planform.

Table 2 Fully developed (fRe), Nusselt and Sherwood numbers for the feed air, the solution, the cooling channels, water falling film and the sweep air stream [17]. H/dfe

(fRe)

NuT

NuH

NuC,fe

ShC,fe

NuC,s

Shm,s

NuC,w

NuC,sw

ShC,sw

20.0 40.0 50.0 70.0 80.0 100.0

89.56 92.38 93.05 93.66 94.01 94.42

6.72 7.14 7.23 7.32 7.35 7.39

7.30 7.71 7.84 8.00 8.04 8.09

7.02 7.22 7.31 7.40 7.43 7.47

7.08 7.44 7.53 7.61 7.65 7.68

7.23 7.64 7.74 7.85 7.89 7.92

18.64 16.08 15.18 13.88 13.38 12.63

7.64 7.67 7.71 7.73 7.74 7.75

7.29 7.34 7.36 7.38 7.42 7.44

7.39 7.43 7.46 7.48 7.51 7.55

Further, effects of the deformation on the conjugate transports in the deformed channels should be considered in future. 2.1.2. Performance studies of the cross-flow membrane contactors In the cross-flow membrane contactors, the conjugate transport data in the channels had been obtained. The lumped heat and mass transfer models could be solved based on the conjugate transport data. Performance analysis and parametric optimizations of the membrane contactors could be performed under various structural parameters, operation conditions and membrane properties. Sebai et al. [25] developed a numerical lumped parameter model to study the transports for the cross-flow PMC for air-to-air energy recovery. They found that the direction for decreasing heat and mass fluxes was seen to be parallel to the diagonal line of the membrane. Das and Jain [26] developed a steady-state model for the semi-permeable membrane-based air-to-air indirect contactors. Parametric analysis had been conducted to disclose the influ-

ences of contactor structures, membrane characteristics, flow rates, solution concentration, and ambient conditions on the performances, which were depended significantly mainly on the membrane features of the porosity, pore size, and thickness. Yaïci et al. [27] investigated the performances of the cross-flow membrane contactors as energy recovery ventilators. It could be found that the effectiveness decreased noticeably and slowly with the increased in membrane spacing and thickness. The latent effectiveness increased significantly with the diffusivity of water in the membrane. The outdoor temperature and humidity had only minor effects on contactor performances. For the membrane contactors used for air humidification, Sabharwal et al. [28] presented a two-dimensional steady-state model in the cross-flow air/water membrane contactor for air humidification. Further, the sensitivity analysis of the model was performed under various operating conditions and geometric parameters. The water transfer rate increased with the velocities at the wet side inlet pressure, wet

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side inlet humidity, dry side inlet temperature, and the membrane numbers increasing. The experimental and numerical studies of the transports in the cross-flow air-to-air membrane contactors for energy recovery or air humidification had been conducted. For LDAD, Gao et al. [29] studied the influences of the desiccant and air inlet parameters on the dehumidifier performances. The moisture effectiveness was mainly influenced by the solution flow and air flow rates. As we known, the membrane properties had tremendous influences on the transport performances of the membrane contactors. The membrane parameters contained the thermal conductivity, moisture diffusivity, maximum moisture uptake, moisture sorption capacity, etc. The sensible effectiveness showed a small change with the membrane parameters. However the latent effectiveness changed more significantly with the membrane parameters. Further, the enthalpy effectiveness was closer to the latent effectiveness [30]. Zhang et al. [31] studied three different membrane materials. Both the numerical and experimental results indicated that the moisture resistance through membranes was codetermined by the thickness, sorption slope, and sorption potential. Moisture diffusivities in various materials were in the same order. When the membrane plate was fixed, the latent performance increased with an increase in the sorption slopes. Xing et al. [32] applied a novel thin flat sheet zeolite membrane for air dehumidification. They achieved a moisture/air separation factor and a water permeance as large as 317 and 6.8  10-6 mol m2 1 s Pa1, respectively. Bai et al. [33] assessed the effects of the operating parameters on dehumidification performances, which included the number of heat transfer units (NTU), solution inlet temperature (Ts,in), ratio of solution to air mass flow rates (m*), and concentration (Cs). The sensible, latent, and enthalpy effectiveness increased with an increase in the m* and NTU. The latent and enthalpy effectiveness increased with the Cs increasing, but the sensible effectiveness had no change. The larger the solution inlet temperature was, the larger the effectiveness was. 2.2. Quasi-counter flow parallel-plate membrane dehumidifiers 2.2.1. Fluid flow, heat, and mass transfer in the quasi-counter flow membrane contactors 2.2.1.1. Hexagonal parallel-plate membrane channels (HPMC). The cross-flow PMC was a typical structure in the real applications. The counter-flow contactor might have larger performance compared to the cross-flow one. But a contactor with the pure counter flow configuration was rather difficult in channel sealing between the neighboring fluids. Therefore a membrane contactor with a quasi-counter flow configuration might be a better choice. Zhang [34] proposed a quasi-counter flow parallel-plate air-to-air total heat exchanger for energy recovery from the exhausted air, as presented in Fig. 4(a), which was similar to a heat and mass exchanger containing a series of hexagonal parallel-plate membrane channels (HPMC). The feed and the exhausted air streams flowed in the neighboring channels with a combination configuration of the cross-flow and the counter flow configurations (quasi-counter flow). The equations describing the three-dimensional momentum, heat, and mass transports were numerically solved. The sensible effectiveness for the NTU, m*, Ts,in, and Cs for the HPMC lay between those for the pure-counter flow and those for the cross flow ones. The flow could be considered to be 3 domains: two cross-like domains and a counter flow domain. The less the cross-flow domains were, the larger the counter flow domain was, and the larger the effectiveness were. Al-Waked et al. [35] conducted the similar studied of the HPMC for air-to-air total heat recovery. The lateral mesh element length had minimal effect on the performance. Liu et al. [36] set up a test rig to study the performance of the HPMC for air-to-air energy recovery in relatively cold cli-

Air out

Air or solution in

Hexagonal membranes Air in

Duct sealings

Air or solution out

(a) Solution out Cooling water in Cooling tubes

Duct sealing

Plate-type membrane

Air in Air out

H

Air in

Solution in Cooling water out

L

S

(b) Fig. 4. Structure of the quasi-counter flow hexagonal parallel-plate membrane contactor (HPMC). (a) No cooling tubes [34–37]; (b) With Z-shaped cooling tubes in the solution side [38].

mates. The sensible effectiveness was insensitive to the operating conditions under various air flow rates, while the latent effectiveness was sensitive the operation conditions in specific larger air flow rates. Huang et al. [37] used the similar model with Zhang [34]. However it was an air-solution type. The influences of the fluid parameters and deformation heights (2H) on heat transfer and fluid flow in the HPMC were investigated based on the assumptions of impermeable membranes and a uniform temperature condition. When the Re and fluids were fixed, the larger the 2H were, the smaller the fm were. For the fixed Re and 2H, the Num for the LiCl solution stream were 1.51–2.85 and 1.21–1.38 times of those for the air and the water streams, respectively. In order to take away the absorption and mixing heats generated by absorbing water vapor from the air stream on the solution, Qiu et al. [38] used an internally-cooled hexagonal plate membrane contactor, which contained a series of internally-cooled hexagonal parallel-plate membrane channels (IHPMC) for the solution stream depicted in Fig. 4(b). The (fRe)m and Num for the solution stream under a uniform temperature boundary condition with various N and do under various Re were shown in Table 3. As seen, when the do was equal to 0.002 m, the (fRe)m increased with the N increasing, while the Num decreased with an increase in the N. When the N was fixed as 3, the (fRe)m increased with an increase in the do, and the Num decreased with an increase in the do. 2.2.1.2. Side in and side out parallel-plate membrane channels (SPMC). The above-mentioned quasi-counter flow membrane contactors were like the hexagonal heat and mass exchangers, where the neighboring fluids were in a quasi-counter flow configuration

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Table 3 Mean (fRe)m and Nusselt number (Num) for the solution stream flowing through the IHPMC with various tube number (N) and tube outer diameters (do) under various Reynolds numbers (Re), L = 0.035 m, S = 0.05 m, H = 0.005 m [38]. N or do; Re?

(fRe)m 21.01

Num

(fRe)m 105.05

Num

(fRe)m 210.10

Num

(fRe)m 315.14

Num

(fRe)m 420.19

Num

N

0

50.56

10.81

51.85

16.37

52.96

21.12

53.60

25.04

54.03

28.55

N

1 2 3 4 5

59.45 67.68 78.71 90.73 106.15

33.94 30.43 26.45 24.66 21.89

65.48 76.72 89.93 103.79 119.94

40.00 39.39 38.36 37.95 36.71

72.49 87.37 103.11 118.78 135.98

49.99 49.68 48.84 48.38 47.18

78.48 96.04 113.37 130.32 148.22

58.37 58.55 57.93 57.18 55.81

83.62 103.30 121.65 139.58 157.98

65.96 66.49 65.94 64.73 63.12

0.001 0.002 0.003 0.004

67.95 78.71 88.35 98.70

32.55 26.45 20.64 16.14

75.67 89.93 102.72 113.20

39.50 38.36 36.27 26.10

83.60 103.11 121.64 132.67

49.60 48.84 50.00 38.19

89.75 113.37 138.59 152.35

58.23 57.93 59.66 50.36

94.86 121.65 153.32 171.82

65.73 65.94 67.56 61.46

(do = 0.002 m)

do (N = 3)

formed in the hexagonal structure. However the quasi-counter flow configuration could also be realized in a side in and side out structure. Huang et al. [39] proposed another structure of the side in and side out parallel-plate membrane channels (SPMC), as shown in Fig. 5(a). The air side and the solution side channels were the same. The fluids entered from the right headers of the channels and leaved from the left headers. The governing equations were obtained and solved under a uniform temperature membrane surface condition. The (fRe)m and Num for the water stream inside the

SPMC under various entrance ratios (xin/x0) and Re were presented in Table 4. Both the (fRe)m and Num increased with an increase in the Re. Yang et al. [40] set up a test rig to study the performances of the SPMC for air humidification. The performances of the SPMC increased with the water flow rate and temperature increasing. For the SPMC, Zhang et al. [41] proposed a structure that was used for realizing LDAD, as schematically depicted in Fig. 5(b). The air stream flowed uniformly in a straight path, while the solution stream entered from the right header and left from the left header.

Fig. 5. Structure of the quasi-counter flow PMC. (a) Side in and side out type without cooling tubes [39]; (b) With Z-shaped cooling tubes in the solution side [41]; (c) Used for liquid desiccant air dehumidification [41].

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Table 4 Mean (fRe) and Nusselt number (Num) for the water stream flowing through the SPMC under various entrance ratios (xin/x0) and Reynolds numbers (Re), x0 = y0 = 0.1 m, H = 0.005 m (x0/H = 20) [39]. xin/x0;

(fRe)m

Re?

50

0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 1.0

45.01 58.02 71.02 76.37 81.72 83.61 85.49 87.23 88.96 93.35

Num

(fRe)m

Num

100 5.78 6.16 6.48 6.78 7.08 7.28 7.47 7.75 8.03 8.09

(fRe)m

Num

(fRe)m

200

51.46 68.05 79.28 84.75 90.22 91.46 92.69 93.16 93.63 96.92

6.14 6.64 6.95 7.35 7.74 7.96 8.18 8.62 9.02 8.97

64.94 83.57 96.25 101.95 107.65 107.05 106.45 104.17 101.88 103.99

Num

(fRe)m

300 6.20 6.88 7.35 7.99 8.62 8.93 9.23 9.95 10.66 10.38

Num

(fRe)m

400

78.47 98.96 113.67 119.24 124.81 122.87 118.92 114.23 109.54 110.87

6.25 7.01 7.57 8.51 9.44 9.78 10.11 11.07 12.02 11.49

Num

500

91.53 114.69 131.11 136.22 141.33 136.87 130.41 123.71 117.01 117.59

6.43 7.05 7.84 9.02 10.19 10.55 10.91 12.07 13.22 12.39

105.33 130.36 148.05 152.36 156.66 149.24 141.82 133.09 124.35 124.17

6.61 7.26 8.18 9.57 10.95 11.32 11.65 13.56 14.31 13.17

Table 5 Mean (fRe), Nusselt number (Num) and Sherwood number (Shm) under the uniform temperature boundary condition for the air and the solution fluids flowing through the SPMC under various aspect ratios (x0/H), x0 = y0 = 0.1 m, xin/x0 = 0.2, (Re)a = 500, (Re)s = 100 [41]. x0/H;

(fRe)m,a

Fluids?

Air streams

20 30 50 75 100

127.32 117.64 107.39 103.19 101.39

(NuT)m,a

(NuH)m,a

(NuC)m,a

(ShC)m,a

(fRe)m,s

(NuT)m,s

(NuH)m,s

(NuC)m,s

(ShC)m,s

11.31 11.37 10.95 10.59 10.26

11.36 11.16 10.66 10.49 10.39

28.20 27.44 23.31 20.95 19.39

LiCl solution streams 8.30 8.00 7.71 7.60 7.56

10.94 10.63 10.37 10.35 10.15

10.83 9.78 8.92 8.68 8.64

10.90 10.38 9.82 9.80 9.84

The momentum, heat, and mass conservation equations describing the conjugate transport phenomena were solved. The (fRe)m, Num, and Shm under the conjugate transport boundary conditions with various length-width ratios (x0/y0), length-height ratios (x0/H), and entry ratios (xin/x0) were listed in Tables 6 and 7. As seen, the air side and the solution side Num and Shm in the quasicounter flow were nearly greater than those for the cross-flow. For the small x0/y0 for the air stream or large x0/y0 for the solution stream, the Num and Shm for the quasi-counter flow were greater than those for the counter flow (see Table 5). For the LDAD process, the solution temperature would increase because of the absorption and mixing heats, which could largely reduce the moisture absorption performance. An internallycooled parallel-plate membrane contactor formed by a series of the internally-cooled side in and side out parallel-plate membrane channels (ISPMC) shown in Fig. 5(c) had been designed to solve the problem on the basis of the former type [42,43]. The cooling tubes were populated in the solution channels. The solution and the cooling water streams flowed in a counter flow configuration. The (fRe)m, Nusselt numbers and Sherwood numbers under the uniform temperature boundary condition and under the conjugate heat and mass transfer boundary conditions are obtained and listed in Tables 7 and 8, respectively [42,43]. Effects of the tube numbers, tube outer diameters, Reynolds numbers, and various

74.79 66.95 61.44 57.86 56.91

8.19 7.81 7.47 7.15 7.02

tube shapes on the basic data are analyzed. It can be found that both the tube numbers and the tube outer diameters have small influences on the mean Nusselt numbers and Sherwood numbers for the air stream, while they have large effects on those for the solution stream. They would be useful for the structural design and performance optimization for the ISPMC used for air/solution/water heat and mass exchanging. 2.2.2. Performance studies of the quasi-counter flow membrane contactors Besides for the air humidification and dehumidification, the SPMC could also be employed as the air-solution run-round membrane energy exchangers (RAMEE) [44,45], which actually were the heat and mass exchangers between the air and the solution streams in the combination of the counter flow and cross-flow configurations, as shown in Fig. 6(a). The energy recovery from the air was realized by absorbing the moisture from the air using the solution steam. The transport features between the membrane-based dehumidifiers and the RAMEE were the same, while their purposes were different. The former ones were for air dehumidification, while the latter ones were for energy recovery. Vali et al. [44] found that the phase change energy had a relatively influence in the temperature and humidity distributions. The aspect ratio and entrance ratio were recommended to be less than 0.2 and 0.1 in

Table 6 Mean (fRe), Nusselt number (Num) and Sherwood number (Shm) under the uniform temperature boundary condition for various fluids flowing through the SPMC under various entrance ratios (xin/x0), x0 = y0 = 0.1 m, H = 0.005 m (x0/H = 20), ma = 1.76 kg/h, ms = 21.9 kg/h [41]. xin/x0;

(NuC)m,a

Fluids?

Air streams

0.1 0.2 0.3 0.4 0.6 0.8 1.0

11.34 10.75 10.62 10.56 10.47 10.22 9.87

(Sh)m,a

(fRe)m,s

(NuT)m,s

(NuH)m,s

(NuC)m,s

(ShC)m,s

7.93 8.36 8.47 8.51 8.63 8.80 8.97

11.43 11.53 11.56 11.56 11.65 11.69 11.75

11.62 11.60 11.56 11.54 11.48 11.45 11.42

28.28 29.12 29.28 29.27 29.36 29.82 30.52

LiCl solution streams 10.97 10.88 10.85 10.84 10.83 10.76 10.67

88.94 77.60 78.80 80.76 82.61 84.83 94.25

Notes: (fRe)m,a = 127.32, (NuT)m,a = 8.30, (NuH)m,a = 10.94. The air channel structure is not be changed with the entrance ratios (xin/x0).

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Table 7 Mean (fRe), Nusselt numbers (Num) and Sherwood numbers (Shm) under the conjugate heat and mass transfer boundary condition for the air stream flowing through the ISPMC with various tube numbers (N) under various Reynolds numbers (Rea), x0 = y0 = 0.1 m, xin/x0 = 0.2, H = 0.005 m, Res = 600, mw = 1.22 kg/h. N;

(NuC)m,a

Rea?

100

0 3 4 5 6

7.30 7.18 7.16 7.10 7.14

(fRe)m,a (NuT)m,a (NuH)m,a

100.57 7.17 8.18

(ShC)m,a

(NuC)m,a

(ShC)m,a

(NuC)m,a

300 7.68 7.67 7.65 7.68 7.71

(ShC)m,a

500

8.20 7.83 7.83 7.75 7.81

8.35 8.33 8.31 8.35 8.44

9.10 8.47 8.50 8.41 8.49

114.96 7.70 8.82

(NuC)m,a

(ShC)m,a

700 8.90 8.97 8.95 9.03 9.22

128.44 8.11 9.51

10.06 9.21 9.17 9.07 9.21

9.48 9.60 9.59 9.71 10.00

140.68 8.51 10.21

Table 8 Mean (fRe), Nusselt numbers (Num) and Sherwood numbers (Shm) under the conjugate heat and mass transfer boundary condition for the solution stream flowing through the ISPMC with various tube numbers (N) under various Reynolds numbers (Res), x0 = y0 = 0.1 m, xin/x0 = 0.2, H = 0.005 m, Rea = 500, mw = 1.22 kg/h. N;

(fRe)m,s

Res?

11.42

0 3 4 5 6

57.51 177.09 256.11 333.78 396.80

(NuC)m,s

(ShC)m,s

(fRe)m,s

(NuC)m,s

(ShC)m,s

34.26 8.52 11.72 10.97 12.00 16.14

15.70 22.22 24.15 26.35 39.42

61.77 191.77 271.79 352.31 419.74

(fRe)m,s

(NuC)m,s

(ShC)m,s

68.52 9.13 12.19 12.08 12.53 15.06

21.67 30.55 33.38 36.04 50.46

the SPMC, respectively. Optimization studies [45] showed that the RAMEE systems with the same supply and exhaust energy exchangers could obtain the highest performance in most conditions. The solution flow rate could be adjusted to enhance the annual energy recovery rate up to about 7%. Kassai [46] studied the effectiveness and humidification capacity of the RAMEE at winter air conditions with low heat capacity ratios. The humidification performance increased with the heat capacity ratio increasing. The highest performance (4.53 g/kg) could be obtained with the inlet air temperature of 1.7 °C. The RAMEE was also simply named as the liquid-to-air membrane energy exchanger, not to show its flow configuration. Moghaddam et al. [47] measured the steady-state performance of the small-scale single-channel LAMEE. The results showed that 138% increase of the convective heat transfer coefficient in the air channel with Rea = 1570 led to 11% increase in the LAMEE effectiveness. The latent effectiveness increased by about 11% with the moisture diffusion resistance of the membrane decreasing from 56 s/m to 24 s/m. Sabek et al. [48] examined the inlet fluids properties of the inlet fluids temperature, solution mass fraction of solution, and inlet air humidity ratio. A numerical solution was conducted to optimize the operating properties of the fluids to enhance the performance. Abdel-Salam et al. [49] established a numerical model to show the sensitivity of the steady-state performance of a flat-plate LAMEE to the air and solution channel widths. The optimum air channel width might be 5–6 mm and that for the solution one was 1–2 mm. Ge et al. [50] investigated the transport performance of the LAMEE employed as the dehumidifier and regenerator to disclose the effects of air and solution flow rates, solution temperature and concentration, and air temperature and humidity. Moghaddam et al. [51] presented the solution-side effectiveness for a single-panel small-scale LAMEE when it was used to regenerate the solution. The difference between the air side and the solution side effectiveness were neglected. Namvar et al. [52] experimentally studied the steady-state and transient performance of the LAMEE. The effectiveness increased with an increase in the solution mass flow rate or with the air mass flow rate decreased. The above-mentioned RAMEEs and LAMEEs were 2-fluid LAMEE (liquid to air types). The internally-cooled types of the 3-fluid

68.49 209.01 292.88 376.42 451.00

(fRe)m,s

(NuC)m,s

(ShC)m,s

11.28 15.47 15.76 16.85 20.20

31.15 43.93 47.62 50.14 61.71

114.19 10.06 13.71 13.86 14.44 18.77

26.66 37.71 43.66 43.86 57.05

77.54 229.96 318.68 405.60 490.61

LAMEE might have better performances compared to those of the 2-fluid types, which were similar to the IHPMC and ISPMC. Abdel-Salam et al. [53] proposed a novel 3-fluid LAMEE prototype shown in Fig. 6(b). Compared to a 2-fluid LAMEE, the cooling tubes were populated in the solution channels to control the solution temperature. The sensible, latent, and total effectiveness, sensible cooling performance, and moisture removal rate of the 3-fluid LAMEE were increased about 69%, 28%, 39%, 54%, and 140%, respectively. This structure could also be used in solution regeneration operating [54]. Compared to the 2-fluid LAMEE for solution regeneration, the sensible, latent, and total effectiveness, and moisture removal rate of the 3-fluid LAMEE were increased about 104%, 141%, 128%, and 17 times, respectively. The effect of the flow maldistribution caused by the membrane deformations on the performance of a 3-fluid LAMEE were investigated [55]. The effectiveness of the 3-fluid LAMEE might be decreased 7%. Based on the definitions of the traditional effectiveness equations of the 3-fluid LAMEE only considering the performance of the heat and mass transfer process between the air and the solution, the novel overall effectiveness equations were presented considering the heat and mass transfer between the air and the solution and the heat transfer between the cooling water and the solution, which were more universal for the 3-fluid LAMEE [56]. In future, the conjugate heat and mass transfer transports in the LAMEE under various membrane deformations should be studied, which might affect the fundamental data of the transport phenomena by the deformations. 2.3. Counter flow parallel-plate membrane dehumidifiers Besides the quasi-counter flow membrane contactors (HPMC, IHPMC, SPMC, or ISPMC), there also had the counter flow ones or approximate counter flow. Houreh and Afshari [57] established a three-dimensional model to study and compare the performances of the humidifiers with the parallel flow and counter flow configurations. The results showed that the water vapor and heat transfer with the counter flow were more than those of the parallel-flow, which led to a higher dew point at dry side outlet. An increased in temperature and a decreased in mass flow rate at dry side inlet

S.-M. Huang et al. / International Journal of Heat and Mass Transfer 139 (2019) 881–906

891

Fig. 6. Schematic of a counter-cross flat-plate LAMEE. (a) A 2-fluid LAMEE prototype [44,45]; (b) A 3-fluid LAMEE prototype [54].

led to a better humidifier performance. Sharqawy et al. [58] developed an effectiveness-mass transfer units (e-MTU) model for an ideal pressure retarded osmosis (PRO) membrane contactor. The concept was essentially a mass exchanger, which like a heat exchanger. The new e-MTU model could be used as a design tool for the PRO systems using a linearized osmotic pressure function and ideal membrane features. Cave and Merida [59] conducted an experimental study into the effects of the operating conditions on a single channel air-to-air membrane contactor based on the counter flow operation and the single-phase vapor-to-vapor. The heat loss to the surroundings affected the overall performance significantly. Lower flow rates led to larger outlet dew points of the receiver stream, which could be related with longer residence times. Based on the lumped transport model in the IMLDD with the cooling tubes inside [60] and the adjacently IMLDD [61] respectively shown in Fig. 5(b) and Fig. 3, the analytical effectiveness correlations were shown in Table 9, which could be used for performance analysis. The correlations were derived by considering the couplings between the air, the cooling water, and the solution streams through the membranes and the tubes. The IMLDD shown in Fig. 5(b) had been simplified as the pure counter flow

configuration for easy to obtain the analytical solution, which was reasonable due to as large as about 90% of the channel length being in a counter flow configuration [62]. Based on this method, Ge et al. [45] applied the analytical effectiveness correlations proposed in the HFMC [63] to study the performances of the RAMEE, indicating that the correlations were matched with the experimental results.

3. Hollow fiber membrane dehumidifiers 3.1. Counter flow hollow fiber membrane dehumidifiers As mentioned, the HFMC could also be used as the dehumidifiers, which usually were counter flow, cross-flow, and inclined flow. Each contactor was formed by a hollow fiber membrane tube bank (HFMTB) placed in an outer plastic shell. Besides air humidity control, the contactors were also used to seawater desalination, formaldehyde removal from air, CO2 capture, dialysis, etc. The tube bank might be in-line, staggered, or randomly distributed configurations. In the counter flow HFMC for LDAD, the solution flowed

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Table 9 Models and results of the counter flow parallel-plate membrane dehumidifiers. Studies [60]

[61]

[62]

Model descriptions (1) A lumped parameter model in a unit cell containing an air channel, a plate membrane, a solution channel, and a row of the cooling tubes; (2) Analytical solution. (1) A lumped parameter model in a unit cell containing half of a feed air channel, a membrane, a solution channel, and half of a cooling channel; (2) Analytical solution. (1) A lumped parameter model in a unit cell containing all the hollow fiber membranes; (2) Analytical solution.

Effectiveness correlations

e e

T ao T ai k1 k2 k3 cool ¼ T si T ai ¼ 1  k11 C 1 e  k12 C 2 e  k13 C 3 e xao xai k1 k2 k3 deh ¼ xsi xai ¼ 1  k21 C 1 e  k22 C 2 e  k23 C 3 e

ecool ¼ 1  k11 C 1 ek1  k12 C 2 ek2  k13 C 3 ek3  k14 C 4 ek4  k15 C 5 ek5 edeh ¼ 1  k21 C 1 ek1  k22 C 2 ek2  k23 C 3 ek3  k24 C 4 ek4  k25 C 5 ek5

esen ¼ T ao ¼ 1  C 1 ek1  C 2 ek2 eLat ¼ xao ¼ 1 þ K 1 C 1 ek1  K 2 C 2 ek2

inside the fibers, while the process air flows axially between the fibers in a counter flow configuration. In order to disclose the characteristics of the transport phenomena in the HFMC, three aspects of the membranes, tube sides, and shell sides would be reviewed here. 3.1.1. Heat and mass transfer inside membranes The membrane-based LDAD method was an air-solution contacting process, where the solution should be prevented from permitting through the membranes, while the moisture could be transferred. Therefore the hydrophobic membranes were commonly applied [10]. The heat and mass transports inside the membranes played a key role in the liquid dehumidification performances. The membranes were so thin (about 100 lm) compared to their widths and lengths that only the transports in the membrane thickness directions should be considered [19,41]. Further, their governing equations of the heat and mass conservations in the membranes had been presented based on the isotropic and asymmetric assumptions, which were second order types [19]. Actually, the heat and mass transports across the hydrophobic membranes could be simplified as the first order types. The Newton’s law of cooling and Fick’s law had been respectively used for describing the transports in the membranes [10] where the heat conductivities and the mass diffusion coefficients were the key parameters. The porous membranes were comprised of the membrane solids and the pores. The heat conductivities of the porous membranes could be calculated by the effective values based on the combinations of the solids and the air inside the pores [64], while those for the dense membranes could directly obtained by the heat conductivities of the solids themselves. Compared to the heat transfer through the membranes, the mass transport accounted for more percentages of the overall heat and mass transfer coefficients, which were respectively approximately 1% and 70% [64]. For the porous membranes, the dustygas model was applied to describe the mass transport [65], which was combined all relevant transport mechanisms across the membrane (Knudsen diffusion, molecular diffusion and viscous flow). Further, constant membrane mass transport features for variable temperature and pressure conditions were used. The single gas permeation experiments were used to obtain the viscous flow and Knudsen diffusion membrane parameters, while the binary gas diffusion experiments were applied to obtain the molecular diffusion membrane parameter. The effects of the pore diameters and distributions of the porous membranes on the moisture diffusion coefficients had been studied based on the fractal model [66]. It could be found that the pore size distribution had a determinative effect on diffusion mechanism. When the membrane maxi-

Valid conditions (1) IMLDD with cooling tubes inside; (2) Counter flow; (3) Tao  Tai < 5 °C, Xao  Xai < 0.03. (1) Adjacently IMLDD; (2) Counter flow (3) Tao  Tai < 5 °C, Xao  Xai < 0.03. (1) HFMC; (2) Counter flow (3) Tao  Tai < 5 °C, Xao  Xai < 0.03.

mum pore size exceeded 15 lm, the fractal model provided a more accurate alternative method for gas permeation analysis in porous membranes. For the dense membranes, the solution-diffusion model was commonly used for describing the mass transport inside the membranes [67]. Zhang and Huang [64] studied the heat and mass transports in various parts of the composite membranes, which were comprised of two layers: a dense PVAL skin layer and a porous polymer PVDF layer. The mechanism for moisture diffusion in the porous layer was a combination of the Knudsen and ordinary diffusion, while that in the skin layer was the solution-diffusion. The resistance model was used to obtain the effective heat conductivity and mass diffusion coefficient of the composite membranes. 3.1.2. Heat and mass transfer inside the tubes As-mentioned, in the HFMC, the solution and the air streams flowed in the shell side and the tube side, respectively. The hollow fibers were circular where the fluid flow and heat transfer were well studied. The well-established correlations had been given [68], while they might not be suitable for the hollow fiber tubes for LDAD because of their conjugate transport features. Zhang et al. [69] used a free surface model to investigate the conjugate transports in the counter flow HFMC for LDAD. The tube surface boundary conditions were non-constant value and non-constant flux conditions. The developing lengths of the heat and mass transfer for the solution stream in the circular tube channel were about 2.4 cm and 25.5 m, respectively. The fully developed Nusselt and Sherwood numbers inside the tubes under the conjugate transport boundary conditions were 4.38 and 4.46, respectively, which was nearly equal to NuH (4.36), but larger than NuT (3.66) [68]. For the free surface model, only one tube and the surrounding fluid were selected as the calculated domain. Then Huang et al. [70] investigated the influences of the fiber-to-fiber interactions on the conjugate transports in the HFMC based on the unit cells containing two fibers (in-line) or three fibers (staggered). The transports inside the tubes were the same as those in obtained by the free surface model [69]. Further, the transport phenomena were almost independent on the various tubes and tube distributions. 3.1.3. Heat and mass transfer outside the tubes 3.1.3.1. Regularly distributed types. The air stream commonly flowed in the shell side between the tubes, which were more complex than the tube side. As-mentioned, the tubes might be regularly distributed (in-line or staggered) and randomly distributed. Zhang et al. [69] studied the conjugate transports in the shell side with a regular configuration. The NuC,a was between NuH and NuT. NuC, a was closer to NuH. The ShC,a were somewhat less than the NuC,a. Both NuC,a and ShC,a decreased with an increase in the packing frac-

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and semi axis ratios (b/a) were listed in Table 11. The NuC,a was close to NuH. The ShC,a was somewhat smaller than the NuC,a. Compared to the HFMTB, NuC,a and ShC,a were deteriorated about 6–18% and 5–15% respectively. The results were obtained based on the free surface model, not considering the fiber-to-fiber effects. Huang and Yang [76] investigated the axial fluid flow and heat transfer between the EHFMTB with a constant wall temperature boundary condition in the regularly distributed unit cells containing two fibers (in-line) or three fibers (staggered) considering the fiber-to-fiber effects. The (fRe) and NuT with various elliptical semi axis ratios (b/a) and pitch-to-diameter ratios (SL/d) were listed in Table 12. Huang et al. [77] also studied the effects of the various longitudinal and transverse pitch-to-diameter ratios (SL/d, ST/d) and semi axis ratios (ble/ale, bri/ari) on the axial transports between a regularly distributed (in-line or staggered) EHFMTB. The (fRe) and NuT with various elliptical semiaxis ratios (b/a) of the left and the right tubes were listed in Table 13. The fRe and NuT had their highest values when the ble/ale and bri/ari both were equal to 1.0 and 1.5 for the in-line and the staggered configurations, respectively. The SL/d and ST/d had larger effects than the ble/ale and bri/ari on the fRe and NuT. The EHFMTB could also be the randomly distributed elliptical hollow fiber membrane tube banks (REHFMB). The longitudinal transport phenomena between the REHFMB were studied [78] based on an element including 20 fibers by the Voronoi tessellation method. The randomly elliptical fibers were divided into a series of polygonal cells. The random distributions in positions, elliptical semi axis ratios, and rotation angles all had influences on the axial transport phenomena.

tion. The free surface model method was only a approximate treatment. The interactions between the neighboring fibers were not taken into account seriously. In order to accurately reflect the transport phenomena in the HFMTB, Huang et al. [70] studied the fiber-to-fiber effects on the conjugate transport phenomena. In the shell side, the fully developed (fRe)a, NuC,a and ShC,a under the conjugate boundary condition with various packing fractions (u) and pitch-to-diameter ratios (SL/(2ro)) were listed in Table 10. The air side NuC,a and ShC,a decreased with an increasing in the u. For the in-line and the staggered configurations, the NuC,a and ShC,a were somewhat larger than NuH for a sparse tube bank. However the NuC,a and ShC,a were less than NuH for a dense one. 3.1.3.2. Randomly distributed types. The tubes might also be randomly distributed, which were more complex than the regular ones. Zheng et al. [71] studied the effects of the random distributions of the HFMTB on the shell side mass transfer performances. The fiber random configuration caused a slight decrease (9–16%) on the shell side mass transfer coefficient. The ratio between the mass transfer coefficient in random packing and the coefficient in uniform one was in the range of 0.91–0.84, which decreased slightly with the packing fraction increasing. Li and Zhang [72] found that the flow maldistribution had greater impacts on membrane contactors. In air humidification, the sensible cooling effectiveness could be deteriorated by 17–36% under various air flow rates, while the humidification effectiveness could be deteriorated by 21–39%. It was obvious that the transport deteriorations might be caused by the irregularity of fiber spacing, the polydispersity of the fiber diameters, and the fiber movement during operation, the effect of the contactor walls, and the inlet and outlet effects [73]. Therefore the regular HFMC should be made to improve the performances. Huang and He [74] firstly invented an automatic arranging machine for hollow fibers and a membrane contactor production device, which relied on a membrane fiber automatic filling machine to accurately control the fibers, and realized passing the automatic membrane fiber through the membrane fiber bracket into a tube by using the fiber traction apparatus to cooperate with a related mechanical walking structure. The automatic control technology utilized in the whole process was mature. The manual intervention process was less. The consumables were used less. The manufactured components were with high quality. The production efficiency was high. The produced membrane contactor had a good performance consistency.

3.2. Cross-flow hollow fiber membrane dehumidifiers The HFMC for gas-liquid contact were usually designed in the counter flow, co-current flow, or counter flow patterns, respectively. The primary limitations of these so-called counter flow contactors were the shell-side flow channeling or mal-distribution because of the non-uniformly fiber population [79]. Therefore a cross-flow HFMC was proposed depicted in Fig. 8(a). The employment of the cross-flow HFMC could ensure a significant reduction of temperature polarization effect at a relatively low range of Reynolds numbers [80]. The structure of the cross-flow HFMC was usually in in-line or staggered configurations. Alrehili et al. [81] found that the cross-flow HFMC with the staggered configurations performed much better than those with the in-line configurations. Dindore and Versteeg [82] found that the mass transfer performances of the cross-flow HFMC involving incompressible fluids could be directly predicted by the heat transfer analogy when the mass transfer was not aided by the chemical reactions. Song et al. [83] investigated the membrane flux curve characteristics under various experimental conditions. It was most likely to obtain a high membrane flux under the conditions of a high feed inlet temperature, a high feed flow rate, a low distillate inlet temperature, and a high distillate flow rate. Zuo et al. [84] conducted

3.1.3.3. Elliptical hollow fiber membrane types. However in real applications, circular hollow fibers in the HFMTB might be deformed to be the elliptical ones because of their weak mechanical strength. In order to disclose the influences of the shape changes of the fibers, Huang et al. [75] investigated the conjugate transports in the counter flow HFMC with an elliptical hollow fiber membrane tube bank (EHFMTB), as depicted in Fig. 7 used for LDAD. For the air stream, the (fRe)a and NuC,a under different u

Table 10 Fully developed (fRe)a, Nusselt and Sherwood numbers under the conjugate heat and mass transfer boundary condition for the axially air flowing between the HFMTB, ST = SL [70]. SL/(2ro)

3.0 2.75 2.5 2.25 2.0 1.75 1.5 1.25

Packing fraction (u)

0.087 0.104 0.126 0.155 0.196 0.256 0.349 0.502

In-line configuration

Staggered configuration

NuH

NuT

NuC,a

ShC,a

(fRe)a

NuH

NuT

NuC,a

ShC,a

(fRe)a

25.38 22.93 20.57 18.27 15.92 13.83 11.50 8.42

24.61 21.99 19.50 17.08 14.63 12.01 8.80 4.69

25.87 23.21 20.56 18.11 15.45 12.76 9.68 5.66

24.72 22.88 19.73 17.67 14.86 12.32 9.42 5.43

251.90 228.37 205.82 184.15 163.18 142.41 120.32 90.63

25.49 23.01 20.66 18.44 15.95 14.52 13.10 11.14

24.48 21.97 19.56 17.22 14.95 12.69 10.32 7.55

25.71 23.14 20.61 18.07 15.55 13.43 11.51 8.65

24.81 23.02 19.83 17.54 14.86 12.05 9.46 6.65

247.91 225.48 203.77 183.11 163.45 144.58 126.07 105.69

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Air in

Solution in

Solution out

Elliptical hollow fiber membrane tubes

Air out

(a)

Membrane δ

ao bi

rf

bo ai

Free surface

(b) Fig. 7. Schematic of the HFMC [75]. (a) The shell and the tube structure; (b) The unit cell with a free surface.

Table 11 Fully developed (fRe)a, Nusselt and Sherwood numbers for the axially air flowing between the EHFMTB with a free surface unit cell [75]. Packing fraction (u)

0.063 0.162 0.203 0.253 0.281 0.360 0.422 0.461

bo/ao = 1.0 (bi/ai = 1.0)

bo/ao = 0.9 (bi/ai = 0.88)

NuH

NuC,a

ShC,a

(fRe)a

NuT

NuH

NuC,a

ShC,a

(fRe)a

NuT

NuH

NuC,a

ShC,a

(fRe)a

7.05 5.51 5.24 4.99 4.89 4.66 4.52 4.42

7.49 5.97 5.67 5.43 5.32 5.10 4.95 4.82

7.44 5.94 5.64 5.40 5.30 5.07 4.92 4.80

7.30 5.84 5.57 5.34 5.22 4.98 4.86 4.75

13.81 16.72 17.53 18.39 18.81 19.85 20.55 21.25

6.97 5.43 5.16 4.91 4.81 4.58 4.44 4.33

7.41 5.89 5.59 5.35 5.24 5.02 4.87 4.73

7.36 5.86 5.56 5.32 5.22 4.99 4.84 4.72

7.22 5.76 5.49 5.26 5.14 4.89 4.78 4.67

13.81 16.72 17.53 18.39 18.81 19.85 20.55 21.25

6.85 5.31 5.04 4.79 4.69 4.46 4.32 4.21

7.29 5.77 5.47 5.23 5.12 4.9 4.75 4.61

7.24 5.74 5.44 519 5.11 4.87 4.72 4.58

7.10 5.64 5.37 5.14 5.02 4.78 4.66 4.55

13.80 16.71 17.52 18.38 18.80 19.84 20.54 21.24

bo/ao = 0.7 (bi/ai = 0.64)

0.063 0.162 0.203 0.253 0.281 0.360 0.422 0.461

bo/ao = 0.8 (bi/ai = 0.76)

NuT

bo/ao = 0.6 (bi/ai = 0.52)

bo/ao = 0.5 (bi/ai = 0.41)

NuT

NuH

NuC,a

ShC,a

(fRe)a

NuT

NuH

NuC,a

ShC,a

(fRe)a

NuT

NuH

NuC,a

ShC,a

(fRe)a

6.73 5.19 4.92 4.67 4.57 4.34 4.20 4.11

7.17 5.65 5.35 5.11 5.01 4.78 4.63 4.49

7.12 5.62 5.32 5.08 4.98 4.75 4.61 4.48

6.98 5.52 5.25 5.02 4.89 4.66 4.54 4.43

13.78 16.69 17.49 18.36 18.78 19.82 20.52 21.22

6.62 5.08 4.81 4.56 4.46 4.23 4.09 3.98

7.06 5.54 5.24 5.01 4.89 4.67 4.52 4.38

7.01 5.51 5.21 4.97 4.87 4.64 4.49 4.37

6.87 5.41 5.14 4.91 4.79 4.55 4.43 4.32

13.74 16.65 17.46 18.32 18.74 19.78 20.48 21.18

6.33 4.79 4.52 4.27 4.17 3.94 3.81 3.69

6.77 5.25 4.95 4.71 4.61 4.38 4.23 4.09

6.72 5.22 4.89 4.68 4.58 4.35 4.2 4.08

6.58 5.12 4.67 4.62 4.5 4.26 4.14 4.03

13.61 16.52 17.33 18.19 18.61 19.65 20.35 21.05

detailed studies of the cross-flow HFMC with using Aspen Plus to know the effects of the operation parameters (membrane area, feed temperature, feed and permeate velocities) on the water flux/production. Based on the previous works, Zhang [85] stepped forward to study the coupled transport phenomena in an application-scale cross-flow HFMC for air humidification. The

effects of overall transport properties on the humidification effectiveness were analyzed. It could be found that the heat transfer resistance in membranes could be neglected. However the mass transfer resistance in membranes was substantial. The packing density was a determining factor. The results were obtained by the pressure drop and convective heat mass transfer correlations

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S.-M. Huang et al. / International Journal of Heat and Mass Transfer 139 (2019) 881–906 Table 12 Fully developed (fRe) and Nusselt number under uniform temperature boundary condition (NuT) for the fluid flowing axially between the EHFMTB, ST = SL [76]. SL/d

b/a

SL/(2b)

ST/(2a)

In-line

Staggered

SL/d

b/a

SL/(2b)

ST/(2a)

(fRe)

NuT

(fRe)

NuT

1.5

0.5 0.6 0.7 0.9 1.0 1.2 1.5 1.8 2.0

2.31 2.03 1.84 1.58 1.0 1.38 1.26 1.19 1.16

1.16 1.22 1.28 1.43 1.0 1.65 1.89 2.14 2.31

101.10 106.55 111.38 117.09 117.75 115.94 110.03 104.34 101.28

7.38 7.58 7.73 7.85 7.86 7.82 7.67 7.49 7.38

102.02 108.24 114.29 123.82 126.88 130.03 130.54 129.56 128.90

2.5

0.5 0.6 0.7 0.9 1.0 1.2 1.5 1.8 2.0

3.85 3.39 3.06 2.64 2.5 2.30 2.10 1.98 1.93

1.93 2.03 2.14 2.38 2.5 2.76 3.16 3.57 3.85

197.24 198.52 199.40 200.28 200.40 200.23 199.43 198.41 197.51

18.09 18.18 18.39 18.44 18.53 18.42 18.36 18.19 18.09

198.51 200.19 201.54 203.41 204.03 204.82 205.25 205.22 205.05

In-line

Staggered

(fRe)

NuT

(fRe)

NuT

7.56 7.97 8.47 9.66 10.25 11.01 11.12 10.66 10.36

2.0

0.5 0.6 0.7 0.9 1.0 1.2 1.5 1.8 2.0

3.08 2.71 2.45 2.11 2.0 1.84 1.68 1.59 1.54

1.54 1.63 1.71 1.90 2.0 2.21 2.53 2.86 3.08

153.81 155.94 157.56 159.22 159.43 159.01 157.35 155.42 154.22

12.89 13.12 13.34 13.45 13.57 13.39 13.27 13.05 12.89

154.37 157.53 159.81 163.12 164.22 165.62 166.47 166.61 166.53

13.10 13.50 13.89 14.56 14.83 15.18 15.39 15.40 15.36

18.38 18.66 18.91 19.23 19.43 19.62 19.75 19.77 19.76

3.0

0.5 0.6 0.7 0.9 1.0 1.2 1.5 1.8 2.0

4.63 4.06 3.67 3.17 3.0 2.76 2.53 2.38 2.31

2.31 2.44 2.60 2.85 3.0 3.31 3.79 4.29 4.63

241.16 242.19 243.23 243.68 244.14 244.01 243.31 242.38 241.74

23.23 23.36 23.49 23.55 23.60 23.54 23.48 23.36 23.23

242.70 244.26 245.44 246.91 247.34 247.78 247.81 247.42 247.04

23.59 23.83 24.01 24.28 24.37 24.48 24.54 24.53 24.49

Table 13 Fully developed (fRe) and Nusselt numbers under the uniform temperature boundary condition (NuT) for the fluid flowing axially between the EHFMTB, SL/d = ST/d = 2.0 [77]. ble/ale

bri/ari

In-line

Staggered

ble/ale

fRe

NuT

fRe

NuT

bri/ari

In-line

Staggered

fRe

NuT

fRe

NuT

0.5

0.5 0.6 0.8 1.0 1.2 1.5 2.0

153.80 155.08 156.85 157.49 157.97 157.84 156.83

12.90 13.03 13.25 13.38 13.43 13.38 13.12

156.80 158.04 160.32 161.43 162.34 162.66 162.15

13.11 13.33 13.69 13.94 14.10 14.13 13.90

1.0

0.5 0.6 0.8 1.0 1.2 1.5 2.0

157.51 158.09 159.11 159.43 159.09 158.39 156.82

13.38 13.44 13.54 13.57 13.53 13.39 13.06

161.35 162.49 164.63 165.66 166.51 166.82 166.29

13.92 14.13 14.54 14.83 15.02 15.05 14.72

1.5

0.5 0.6 0.8 1.0 1.2 1.5 2.0

157.84 158.13 158.67 158.39 158.11 157.33 155.58

13.37 13.38 13.39 13.39 13.36 13.26 13.02

162.57 163.62 165.70 166.79 167.73 168.26 168.09

14.08 14.26 14.65 15.01 15.27 15.45 15.32

2.0

0.5 0.6 0.8 1.0 1.2 1.5 2.0

156.83 156.96 157.22 156.82 156.46 155.58 154.22

13.12 13.08 13.07 13.06 13.06 13.02 12.88

162.07 163.08 165.10 166.29 167.35 168.11 168.29

13.82 13.95 14.28 14.65 14.97 15.30 15.43

to calculate the friction factors, Nusselt and Sherwood numbers across the tube banks with various parameters proposed by Zhang and Li [86]. Compared to the previous researches and the available correlations, their results were more appropriate for the cross-flow HFMTB operated under the relatively low Reynolds numbers from 100 to 500 with turbulent and the least transitional flow behaviors. The cross-flow HFMC was composed of not only the hollow fibers (the core), but also the inlet/outlet headers, which may lead to unevenly distributed flow at the inlet of the core. Zhang et al. [87] investigated the influences of structure-induced flow maldistributions affected by the inlet header, the core, and the outlet header on the deteriorations in performances. The packing fraction affected the flow maldistribution substantially. Then the cooling and humidification effectiveness were influenced significantly, which could be deteriorated by about 3–30% and 26–58%, respectively. The HFMC was formed by the HFMTB with a number of hollow fibers (100–5000) packed in the shell and the air flows over the tube bank. The direct modeling of the whole tube bank was so difficult that three approaches of the free surface model, periodic unit cell model, and multiple tube model were applied to describe the fiber-to-fiber interactions.

3.2.1. Free surface model The free surface model was proposed to solve this problem. A unit cell including a single fiber, the solution stream inside the fiber, and the air stream outside the fiber was considered. The air stream outside the fiber had an outer free surface. The crosssections of the conjugated tube channel and air channel (separated and coupled by the membrane) were a circle and an annulus respectively, as shown in Fig. 8(b). Zhang et al. [88] studied the conjugate transports in the cross-flow HFMC for LDAD with assuming that both the air and the solution fluids being laminar flow because of the Re less than 300. The air side Nusselt and Sherwood numbers were larger than those in the counter flow configuration when the Re were larger than 35. Huang et al. [89] applied the Happel’s free surface model and the low-Re k-e turbulent model to capture the air flow with the Re ranging between 300 and 600, while a laminar model was applied for the solution flow in the tubes. Both the Nusselt and Sherwood numbers increased with an increase in the packing fractions in the shell side. As-mentioned, the conjugate transports could be deteriorated in the counter HFMC. However the elliptical cross-sectional hollow fibers were better without any enhanced components in the crossflow ones. Huang and Yang [90] studied the conjugate heat and

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Fig. 8. Schematic of a cross flow hollow fiber membrane contactor, a free surface cell selected as the calculation domain [79]. (a) The shell and the tube structure; (b) The free surface cell [88].

mass transfer in the elliptical hollow fiber membrane contactor (EHFMC) employed for LDAD. The air side Nusselt and Sherwood numbers were increased with the elliptical semiaxis ratios (b/a) decreasing. The comprehensive performances both in the air and the solution channels were improved. 3.2.2. Periodic element model Obviously, the effects of the fiber-to-fiber interactions were not taken into account seriously for the free surface model. Therefore the results were not accurate enough. The calculating domains of the periodic areas surrounded by the neighboring fibers were taken to account for the interactions. Huang et al. [91] solved the transport problems by considering the interactions between the neighboring fibers with assumption that air and the solution streams were laminar flows with the Re less than 300. The friction factors and the Nusselt numbers estimated by the free surface model could deviate as larger as 20% from this approach for an in-line configuration and/or a staggered configuration. Zhang et al. [92] stepped forward to investigate the same problem with considering turbulent flow in air side with the Re ranging from 300 to 600. Compared to the available experimental data, the low-Re k-e turbulent model was better than a laminar model when the air side Re were higher than 300. For the EHFMTB for air humidification, Huang et al. [93] studied the heat transfer and fluid flow across the tube bank under the transitional flow zone with the air Re between 300 and 600. Compared to the HFMTB, the comprehensive heat transfer of the air flow across the EHFMTB along the minor semiaxis direction was

enhanced. Further, the heat transfer enhancement was more obvious with the semiaxis ratios (b/a) decreasing. The tube bank was always randomly populated in real applications because of the convenience in the manufacturing process. Jiang et al. [94] attempted to disclose the influences of the fiber configurations on the performances. The EHFMTB with random distribution had about 30.5–89.8% lower fm and about 5.5–61.1% lower Num than those with regular configuration. 3.3. Inclined flow hollow fiber membrane dehumidifiers The researches mentioned above had been focused on either counter flow or cross-flow. However they were rather ideal flow conditions. Further, the pure counter flow configuration was very difficult in channel sealing. The performances of the pure counter flow HFMC were commonly better than those for the pure crossflow configuration. In practical applications, the fluid inlet was small but the fibers were long in the HFMC. Therefore the fluid could usually impinge the tube bank with an inclined angle. The inclined flow might be a more common type. Ali and Vafai [95] analyzed the heat mass transfer between a falling film and air along a single inclined flow fiber. The inclined angle could largely enhance the dehumidification performance compared to counter flow. Igarashi and Mayumi [96] studied the fluid flow and heat transfer features around a cylinder at small attack angles and gave the local heat transfer coefficients. Choietal [97] tested the flow field around a cylinder with the inclined flow and observed the effect of inclination on the vortices and wakes. These researches

S.-M. Huang et al. / International Journal of Heat and Mass Transfer 139 (2019) 881–906

were focused on the oblique flow around a single fiber without considering the interactions between the neighboring fibers. Ouyang et al. [98] studied the laminar inclined fluid flow and heat transfer across the HFMTB with the regular distribution under the uniform surface temperature boundary condition, as plotted in Fig. 9. The influences of the impinging angles from 0° to 90° on the laminar flow and heat transfer were discussed. The friction factors and Nusselt numbers increased with the inclined angles increasing. At the same inclined angle, the staggered arrangement had larger pressure drops and heat transfer coefficients than those for the in-line configuration. Based on the former work, Zhang et al. [99] stepped forward to investigate the inclined flow over a circular tube bank under the constant wall heat flux boundary condition. The convective heat transfer to pressure drop ratio, j/f factor decreased with the inclined angles increasing under the uniform heat flux boundary condition, meaning that the comprehensive performances for the counter flow configuration were relatively better than that for the cross-flow one. The M-LDAD process was a conjugate problem. Ouyang and Zhang [100] studied the conjugate transports in a skewed-flow HFMTB used for LDAD. The heat and mass transfer rates and friction factors for the inclined flow configuration were higher than those of the parallel flow one but lower than those of the cross-flow one.

4. Membrane-based liquid desiccant air dehumidification systems Various membrane-based dehumidifiers had been introduced in above-mentioned researches. The continuous LDAD process should connect with other devices involving the dehumidifier, regenera-

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tor, container, pump, heater, cooler, etc. to form the membranebased liquid desiccant air dehumidification (M-LDAD) systems. Depending on the dehumidifier numbers, the systems could be classified into the single-stage and multistage ones. Also the systems could be divided into the operational ones driven by the pumps and the natural ones driven by the solution convection.

4.1. Single-stage systems 4.1.1. Operational systems driven by pumps 4.1.1.1. The M-LDAD systems connected with other air-conditioning systems. These M-LDAD systems were driven by the air and the solution pumps to form the air flow and the solution flow loops, respectively. In the practical applications, the M-LDAD systems were usually connected with other air-conditioning systems. Abdel-Salam and Simonson [101] proposed a membrane-based liquid desiccant air conditioning (M-LDAC) system. The schematic diagrams of the M-LDAC and the M-LDAC-ERV systems were presented in Fig. 10. The life cycle cost and the annual primary energy consumption of the M-LDAC system were respectively 12% and 19% lower than those of the conventional air conditioning (CAC) system. Based on the former research, a heat pump was used to cover the solution cooling and heating loads simultaneously instead of using separate heating and cooling equipments. The performances of the heat pump evaporator and condenser were matched to meet the solution cooling and heating demands [102]. Abdel-Salam et al. [103] used a solar thermal system (STS) to regenerate the diluted solution, which could largely improve the economic and environmental performances. Xiao et al. [104] proposed a novel dedicated outdoor air system (DOAS) consisting

Fig. 9. Schematic of oblique flows across the HFMTB [98–100]. (a) In-line layout; (b) Staggered layout.

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Fig. 10. Schematic diagrams of the LDAC systems [101]. (a) M-LDAC systems; (b) M-LDAC-ERV systems.

of a liquid dehumidifier, a membrane-based total heat exchanger, a dry cooling coil and a regenerator, which could control the temperature and humidity independently for the air flow. The application of the total heat exchanger could enhance the system energy performance by 19.9–34.8%. Ge et al. [105] presented a model-based dedicated outdoor air-chilled ceiling (DOAS-CC) system to optimize the overall system performance. The optimal strategy could save about 4.7% and 10.5% energy in summer and spring days compared with the conventional control strategy. Safizadeh et al. [106] proposed a novel air-conditioning system containing a membrane unit, a novel evaporative cooled sportive dehumidification system (ECOS), and an electricity-efficient cooling technology. An optimization approach of the genetic algorithm was developed to obtain an optimized balance of three air-conditioning components. An integration of a small ECOS, a relatively large membrane unit, and a chiller operating at an elevated evaporation temperature was the most cost effective combination meeting comfort criteria. Lin et al. [107] conducted a thermodynamic analysis on a hybrid system combing M-LDAD and M-cycle dew point evaporative cooling technology, which could generate the processing air at 10.9 g/ kg humidity ratio and 18.3 °C temperature under summer conditions. Chen et al. [108] developed a membrane-based hybrid liquid desiccant dehumidification cooling system to realize efficient temperature and humidity controls in hot and humid zones. As the inlet air relative humidity increased from 46% to 70% at the constant temperature of 34.6 °C, the moisture removal rate was doubled and the dehumidification effectiveness was improved by 36.9%. With the solution concentration ranging from 30% to 42%, the dehumidification performance was improved from 0.05 g/s to 0.14 g/s and the cooling output was doubled. Chen et al. [109] stepped forward to study a membrane-based liquid desiccant dehumidification cooling system for energy efficient airconditioning with independent temperature and humidity controls. The higher the solution concentration was, the better the dehumidification performance was. Bai et al. [110] experimentally studied the performances of a cross-flow M-LDAD system. The influences of the main operating parameters including the NTU,

m*, Ts,in, and Cs on the sensible, latent, and total effectiveness had been assessed. The sensible, latent, and total effectiveness reached the maximum values of 0.49, 0.55, and 0.53, respectively at m* = 3.5 and NTU = 12. Bai et al. [111] stepped forward to study the operating parameters of the same M-LDAD system. The m* and NTU were two key parameters affecting the system effectiveness. 4.1.1.2. The liquid-to-air energy recovery systems. For the liquid-air energy recovery processes with the same transport phenomena as those for the LDAD, Mahmud et al. [112] designed a novel run-around membrane energy exchanger (RAMEE) system. Each liquid-to-air membrane energy exchanger (LAMEE) connected by a solution flow loop in the RAMEE system was shown in Fig. 11. In the system, an exchanger was located in the supply air stream entering the building, and the other was located in the exhaust air stream leaving the building. Under the summer conditions, the total effectiveness increased with the solution flow rate increasing, while decreased as the air flow rate increasing. Under the winter conditions, the total effectiveness changed a little with the variations in the air and solution flow rates. Hemingson et al.

Fig. 11. Schematic of a run-around membrane energy exchanger (RAMEE) system in a HVAC application [112].

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[113] studied the steady-state performance of the RAMEE system for a wide range of the outdoor air conditions. The effectiveness were rather dependent on the outdoor conditions, which resulted in some values exceeding 100% or being less than 0 for some specific outdoor air conditions. Abdel-Salam et al. [114] then proposed a M-LDAC system based on the counter flow LAMEE, not the RAMEE. However the counter flow LAMEE was difficult in channel sealing between the air and the solution streams. Further, there was not experimental works in this research. Only modeling works were conducted. The system COP was 0.68. The sensible heat ratio (SHR) for the LAMEE was between 0.3 and 0.5. It was obvious that the RAMEE system was the energy recovery system, not only for the heat recovery or only for the mass recovery. For the only sensible heat recovery system, Vali et al. [115] developed a steadystate model to study only the heat transfer problem. The effectiveness correlations for the heat exchangers and the whole heat recovery system were developed. For a given area of the exchangers, the largest sensible effectiveness was reached with a small aspect ratio and short inlet and outlet lengths of the side in and side out solution channel.

4.1.1.3. The M-LDAD systems based on the internally-cooled membrane contactors. Based on the IMLDD to enhance the solution absorption performance, Woods and Kozubal [116] proposed a LDAC system consisting of a parallel-plate IMLDD and an indirect

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evaporative cooler. The desiccant-enhanced evaporative (DEVAP) system was presented in Fig. 12. In the first stage, the solution films formed in the solution channels removed the water vapor from the air stream through the membranes. The water films wetted the surfaces of the cooling channels, which took away the enthalpy from the liquid desiccant into an exhaust air stream. Therefore the heated solution was cooled by the cooling channels. The second stage was the counter flow indirect evaporative cooler. Dong et al. [117] then used district heat from a combined heat and power (CHP) system in the DEVAP system as the heat source, which produced 40.5% less CO2 and consumed 46.2% less primary energy compared with the system using the conventional gas boiler. Kim et al. [118] comparatively evaluated the energy performances of a dedicated outdoor air system (DOAS) and the DEVAP system. The energy saving potential of the DOAS was larger than that of the DEVAP system. Further, the DOAS with the ceiling radiant cooling panels was 20% less primary energy consumption compared to that of the DEVAP system. However the waste energy sources (industry waste heat) and the renewable energy sources (geothermal energy, solar thermal energy, etc) could be applied in the DEVAP system to regenerate the diluted solution. Therefore the LDAD technology was still a promising technology.

4.1.1.4. Solution regeneration problems in the LDAD systems. For the comprehensive LDAD system, the solution regenerator was also an

Fig. 12. Schematic of the desiccant-enhanced evaporative (DEVAP) system [116]. (a) Four stacked channel pairs (b) Top view, one channel pair.

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important component, which largely affected the total performance of the whole system. Keniar et al. [119] investigated the feasibility and performance of an M-LDAD system combined with solar energy. The solar energy was used as the source of the thermal energy for the desiccant regeneration and seawater was used to cool the desiccant, which was feasible. The payback period to return the initial cost of the equipment was 7 years and 8 months. Zhang et al. [120,121] presented a compression heat pump driven M-LDAD system shown in Fig. 13, which consisted of a solution regeneration module, an air dehumidification module and a compression heat pump system. Both the dehumidification and the regeneration modules were the cross-flow HFMC. The exhaust air was used as the regeneration air source. The diluted salt solution was heated and regenerated by the heating energy of the condenser, while the concentrated solution was cooled by the cooling energy of the evaporator. They are both the output energy of the heat pump system. The energy efficiency ratio (EER) was defined as the ratio between the heating energy of the condenser and the cooling energy of the evaporator to the compression power. The EER of the heat pump was all greater than 3.75, which was obtained under some specific operating conditions. However the seasonal energy efficiency ratio (SEER), which was usually used to present the average performance of the heat pump, was still not obtained and analyzed. The solution flow rate was increased to improve the dehumidification performances of the whole system. Then the variation of the solution temperature was relatively small. The heating and cooling loads were small. Therefore the EER was not large. The dehumidification COP of the system was varied from 0.4 to 0.9, indicating that the energy performance of the system performed well. In addition, the transient behaviors of the system were analyzed both at the start-up and in the normal operation periods. The initial concentration of the solution and the volume of the stored solution played key roles at the start-up period, while the initial temperature had less influences. Controlling the compressor speed might be the feasible method to match the load and weather fluctuations. Su and Zhang [122] firstly proposed a LDAD system of a frost-free air source heat pump (ASHP) combined with M-LDAD and humidification. The membrane dehumidifier was used to dehumidify the air before entering the outdoor coils to prevent the frosting. Rattner et al. [123] modeled a thermally driven parallel-plate air-gap membrane distillation desiccant regenerator for LiCl in the dehumidification applications. The membrane materials had little effect on the regeneration performance. Duong et al. [124] used membrane distillation (MD) for regeneration of LiCl solution in the LDAC system. The MD process at the feed temperature of 65 °C could increase the LiCl concentration up to 29 wt% without any observable LiCl loss. It was obvious

Fig. 13. Schematic of a heat pump driven membrane-based liquid desiccant air dehumidification (M-LDAD) system [120,121].

that the MD for the solution regeneration driven by the waste energy sources (industry waste heat) and the renewable energy sources (geothermal energy, solar thermal energy, etc) might be a promising approach. 4.1.2. Natural systems driven by solution convection The above-mentioned M-LDAD systems were driven by the air and the solution pumps using high grade energy (electricity) to form the two flow loops. In order to save electricity, Fazilati et al. [125,126] firstly proposed and approved the application of the M-LDAD system driven by the natural convection of the solution in the practical operating conditions, as shown in Fig. 14, which consisted of two counter flow HFMC as the absorber and the regenerator connected to form a closed loop. There was not the solution pump in this system, where no parasitic pump energy was required. In each membrane contactor, the hollow fibers separated the air and solution streams. The cooling and the heating water streams flowed in the outer jackets of contactors. When the absorber and the regenerator were placed in the same level, the first type system, which had relatively high time constant, was formed [125]. This phenomenon should be eliminated in the operation. Therefore the regenerator was placed above the absorber to enhance the natural convection influence by the gravity to form the second type system, where the loop was redesigned to resolve the shortcomings of first type system [126]. The performance of the system was experimentally investigated under two inlet air conditions which resembled two conditions with relative humidity of 78% and 37% and temperature of 31.5 °C. The results showed that the loop could act as the natural convection LDAC system. Its COP values were 3.95 and 4.56 for higher and lower humidity ratio air inlets, respectively. It was obvious that this system performed better in more humid inlet air condition. In this system, the major part of the electric power was used for driving the cooling water stream. If it was replaced with free cold water source, the COP of the system would increase in a relatively large extent. It was obvious that the performance of the natural convection M-LDAD system omitting the solution pump might be better than those driven by the heat pumps. However the response times of the start-up and stop were long. Further, the solution flow rate might be difficult to be adjusted, which made the whole system difficult to be controlled. However the natural convection system could be used in the specific condition, which was safe enough and undemanding. 4.2. Multistage system As mentioned, the M-LDAD systems could be classified as the single-stage and the multistage ones depending on the numbers of the membrane-based dehumidifiers. The latter ones could alleviate the temperature rise in a certain degree by the intercooling of the solution during the LDAD process and realize larger air flow rate. Zhang et al. [127] firstly proposed and studied a novel compression heat pump driven and two-stage M-LDAD system, as presented in Fig. 15, which contained two dehumidifiers (HFMC), two solution regenerators (HFMC), and a compression heat pump system. The heated solution before into the dehumidifiers was cooled by the evaporators to form the internally-cooled type, while the solution before into the regenerators was heated by the condensers to enhance the regeneration ability. The concept was similar to an internally-cooled HFMC. Compared with the single-stage M-LDAD system, the two-stage system had a lower solution concentration exiting from the dehumidifier and a lower condensing temperature. Further, the COP could increase about 20%, which was about 4.5–5.5. Zhang et al. [128] stepped forward to present a novel capacity matching index (CMI), which considering the two regenerators, two evaporators, two condensers, and two dehu-

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Fig. 14. The M-LDAD system driven by the natural convection of the solution [125,126].

Fig. 15. Schematic of a heat pump driven two-stage M-LDAD system [127–129].

midifiers, to evaluate the energy capacity matching of the same system. The CMI was usually smaller than 1 under the typical hot and humid weather condition and the demand, while the energy supply in the system was mismatched. The dehumidification rates, CMI, EER and COP decreased with an increase in the inlet air temperature, while the CMI was constant under various inlet air humidity ratios. Based on the former researches, for the same system, Zhang et al. [129] used the genetic algorithm to minimize the energy consumption to solve the optimization problem based on a set of the transport models, which were validated by an experimental work. The hourly optimal regulation could be hourly regulated under hot and humid weather conditions. The energy

consumption was reduced by more than 20% to satisfy the indoor air humidity demand. This optimization strategy was suitable for controlling and monitoring the systems under hot and humid weather conditions. Based on the concept of the two-stage MLDAD system, Zhang [130] proposed a multiple-stage M-LDAD system, which contained three HFMC as the dehumidifiers and also three HMFC as the regenerators. The other components were the same as those for the two-stage one. The three-stage system was suitable for the relatively large air flow rate. As above-mentioned, the natural convection M-LDAD systems have the relatively large potential to reduce energy consumption. Huang et al. [131] firstly proposed a zero-energy driven M-LDAD

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Fig. 16(a) and (b), respectively. N (larger than or equal to 2) dehumidifiers and regenerators, which were the HFMC, were applied in the system. If the air flow rates were large, more multistage system should be used to reduce the air flow pressure drop. The cooling

system without any solution pump, as presented in Fig. 16. The solution stream flows in its loop by the natural concentration and thermal convections. Two types of the dehumidifier and the regenerator on the same and different levels were shown in

Cooling water in

Solution container

Valve Hot water in

Dehumidfier Humid air in Regenerated air in Regenerator

Cooling coils

(a)

Pre-heater

Pre-cooler

(b) Fig. 16. Schematic of a zero-energy driven M-LDAD system [131]. (a) The dehumidifier and the regenerator in the same level; (b) The dehumidifier and the regenerator in the different level.

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water stream flows inside the HFMC to form the internally-cooled dehumidifiers and pre-heated regenerators to improve the performance of the whole system. However their structures were different with those introduced by Zhang et al. [127–130]. The cooling coils were installed inside the HFMC to form a compact structure, not to separate the various stages shown in Fig. 15.

5. Conclusions and perspectives The membrane-based liquid desiccant air dehumidification (MLDAD) technology had been reviewed based on the momentum, heat, and mass transports, the performance analysis, and the thermodynamics. The parallel-plate membrane dehumidifiers (PMC) and hollow fiber membrane contactors (HFMC) with the counter flow, cross-flow, and quasi-counter flow configurations were summarized. The adiabatic and internally-cooled membrane dehumidifiers were introduced. The single-stage and multistage M-LDAD systems to improve their performances were analyzed. This review had given a comprehensive analysis of these researches. Following results could be found: (1) For the PMC, the performance of the counter flow PMC was better than that for the cross-flow one, while the pure counter flow PMC was rather difficult in channel sealing in the practical applications. Therefore the quasi-counter flow PMC of the side in and side out type and hexagonal type were the promising alternatives. These two types could be selected by the structures of the spaces inside the M-LDAD systems. Further, the internally-cooled quasi-counter flow PMC might be the best choice in the real applications where the solution temperature could be effectively adjusted and its moisture absorption ability could be improved. Effects of the membrane deformations on the transport phenomena in the PMC should be studied in future, which was a complex coupling problem between fluids, solids, momentum, heat, and mass transports. (2) For the HFMC, the counter flow HFMC was suitable for the small air flow rate, while the cross-flow one could control the pressure drop with the large air flow rate. The inclined flow type should be avoided to prevent the performance from deteriorating. The cross-flow elliptical hollow fiber membrane tube bank (EHFMTB) could be used to enhance the heat and mass transfer both in the tube and the shell side with the air stream flowing across the tube bank along the minor semi-axis. The internally-cooled HFMC was formed by installing the cooling coils into the solution channels between the neighboring short sub-contactors. The number and length of the sub-contactors and the area of the cooling coil should be optimized. Further, the performance correlations should be derived in future. Influences of the membrane deformations on the transport phenomena in the HFMC should be studied. The randomly distributions of the hollow fiber tubes should be avoided. The automatic processing technology of the regularly distributed HFMC would be a very meaningful promising technology, which were developed and generalized into market by the authors’ research group. (3) For the M-LDAD systems, the PMC and HFMC with better performances could be used in the systems. The internallycooled multistage M-LDAD system based on the internallycooled PMC and HFMC was the better choice in the real application. The membrane distillation (MD) for the diluted solution regeneration driven by the waste energy sources (industry waste heat) and the renewable energy sources (solar thermal energy, geothermal energy, etc.) might be a

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promising way. Although the performance of the natural convection M-LDAD system might be better because of omitting the solution pump, the response time of the startup and stop were long, which made the system difficult to be controlled. However the natural convection system could be used in the specific condition, which was safe enough and undemanding. The thermodynamic and dynamics of the internally-cooled multistage M-LDAD system should be investigated to match the energy between various components. (4) Although the membrane contactors used for LDAD could overcome the vital shortage of the liquid crossover problem encountered in the direct-contacting type, heat and mass transfer performances might be reduced because of the membrane separations between the air and the solution streams. It could be addressed by improving the permeability of the membrane and enhancing the heat and mass transfer in the air channel. It should be noted that the water vapor might be condensed in the air channel when the dew temperature of the hot and humid air was higher than the air side membrane surface temperature. Water droplets might be formed on the air side membrane surface, which would largely deteriorate the moisture transport across the membrane. Therefore the condensation of the water vapor in the air should be prevented. The internally-cooled M-LDAD system might be a promising method where the solution temperature can be easily adjusted by the refrigerant in the solution side. In addition, super hydrophobic modification of the air side membrane surface might be anther way.

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