Mar 2, 2018 - building ventilation systems due to their small footprint, simplicity, reduced ... Keywords: Energy recovery ventilation; Enthalpy exchanger; ...
Author’s Accepted Manuscript Heat and moisture transfer modeling in enthalpy exchangers using asymmetric composite membranes Amin Engarnevis, Ryan Huizing, Sheldon Green, Steven Rogak www.elsevier.com/locate/memsci
PII: DOI: Reference:
S0376-7388(17)33465-8 https://doi.org/10.1016/j.memsci.2018.03.007 MEMSCI16001
To appear in: Journal of Membrane Science Received date: 9 December 2017 Revised date: 2 March 2018 Accepted date: 4 March 2018 Cite this article as: Amin Engarnevis, Ryan Huizing, Sheldon Green and Steven Rogak, Heat and moisture transfer modeling in enthalpy exchangers using asymmetric composite membranes, Journal of Membrane Science, https://doi.org/10.1016/j.memsci.2018.03.007 This is a PDF file of an unedited manuscript that has been accepted for publication. As a service to our customers we are providing this early version of the manuscript. The manuscript will undergo copyediting, typesetting, and review of the resulting galley proof before it is published in its final citable form. Please note that during the production process errors may be discovered which could affect the content, and all legal disclaimers that apply to the journal pertain.
Heat and moisture transfer modeling in enthalpy exchangers using asymmetric composite membranes Amin Engarnevisa, Ryan Huizingb, Sheldon Greena, Steven Rogaka aDepartment
bdPoint
of Mechanical Engineering, The University of British Columbia, 2329 West Mall, Vancouver, BC V6T 1Z4, Canada.
Technologies Inc., 1455 E Georgia St, Vancouver, BC, Canada V5L 2A9
ABSTRACT Enthalpy exchangers using water vapor perm-selective membranes are used in building ventilation systems due to their small footprint, simplicity, reduced contaminant crossover, and relatively high efficiency. Moisture permeation properties of the membrane media that vary with operating air humidity and temperature lead to significant changes in the performance of an enthalpy exchanger and thus the whole ventilation system, impacting building energy efficiency. Evaluation of the actual energy savings potential of such energy recovery devices in building ventilation systems requires models that account for this variable membrane performance. A theoretical model is developed for current generation asymmetric composite membranes used in enthalpy exchangers. This model predicts the membrane permeability as a function of local values of air humidity and temperature, based on a limited number of kinetic water vapor sorption tests of the membrane material. The membrane model is coupled with a finite-difference model of the conjugate heat and mass transfer in full cross-flow enthalpy exchanger cores. The model predictions are validated against experimental data of a commercial-scale enthalpy exchanger. The model is used to predict the influence of outdoor air parameters (temperature, humidity) on an enthalpy exchanger and the predictions are compared against a baseline case that assumes constant membrane permeability. Such assumption can result in deviations in effectiveness predictions by up to 15%. Depending on the mode of operation, outdoor air relative humidity can increase or decrease the effectiveness of enthalpy exchangers by up to 12%. In contrast, outdoor air temperature appears to have only a minimal influence on effectiveness parameters.
Keywords: Energy recovery ventilation; Enthalpy exchanger; asymmetric composite membrane; modeling; operating conditions NOMENCLATURE
Symbols temperature wate vapor pressure ⁄
water vapor concentration water vapor activity
⁄
enthalpy of humid air
⁄
specific heat of dry air
⁄
air velocity
̇
⁄
mass flow rate
⁄
volumetric flowrate
⁄
latent heat of vaporization or condensation
⁄
moisture flux
⁄
heat flux
⁄
convective heat transfer coefficient
⁄
convective mass transfer coefficient mixing parameter partial vapor pressure flow area mass transfer resistance
⁄ ⁄
heat transfer resistance
⁄
specific gas constant ⁄
molecular weight
⁄
heat of adsoption ⁄
molar volume of gas at STP permeance permeability
solubility coefficient pore diameter number of pores per unit area ⁄
diffsuion coefficient ⁄
vapor diffusivity in air
⁄
Knudsen diffusion coefficient hydraulic diameter core width core height core depth triangular channel pitch triangular channel base Reynolds number Nusselt number fanning friction factor, ⁄ Lewis number Prandtel number dimensionless channel length
Greek Letters effectiveness, surface porosity aspect ratio, shape parameter humidity ratio (
)
dimensionless temperature dimensionless humidity ratio air density (
)
air dynamic viscosity (
)
latent-to-sensible heat ratio thickness thermal conductivity pore size distribution pore tortuosity, =2.5 fin correction factor
Superscripts M
membrane surface mass transfer heat transfer bulk air interface sensible adsorption vapor
Subscripts coating membrane substrate supply side exhaust side interface sensible latent total inlet outlet heat transfer √
based on square root of flow area
1. INTRODUCTION Energy required for air conditioning accounts for 30% to 40% of the primary energy consumption in the building sector. Air-to-air energy recovery ventilators (ERVs) recycle the sensible and latent energy in exhausted building air and use it to precondition the incoming outdoor ventilation air. Heat and moisture recovery using highefficiency ERVs in HVAC systems could save up to 65% of the energy used for fresh air treatment, while also improving indoor air quality and comfort [1], [2]. Membrane-based ERVs, also known as enthalpy exchangers, use selective water vapor permeable membranes that separate indoor and outdoor air streams. Compared to the incumbent technology (i.e., paper-based and desiccant-based ERVs), membrane-based enthalpy exchangers offer a smaller footprint, simplicity, reduced contaminant crossover, and high efficiency. shows the exchanger core of a typical membrane-based ERV. The supply and exhaust airstreams flow along the channels in a cross-flow pattern and exchange heat and moisture through the membrane. Alternate layers of membranes, separated and sealed, form the supply and exhaust airstream passages. Figure 1(a)
Exhaust Air Outlet
Supply Air Outlet
(a)
Coating layer: 1-5 µm thick
Supply Air Inlet (Outdoor fresh air)
Exhaust Air Inlet (Indoor stale air)
(b)
Microporous substrate: ~100 µm thick
Membrane
(c) Spacer
Figure 1. Geometry of the cross-flow fixed-plate enthalpy exchanger cores: (a) cross-flow core arrangement (in cooling mode), (b) spacer-filled membrane channels, (c) scanning electron micrograph of the asymmetric composite membrane cross-section [3].
Ideally, membranes used for ERVs should be highly permeable and selective for water vapor over other gases and contaminants that might be present in indoor and outdoor air streams. [4]. A practical membrane with high permeance can only be fabricated by making the selective polymer layer as thin as possible. Current generation ERV membranes typically consist of a polymeric microporous layer as the support for a relatively thin (1). (4) Since the membranes are rather thin (~ 100μm), heat and mass transfer through the membrane is considered one-dimensional (in thickness direction). (5) Heat and moisture transfer in the two air streams are two-dimensional (x and y directions) (6) Membrane surfaces are in dynamic equilibrium at the temperature and vapor pressure of the adjacent air.
3.3. Governing Equations Mass Conservation: Supply side: (
)
(26)
(
)
(27)
Exhaust side:
Energy Balance: Supply side:
(
)
̅
(28)
(
)
̅
(29)
Exhaust side:
where,
is humidity ratio in the air streams (
air velocity inside the flow channels (m/s), ( ⁄
),
and
),
is temperature,
is dry air density,
is the mean
is the specific heat
are the convective mass transfer coefficient ( ⁄ ) and convective heat
transfer coefficient ( ⁄ depicted in Figure 3.
), respectively, and
is the triangular channel height as
Boundary and Interface Conditions The changes of enthalpy along the x and y directions in the two air streams on supply and exhaust sides (Eq. (26) to (29) ) are coupled together by heat and mass transfer through the membrane plate, governed by the following equations. Membrane surface on the supply side: (
)
(
(
)
(30)
)
(
(31)
)
Membrane surface on the exhaust side: (
)
( Where,
(
)
(32)
)
(
(33)
)
is the water vapor concentration in the bulk air (
), and
and
are the total heat and mass transfer resistances, respectively.
(34)
(35)
The boundary conditions are defined as Supply side: |
,
(36)
|
Exhaust side: |
,
|
(37)
The dimensionless coordinates, temperature, and humidity ratio are defined as
(38)
(39)
(40)
(41)
3.4. Air-side Heat and Mass Coefficients Zhang et al. [47] showed that the hydrodynamic, thermal, and concentration entry regions account for a large fraction of the total duct length (up to 30%) in plate-type enthalpy exchangers. Therefore, the Nusselt correlations for both thermally and hydrodynamically developing flow for non-circular ducts by Muzychka and Yovanocic [48] are used.
⁄
(
√
[
√
)
({
√
(
)
⁄
⁄
}
{
(
√
√
(42)
)} ) ]
where, constants to are used to describe boundary conditions (uniform wall temperature) and Nu type (local), blending parameter, , is defined as a linear function of Pr, duct geometry is defined by the aspect ratio, , and shape parameter, ( and for an isosceles triangle when
), and the dimensionless length,
, defined
by ⁄
(43)
√
√
where √ is the characteristic length scale defined as the square root of the channel cross-sectional area. The local heat transfer coefficients are then calculated from the definition of the Nusselt number as
√
√
⁄
.
It is understood that the boundary conditions on membrane surfaces are neither uniform wall temperature (UWT) nor uniform wall flux (UWF), which are common approximations in heat transfer problems. Rather, boundary conditions are formed by the coupling between the supply and exhaust air streams. For cross-flow plate-type enthalpy exchangers, Zhang et. al [47] have shown that the boundary conditions lie between those under UWT and UWF. Therefore, the assumption of UWT in equation (42) is expected to introduce an error of less than 9% in approximation of the Nusselt number.
The Chilton-Colburn analogy is used to estimate the convective mass transfer coefficient from the heat transfer coefficient
⁄
(44)
where is the Lewis number, which is about 0.85 for ventilation air at temperatures between 0-40°C [19]. Fin effects of the corrugated aluminum spacers used to separate membrane layers in the studied geometry, representing plate-fin triangular channels of finite fin conductance, has been shown to significantly influence the fully developed Nusselt and Sherwood numbers in membrane channels of enthalpy exchangers [49], [50]. In fact, much smaller efficiencies of aluminum fins for moisture transfer compared to the heat transfer makes the validity of the analogy presented in equation (44) questionable. To address such fin effects, we have used a correction factor for constants and , that represent boundary conditions of equation (42), defined as
√
,
⁄ √
where
√
represents the fully-developed Nusselt number predicted by equation (42)
and is the fully-developed Nusselt number taken from the tabulated data in Zhang [50] for fins of different thermal conductivities; an isothermal fin (infinitely conductive) and an adiabatic fin (non-conductive) are assumed for obtaining heat and mass transfer coefficients, respectively. This correction method approximates variations of local Nu numbers to within 5% of the data in [50] for a channel aspect ratio of 0.5 and UWT boundary conditions with different fin conductance parameters.
3.5. Solution Procedures The forward finite difference scheme is used for discretization of the differential equations on both sides of the membrane. An iterative procedure used to solve the coupled equations is as follows: (1) Initialize the humidity and temperature fields on the two sides of membrane plate at inlet indoor and outdoor air conditions. (2) Calculate membrane surface humidity and temperatures through Eq.(30) to (33). (3) Calculate membrane permeation and heat transfer properties given these surface conditions. (4) Calculate humidity and temperature profiles of the bulk air streams (Eqs. (26) to (29)) on the two surfaces of the membrane plate given surface values calculated from the previous step. (5) Repeat steps (2) to (4) until the values of humidity and temperature fields for the bulk air streams are converged.
Numerical tests were conducted to determine the impact of grid size on the accuracy of the presented results. It was determined that 70 grids in each direction be adequate for the simulations (resulting in less than 0.1% difference from the case with 100 grids).
3.6. Enthalpy Exchanger Performance When the temperature and humidity fields in the exchanger are calculated, the sensible and latent effectiveness are accordingly calculated using mean outlet values assuming equal flow rates for supply and exhaust sides (i.e. a reasonable assumption for most real applications [16]). (
(
)
)
(45)
(46)
The total performance of the energy exchanger is consequently defined as (
where,
)
(47)
is the enthalpy of humid air.
4. RESULTS AND DISCUSSION
4.1. Moisture Permeation through Membrane Composite membrane permeance for water vapor at various supply and exhaust side activities (i.e., the ratio of vapor pressure to saturation vapor pressure) are plotted in Figure 4. It can be observed that the membrane permeance is independently a function of supply and exhaust activities.
Figure 4. Permeance at various supply and exhaust side activities at T = 35°C.
The impact of membrane orientation on the permeation through asymmetric composite membranes can also be realized from the asymmetry of contours about the 1:1 line (dashed line) in Figure 4. Membrane permeance is consistently higher when the dense coating layer is exposed to the feed side with higher humidity content. This difference becomes more substantial at higher feed side activities, lower permeate-to-feed activity ratios, and higher temperatures. It will be shown in section 4.4 that the membrane orientation will result in different performance variations of an enthalpy exchanger in cooling and heating modes. shows the impact of operating temperature on membrane permeance at various levels of supply activity. Higher operating temperatures, on average, result in a moderate decrease in permeability, which is caused by the decreased sorption capacity of dense polymer dominating its diffusivity increase. However, at very low temperatures, such as 5ºC in Figure 5, it can be seen that permeability decreases significantly, especially at very high activities. This is attributed to the high sorption capacity of material at very low temperatures resulting in significant clustering and so a drop in diffusivity dominating the increased sorption at very low temperatures [51]. Figure 5
a) Exhaust activity
b) Exhaust activity Figure 5. Effect of temperature on water vapor permeance at various supply side activities
4.2. Validation of the Enthalpy Core Model The model results were verified by comparing the calculated effectiveness of a commercial-size core (see Table 2 for specifications) with the experimental performance measurements of the core tested under AHRI summer and winter test conditions (as specified in Table 3) at various operating flow rates.
Table 2. Geometrical specifications of the enthalpy exchanger core tested in this study.
Footprint (Dc×Wc) (mm×mm) Height (Hc) (mm) Membrane Flow spacer Number of layers Pitch (a) (mm) Base (b) (mm) Apex angle (α) (degree)
552×552 200 MX4 (see section 2.2) Corrugated Aluminum sheet 72 2.6±0.1 7.4±0.1 54±2
Table 3. AHRI summer test conditions.
Cooling mode (summer)
Heating mode (winter)
Parameter
Outdoor Air
Indoor Air
Outdoor Air
Indoor Air
Dry-bulb Temperature (◦C)
35
24
1.7
21
Wet-ball temperature (◦C)
26
17
0.5
14.5
Relative Humidity (%)
50
50
78
50
The results of two sets of measurements of sensible and latent effectiveness for each mode of operation along with the model predictions are presented in Table A-1 in Appendix. A summary of these results is plotted in Figure 6. Discrete symbols represent measured data, and solid and dashed lines represent calculated sensible and latent effectiveness, respectively. Black symbols show calculated effectiveness values from the difference in measured air parameters on the supply side while hollow symbols show those from the exhaust side differences. The maximum deviations are 6% for sensible effectiveness in cooling mode and 10.2% for latent effectiveness in heating mode, indicating that effectiveness is satisfactorily predicted by the model. The results of both measurements and the model predictions presented in Figure 6 suggest that sensible and latent effectiveness are both affected by weather conditions and are consistently higher for the heating mode. Correspondingly, the total effectiveness is also higher for the heating condition. This is due to the higher permeability of the membrane at lower temperatures (as was shown in Figure 5) and
stronger heat and mass transfer coupling in the cooling mode, the latter due to much larger moisture fluxes through the membrane. This behavior is partly attributed to the material properties of the particular polymer membrane tested in this study and can be significantly different for another membrane with different permeation properties, such as stronger temperature-dependency [52] or increasing diffusivity with moisture content of membrane. Koester et al. [53] discuss such variations of water vapor permeance through different membranes suitable for use in enthalpy exchangers.
a) Cooling mode
b) Heating mode
Figure 6. Variations of enthalpy exchanger performance with air flow rate. The experimental measurements are shown by symbols, and the corresponding model predictions are shown by the faired curves.
4.3. Effect of Variable Membrane Permeability The impact of variable permeability on effectiveness is illustrated in Figure 7 (a) and (b), for cooling and heating modes with a constant operating temperature and variable outdoor relative humidity. Two scenarios are considered in the plot 1) variable membrane permeability, 2) a constant membrane permeability estimated using ideal laminate theory from individual measurements of coating and substrate permeability at 50oC and 50%RH test conditions.
a) Cooling mode
b) Heating mode Figure 7. Variations of enthalpy exchanger performance with outdoor air RH; comparison of constant permeability model and composite membrane model with variable permeability
Assuming a constant permeability can result in deviations in effectiveness predictions of up to 15%. These deviations would lead to significant errors in predicting potential energy savings from ERV installation in buildings. Figure 8 shows the differences in moisture flux prediction over the membrane surface for a constant permeability model (dashed lines) and the composite model (solid lines) during cooling and heating modes.
a) Cooling mode (Toutdoor = 35°C)
b) Heating mode (Toutdoor = 1.7°C) Figure 8. Variations of moisture flux through membrane surface; comparison of constant permeability model (dashed curves) and composite membrane model with variable permeability (solid curves) at 50% outdoor relative humidity.
4.4. Effect of Operating Humidity and Temperatures In a theoretical study by Min et. al [54], the coupled heat and moisture transfer in a plate-type membrane-based enthalpy exchanger was investigated over a wide range of weather conditions. They have shown that depending on the operating conditions, heat and moisture may transfer either in the same direction, or opposite directions, or even partially the same and partially opposite directions over the membrane surface. They further discuss that when the heat and moisture transfer in opposite directions, the enthalpy effectiveness loses its significance since the concept of enthalpy effectiveness is valid only in cases where heat and moisture transfer have the same directions. Similar results are also reported by Simonson and Besant for performance variations of an enthalpy wheel [55]. Therefore, the range of variations of outdoor air relative humidity and temperature in our study is limited to those cases where the heat and moisture transfer in the same direction across the entire membrane surface; from the supply outdoor air to exhaust indoor air during the cooling mode, and vice versa during the heating mode. shows the variation of enthalpy exchanger effectiveness with outdoor relative humidity at three different temperatures for cooling mode operation. Figure 10 is an equivalent plot for heating mode operation. Outdoor air temperature has minimal influence on effectiveness. In both heating and cooling modes, at lower relative humidities, higher operating temperatures result in a slightly lower sensible and latent effectiveness while at higher relative humidities this trend is reversed. Figure 9
Figure 9. Variations of enthalpy exchanger performance with outdoor air state in cooling mode 3 (flowrate = 544 m /hr)
Figure 10. Variations of enthalpy exchanger performance with outdoor air state in heating mode 3 (flowrate = 544 m /hr)
Unlike temperature, outdoor relative humidity has a larger influence on both sensible and latent effectiveness in the cooling mode. The latent effectiveness increases from 45% to 50% as the relative humidity increases from 30% to 90%. Sensible effectiveness decreases slightly from 68.4% to 67.7% with similar variations of relative humidity. This is attributed to the heat and mass transfer coupling effect due to the relatively large moisture flux through the membrane in cooling mode. Total effectiveness is correspondingly varied in this operating humidity range. At lower values of relative humidity its variations are dictated by the decrease in sensible effectiveness while it approaches the latent effectiveness at higher relative humidity. On the other hand, under similar variations of relative humidity in the heating mode, sensible effectiveness remains nearly unchanged while latent effectiveness slightly increases from 46.5% to 48.4% at an operating temperature of 1.7oC. As a result, total effectiveness monotonically increases from 57.5% to 60.6%. The variations in effectiveness with temperature and relative humidity discussed above can be explained by the variations of membrane moisture transfer resistance predicted by the asymmetric composite membrane permeability model. Figure 11 shows an
example of these variations in moisture transfer resistance of the membrane for cooling and heating modes with 30% outdoor air relative humidity.
Figure 11. Contours of local membrane moisture transfer resistance, ( ⁄ ); solid and dashed lines indicate cooling (Toutdoor = 35°C) and heating (Toutdoor = 1.7°C) modes, respectively. (RHoutdoor = 30%)
shows an averaged value of moisture transfer resistance over the membrane plate surface at each operating condition. For higher supply side (outdoor) relative humidity at a fixed exhaust (indoor) RH (which is the case for both cooling and heating conditions), the membrane has lower resistance to moisture transfer resulting in an increase in latent effectiveness in the cooling mode. For the heating mode, this trend is reversed at very high relative humidity of the supply side (RH>70%) which explains the nearly constant latent effectiveness of the core for higher humidity in this mode (see Figure 10). This is believed to be due to the significant clustering of water molecules in the membrane coating at lower temperatures (i.e., resulting in higher solubility (see ref [40]) and thus moisture content inside the polymer) lowering overall membrane permeability. Figure 12
Figure 12. Variations of average membrane moisture transfer resistance with outdoor air state; hollow and black symbols indicate cooling and heating conditions, respectively.
4.5. Effect of Membrane Orientation compares the latent effectiveness calculated for two different scenarios: (I) coated side of membrane exposed to the supplied outdoor air referred to as ‘Membraneon-supply’, (II) coated side of membrane exposed to the exhausted indoor air referred to as ‘Membrane-on-exhaust’ (as shown in Figure 13(a)). Indoor air conditions in these scenarios are fixed according to the AHRI test conditions (see Table 3). Figure 13(b)
(a) Membrane configuration; (I) Membrane-on-supply, (II) Membrane-on-exhaust
(b) latent effectiveness Figure 13. Membrane orientation effect at various outdoor air relative humidities; outdoor air temperatures are 35°C and 1.7°C for cooling and heating modes, respectively.
In the cooling mode, exposure of the membrane to the exhaust air (i.e., Membrane-onexhaust scenario) results in a slightly lower latent effectiveness of the core while it increases the latent effectiveness in the heating mode. The sensible effectiveness in both cases is almost independent of membrane orientation (See Figures A-1 and A-2 in Appendix). 5. CONCLUSIONS The impact of operating conditions (temperature and relative humidity) on the performance of cross-flow membrane-based enthalpy exchangers was studied through material testing and exchanger core modeling. A generalized model for conjugate heat and moisture transfer through asymmetric composite membranes used in the state-of-the-art enthalpy exchangers was developed. The model takes into account permeation properties of the dense polymeric coating layer, extracted from a limited number of kinetic sorption experiments, and the microstructure of the substrate layer, to predict variable membrane permeability in a wide range of operating conditions. It was shown that membrane permeability is a strong function of the relative humidity of the air streams on both sides of the
membrane and a weaker function of operating air temperatures and membrane orientation. A finite-difference heat and mass transfer model was developed to predict the systemlevel enthalpy exchanger performance. This model takes into account the variable membrane permeability as a function of local operating humidity and temperature values on the membrane surfaces. The model was validated against experimental measurements on a commercial enthalpy exchanger. Depending on the mode of operation, membrane permeability variance was shown to influence predicted performance of the exchanger core by up to 15% compared to the case with constant permeability. Other conclusions from this study are: In general, both sensible and latent effectiveness (and thus the total effectiveness) are stronger functions of operating relative humidity and weaker functions of the operating temperature and core orientation. In the cooling mode, both sensible and latent effectiveness depend strongly on the outdoor air relative humidity. Sensible effectiveness decreases with increasing outdoor relative humidity while latent effectiveness increases. At RH values lower than 50%, total effectiveness is more closely aligned with sensible effectiveness while at higher RH values it lies closer to the latent effectiveness. In the heating mode, due to the very low moisture transfer in the enthalpy exchanger, sensible effectiveness remains nearly constant. Latent effectiveness slightly increases with outdoor relative humidity and has a higher value compared to the cooling mode for outdoor relative humidities below 60%. Relative to the baseline case of exposure to the outside air of the membrane coated surface, exposure of the membrane substrate results in a slightly lower latent effectiveness of the core in the cooling mode while it has a more pronounced, beneficial effect on the latent effectiveness in the heating mode. Operating air temperature appears to have minimal impact on the core performance in both heating and cooling modes. ACKNOWLEDGEMENTS This work was funded by the Natural Sciences and Engineering Research Council of Canada (NSERC) through a CRD Grant-CRDPJ485151-15 –sponsored by DPoint Technologies Inc. APPENDIX 1. Validation Data of the Enthalpy Core Model
Table A-1 summarizes the modeling and experimental results for the studied commercial-size enthalpy exchanger core. The experimental data is taken from the dataset provided by dPoint Technologies Inc., based on test reports from Lucerne laboratory (Lucerne University (HSLU), Switzerland), a certified laboratory to test to AHRI 1060 and ASHRAE 84 standards.
Table A-1. Validation core performance measurements and model data Mode
Experimental supply
Cooling (Summer)
Heating (Winter)
Q( 264 401 531 661 801 279 417 561 695 826
exhaust
)
Q( 79.22 73.09 69.40 67.31 65.60 74.55 70.80 68.63 64.73 64.22
61.20 55.35 50.11 46.60 42.62 57.14 51.18 45.95 42.09 39.13
Model
70.42 65.26 61.57 59.15 57.03 62.85 57.35 53.71 49.63 47.44
2. Membrane Orientation Effects plots
254 390 518 641 780 283 424 567 712 839
(Membrane-on-supply)
)
Q( 72.91 68.99 66.59 64.52 63.01 70.23 67.89 65.45 64.22 62.45
62.20 55.75 52.21 49.11 46.80 54.02 47.87 43.48 40.66 38.09
66.83 63.54 60.53 58.91 56.90 59.55 54.32 49.96 47.85 45.68
259 395 524 651 791 281 420 564 704 833
) 76.96 71.84 67.85 64.46 61.20 76.76 71.80 67.68 64.19 61.36
62.52 54.71 48.93 44.30 40.13 59.87 52.06 46.02 41.30 37.72
71.72 65.62 60.99 57.14 53.55 65.66 58.82 53.44 49.13 45.82
Figure A-1. Membrane orientation effect in cooling mode (Toutdoor = 35°C, flowrate = 544 m3/hr)
Figure A-2. Membrane orientation effect in Heating mode (Toutdoor = 1.7°C, flowrate = 544 m3/hr)
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Highlights
1- A mathematical model is introduced for conjugate heat and moisture transfer through current generation asymmetric composite membranes used in enthalpy exchangers 2- The model is validated against experimental data of a commercial-scale enthalpy exchanger core 3- Variability in membrane permeability is shown to influence predicted performance of an enthalpy exchanger core by up to 15% compared to the case with constant membrane permeability. 4- Depending on the mode of operation, enthalpy exchanger effectiveness can increase or decrease with outdoor air relative humidity and temperature 5- Impact of membrane orientation on the exchanger latent effectiveness is disclosed.
