Numerical analysis of heat and energy recovery ventilators performance based on CFD for detailed design

Applied Thermal Engineering 51 (2013) 770e780

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Applied Thermal Engineering journal homepage: www.elsevier.com/locate/apthermeng

Numerical analysis of heat and energy recovery ventilators performance based on CFD for detailed design Wahiba Yaïci*, Mohamed Ghorab, Evgueniy Entchev Renewables and Integrated Energy Systems Laboratory, CanmetENERGY Technology Centre, Natural Resources Canada, 1 Haanel Drive, Ottawa, Ontario K1A 1M1, Canada

h i g h l i g h t s < Performance of heat and energy recovery ventilators (HRVs/ERVs) is investigated. < The numerical simulations are performed for summer and winter Canadian climate. < The effect of a wide range of parameters on ventilator’s effectiveness is analysed. < CFD simulations proved to be an effective tool for detailed design of HRVs/ERVs.

a r t i c l e i n f o

a b s t r a c t

Article history: Received 28 May 2012 Accepted 6 October 2012 Available online 16 October 2012

This paper presents a detailed numerical analysis of heat and membrane-based energy recovery ventilators (HRV/ERV), using computational fluid dynamics (CFD). The CFD model includes conjugate heat and mass transfer mechanisms for laminar flow to investigate the thermal performance of these systems. Both co-current and counter flow designs were investigated under typical summer/winter Canadian conditions. The model was validated with data obtained from the open literature. It was then applied to investigate the effect of a wide range of parameters on ventilators sensible and latent effectivenesses. The numerical results confirmed the superior effectiveness of counter flow over the co-current flow. The results showed a decrease in the HRV/ERV effectiveness with the increase in supply/exhaust air velocity. The ERV effectiveness in the summer was found to be higher than the winter. The effectiveness decreased noticeably and slowly with the increase in membrane spacing and thickness. The latent effectiveness increased significantly with the diffusivity of water in the membrane. The results also indicated that the outdoor temperature and humidity had only minor effects on HRV or ERV performance. CFD simulations proved to be an effective approach for detailed design of HRVs/ERVs. Ó 2012 Elsevier Ltd. All rights reserved.

Keywords: Energy recovery Heat transfer Mass transfer Membrane Effectiveness Modelling CFD simulation

1. Introduction Buildings and their HVAC systems are required to be energy efficient, while satisfying the ever-increasing demand for better indoor air quality, performance and environmental preservation. The goal of HVAC design in buildings is to provide good comfort and air quality for occupants during a wide range of outdoor conditions. Sufficient fresh air supply by ventilation is necessary for occupants health [1,2]. In most industrialised countries, energy consumption by the HVAC sector accounts for 33% of the total energy consumption [1,3]. HVAC of houses and buildings accounts for 23% of Canada total energy use and 22% of its total GHG emissions [4]. Cooling and dehumidifying fresh ventilation air constitutes 20e * Corresponding author. Tel.: þ1 613 9963734; fax: þ1 613 9470291. E-mail address: [email protected] (W. Yaïci). 1359-4311/$ e see front matter Ó 2012 Elsevier Ltd. All rights reserved. http://dx.doi.org/10.1016/j.applthermaleng.2012.10.003

40% of the total energy load in hot and humid regions. Total heat recovery from ventilation air has nowadays become an important topic [3]. There are many researches aimed at improving the HVAC systems in buildings while reducing the energy costs and environmental impacts. The development of energy recovery ventilator (ERV) systems over the last few decades has led to improved performance and capability in recovering both sensible and latent energy, where the latent load constitutes a large fraction of the total thermal load in the HVAC system. It has been found that 70e90% of the energy for fresh air treatment could be saved. With ERV systems, the efficiency of an existing HVAC system can also be improved [1,5e8]. A heat recovery ventilator (HRV) also known as an air-to-air exchanger is a ventilation system that employs a co-current, cross-flow or counter-current flow heat exchanger between the inbound and outbound airflow. An HRV provides fresh air and

W. Yaïci et al. / Applied Thermal Engineering 51 (2013) 770e780

Nomenclature As AH c C d D E F h H* K k l L LMTD LMWD m N n NTU P Q q Re R RH T

heat/mass transfer surface area, m2 absolute humidity, g of moisture/kg of dry air molar concentration of water moisture, mole/m3 heat capacity ratio membrane spacing, mm diffusivity of water in membrane, m2/s fan power, W volume force vector, N/m3 specific enthalpy, J/kg ratio of latent to sensible energy differences between the inlets of two airstreams overall heat transfer coefficient, m/s thermal conductivity, W/m K core length, mm channel length, mm log mean temperature difference ( C) log mean humidity ratio difference (kg/kg dry air) mass airflow rate, kg/s mass flux, kg/m2 s number of channels or normal vector to the boundary number of heat transfer units pressure, Pa total heat transfer, W heat flux, W/m2 Reynolds number reaction rate, kg/m3 s relative humidity, % temperature, K

improved climate control, while saving energy by reducing the heating or cooling requirements. HRVs, as their name implies, recover the heat energy in the exhaust air, and transfer it to fresh air as it enters the building. An ERV is a type of air-to-air heat exchanger that not only can transfer sensible heat but also latent heat. Since both temperature and moisture are transferred, ERVs can be considered total enthalpy devices. An HRV is limited to only transferring sensible heat [1,8]. A summary of the studies on heat recovery technologies for building applications is available in the review paper [9]. An ERV comprises an enthalpy exchange core, two airflow passages and two fans, encased in a housing. Supply and exhaust streams are drawn by the fans to flow through the core, to transfer both heat and moisture from one airstream to the other. The core is made of vapour-permeable membranes like micro-porous polymeric membrane. The main advantages of a membrane system are that it is a static device that does not involve any moving parts, simple to construct and can be easily integrated onto existing conditioning systems [1,10,11]. During the last decade, researchers have been active in studying the performance of membrane heat exchangers for HVAC system. The effectiveness of an enthalpy exchanger under various conditions should be calculated in order to optimise the membrane system. Due to the complex coupling between heat and mass transfer, there is no simple design methodology available for the membrane-based enthalpy exchangers. Zhang and Jiang [6] have developed a numerical model for heat and moisture transfer of membranes with different core and airflow arrangements. Simonson and Besant [11,12] have proposed an approach for energy wheel design by developing useful effectiveness correlations that can help manufacturers and HVAC engineers to produce and select efficient energy wheels. Niu and Zhang [13] have developed a theoretical

u U V V W DP Cp

771

velocity field, m/s overall heat transfer coefficient, W/m2 K air velocity, m/s airflow rate, m3/s humidity ratio, kg of moisture/kg of dry air pressure drop, Pa specific heat, J/kg K

Greek symbols membrane thickness, mm vector operator 3 effectiveness m dynamic viscosity, Pa s r density, kg/m3

d D

Subscripts and superscripts A air DB dry bulb e exhaust H enthalpy i inlet L latent m membrane min minimum o outlet s supply S sensible T transpose of velocity gradient tensor w water

model to describe the heat and mass transfer in the core of a membrane-based ERV. They investigated the effects of air states and membrane adsorption characteristics on the ERV effectiveness and reported that the entering air humidity influenced the ventilator effectiveness but not the sensible effectiveness. Zhang and Niu [13] further developed performance correlations for estimations of membrane-based ERV on cross-flow arrangement. Zhang et al. [14] have studied the effects of operating conditions and material properties on sensible and latent efficiency in application scale parallel-plates enthalpy with novel membrane material. Min and Su [16] have developed a mathematical model to analyse the performance of a membrane-based ERV. They investigated the effects of the ERV core geometric parameters namely the air fluid channel spacing and membrane thickness on the ERV effectiveness and analysed the effects of the membrane parameters on the ERV effectiveness. They interpreted the results in terms of the heat and moisture transfer resistances through the membrane and found that both resistances were affected by not only the membrane parameters but also the operating conditions. More recently, Min and Su [17] have performed a theoretical study to investigate the effects of the outdoor air temperature and humidity on the performance of a membrane-based energy recovery ventilator operating in both hot and cold seasons. The results showed that the outdoor air temperature and humidity affect not only the moisture transfer resistance but also the heat transfer resistance through membrane. The effectiveness-NTU method has been used to develop correlations to estimate the sensible, latent and total heat exchanger performance. The correlations are specific to the operating conditions and the design of HRV/ERV systems, and cannot disclose the insight into the mechanisms of heat and mass transport. Numerical modelling has become an efficient tool to analyse

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energy recovery in such systems. Furthermore, the majority of studies were performed on rotary energy wheels or on cross-flow membrane heat exchangers utilising CFD simulation to investigate the heat exchanger performance and analyse the heat and moisture transfer mechanisms [1,10,11,13,15,18e23], but to the best of our knowledge, few studies have been reported in the open literature on HRV or membrane-based ERV operating on co-current and counter-current flow arrangements for the prediction of their performance using CFD tools. Despite their usefulness, these studies suffer from the limitative nature of the assumptions on which they are based, preventing a detailed analysis for optimal design for any flow configuration and operating conditions. The work presented in this paper is an attempt to represent the performance of an HRV and a membrane-based ERV more generally and accurately using CFD modelling of the conjugate heat and mass transfer while taking into account different flow designs, membrane and air physical properties. The study aims to compare the performances of HRV and membrane-based ERV systems operating on either co-current or counter-current flow arrangements under typical Canadian summer and winter conditions. Effects of a wide range of parameters such as the supply and exhaust air velocities, membrane thickness and spacing, water diffusivity through the membrane and influence of outdoor ambient air temperature and humidity on ventilators performance will be investigated. Useful design information and performances data for different geometries, properties and conditions will be extracted.

Fig. 1. Schematic of a membrane-based energy recovery ventilator: (a) Core in counter flow arrangement; (b) schematic of the physical model in co-current and counter flow arrangements.

2. Simulation model details 2.1. Physical model Fig. 1a illustrates the geometric model of an exchanger used in membrane-based energy recovery ventilators for air conditioning. The physical model consists of supply and exhaust stream channels and a membrane. The supply and exhaust airstreams flow along the channels and exchange heat and moisture in a co-current or counter-current flow arrangements. As mentioned, the difference between total enthalpy exchangers and air-to-air heat exchangers is that in place of metal materials, hygroscopic materials, like paper and polymer membranes are used as the plate materials [1,8]. Alternate layers of membranes, separated and sealed, form the supply and exhaust airstream passages. As the system design is symmetrical, the computational domain used in the simulations represents half the volume of the supply airstream, a membrane, and half the volume of the exhaust airstream, as shown in Fig. 1b.

As the process is assumed to be in steady state, time dependent parameters are dropped from the equations. The resulting equations are as follows: 2.2.1. Conservation of mass

D$ðruÞ ¼ 0

(1)

2.2.2. Conservation of momentum

h



rðu$DÞu ¼ D$  p þ m Du þ ðDuÞT

i

þF

(2)

where u is the velocity vector, r is the density, m is the dynamic viscosity, p is the pressure, T is the transpose of velocity gradient tensor, and F is the volume force vector. As the flow is laminar, there is no force field, F ¼ 0 in Eq. (2).

2.2. Governing equations Two-dimensional (2D), steady-state co-current and countercurrent flow models were built in the CFD COMSOL [24] software. The governing equations of mass, momentum and energy conservation were solved by using the finite element method, based on the following assumptions: - The heat exchangers operate under steady-state conditions; - Air is incompressible and its physical properties are constant; - Water heat conductivity and diffusivity in the plates are constant; - There is no lateral mixing of the two airstreams and no phase change occurs; both fluids are single phase and are unmixed; - Heat conduction and vapour diffusion in the two airstreams are negligible compared to energy transport and vapour convection by bulk flow.

2.2.3. Heat transfer The conjugate heat transfer and laminar flow is used to model slow-moving flow in the HRV/ERV where temperature and energy transport are coupled. Eqs. (1) and (2) are solved together with an energy balance in 2D. The governing equation of heat transfer in fluids is:

rCp u$DT ¼ D$ðkDTÞ þ Q

(3)

where Cp is the heat capacity at constant pressure, T is the temperature, k is the fluid thermal conductivity, and Q is the heat source (or sink), which is zero. 2.2.4. Mass transfer For the ERV membrane, the governing equation of mass transfer by diffusion and convection are:

W. Yaïci et al. / Applied Thermal Engineering 51 (2013) 770e780

773

D$ðDDcÞ þ u$Dc ¼ R

(4)

us jx¼0 ¼ Vsi

(15)

N ¼ DDc þ uc

(5)

Ts jx¼0 ¼ Tsi

(16)

where c is the molar concentration of water, D is the diffusivity of water in the membrane, N is the mass flux, and R is the reaction rate, which is zero.

cs jx¼0 ¼ csi

(17)

2.3. Boundary conditions No slip condition is imposed at the membrane surface, i.e., the velocity of the fluid at the wall is zero. That is, the fluid at the wall is not moving. The condition prescribes:

u ¼ 0

(6)

No penetration or no cross-flow is imposed on the solid wall boundaries, i.e., at the centre of the supply and exhaust channels and at the symmetry plane. From a modelling point of view, this may be reasonable approximation if the important effect of the wall is to prevent fluid from leaving the domain. The constraint is formulated as to set the velocity component normal to the wall to zero:

u$n ¼ 0

(7)

   t$  p þ m Du þ ðDuÞT n ¼ 0

(8)

where n is the normal vector and t is the tangential vector to the boundary. The Dirichlet condition is applied for the pressure and there are no viscous effects at the outlet of the HRV/ERV.

p ¼ p0

mðDu  DuÞT n ¼ 0

(9) (10)

The zero heat flux boundary condition is assigned to the centre of the supply and exhaust channels and at the symmetry plan, assuming the plan wall is perfectly insulated on outside. The equations for this condition for the heat and mass transfers are as follows:

n$ðkDTÞ ¼ 0

(11)

n$ðDDc þ ucÞ ¼ 0

(12)

It is assumed that there is continuity of the heat and mass fluxes at the membrane interface, and the boundary condition for the HRV/ERV outlet is convective flux for heat and mass transfers:

n$ðq1  q2 Þ ¼ 0

(13)

n$ðN1  N2 Þ ¼ 0

(14)

where q is the heat flux, N is the mass flux. The supply and exhaust air velocities, temperatures and water concentrations, are defined as the inlet parameters that will be specified in the next section. The inlet velocity, temperature and concentration profiles are assumed to be uniform. Zero gauge pressure is assigned to the outlet of the HRV/ERV channels in order to obtain the relative pressure drop between inlet and outlet. The inlet conditions prescribe as follows:Supply air for the coand counter-current flows:

Exhaust air for the counter-current flow:

ue jx¼L ¼ Vei

(18)

Te jx¼L ¼ Tei

(19)

ce jx¼L ¼ cei

(20)

Exhaust air for the co-current flow:

ue jx¼0 ¼ Vei

(21)

Te jx¼0 ¼ Tei

(22)

ce jx¼0 ¼ cei

(23)

2.4. Mesh generation Mesh generation is performed, using COMSOL [24]. The surfaces of the HRV and ERV geometric models are meshed using quadrilateral elements. Normal size is used for the 2D models, giving approximately 8000 elements. A mesh independence study of the CFD models is investigated by modifying the mesh density. Coarse, normal, fine and finer grids were tested. It is found that for both models, refining the mesh with fine grid and finer grid provides negligible changes of the main field variables. The relative errors of the computed supply and exhaust outlet temperatures and moisture outlet concentrations obtained using the coarse, fine and finer grids relative to the results obtained using normal grid, were 5%, 0.5% and 0.5%, respectively. This difference between the normal and fine or finer grids was assumed small enough to consider calculations with the normal mesh as grid independent. Therefore, the results obtained from normal grid can be utilised for further heat transfer and fluid flow analysis in the present study. Furthermore, the mesh has been checked for cell skewness and aspect ratio. The convergence criterion has been set to the tolerance of 106 and all residuals were run down to 106, in all cases.

3. Model validation Validation was performed for both the heat and energy recovery ventilators (Table 1). The empirical correlations for co- and counter-current heat exchangers without mass transfer in Eqs. (25) and (26) and given by [25] are used to validate the current numerical model, where C is heat capacity ratio, NTUS is the number of heat transfer units and Q is sensible heat transfer, as presented in Eqs. (26)e(28).For countercurrent flow: 3S

¼

1  exp½  NTUS ð1  CÞ 1  Cexp½  NTUS ð1  CÞ

(24)

For co-current flow: 3S

¼

1  exp½  NTUS ð1 þ CÞ 1þC

(25)

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W. Yaïci et al. / Applied Thermal Engineering 51 (2013) 770e780

Table 1 Input data and dimensions. Parameters

Supply air TDB ( C) RH (%) AH (g/kg) Exhaust air TDB ( C) RH (%) AH (g/kg) Core dimensions L (mm) d (mm) d (mm) Reynolds number for d ¼ 2 mm Re for V ¼ 0.5 m/s Re for V ¼ 2.5 m/s a

a

Validation cases (HRV/ERV)

HRV/ERV study base cases Summer

Winter

35 64 23.02

30 61 16.38

10a 50 0.80

27 47 10.50

26 52 10.96

23 50 8.78

250 2 0.1

250 2 0.1

250 2 0.1

58 291

63 316

103 541

b

Mean temperature for the Canadian climate.

C ¼

Cmin ðmCpÞmin ¼ Cmax ðmCpÞmax

NTUS ¼

UAs Cmin

Q ¼ ðmCpÞs DTs ¼ ðmCpÞe DTe ¼ As U DTLMTD

(26)

(27)

(28)

The sensible effectivenesses are calculated using Eqs. (24) and (25) across a wide range of air velocity (from 0.5 m/s to 2.5 m/s). The heat capacity ratio of 0.956 is used for the validation study. Fig. 2a and b show a comparison of effectiveness results obtained from the present CFD modelling and the empirical correlations at different NTUS values. The results show that there is a very good agreement between the CFD and empirical equation results with accuracy more than 98%. As a result, the current model can be utilised for further HRV analysis with supporting reliability of the calculations. By analogy to heat transfer, the definition of the number of transfer units for heat is used for mass transfer that accounts for the moisture transfer across the membrane of the ERV [6,11,12,15,26,27]:

NTUL ¼

As K



(29)

V min The overall mass transfer coefficient through the membrane is calculated by:

K ¼

VðWsi  Wso Þ DWLMWD

(30)

Similar to the sensible effectiveness, the latent effectiveness is calculated using the following correlations:For counter-current flow: 3L

¼

1  exp½  NTUL ð1  CÞ 1  Cexp½  NTUL ð1  CÞ

(31)

For C ¼ 1: 3L

¼

NTUL 1 þ NTUL

(32)

Fig. 2. Variation of effectiveness for HRV and ERV with NTU for validation study: (a) Co-current flow; (b) counter flow.

For co-current flow:

3L

¼

1  exp½  NTUL ð1 þ CÞ 1þC

(33)

As for the HRV CFD model validation, the membrane-based ERV CFD model is validated using Eqs. (29)e(33) across a wide range of air velocity for the counter and co-current flows. Fig. 2a and b show a comparison of both sensible and latent effectiveness results obtained from the present ERV CFD modelling and the empirical correlations at different NTU values. The results show that there is a very good agreement, within 2%, between the CFD results and empirical equations. This demonstrates the suitability of the CFD models in predicting the performance of an ERV with a good accuracy. Therefore, the CFD models can confidently used as design tool. Two cases were simulated for typical Canadian winter and summer seasons for both the heat and energy recovery ventilators (Table 1). The supply air temperatures have respectively, dry temperatures of 30  C for the summer and 10  C for the winter season. The exhaust air temperatures have respectively, dry temperatures of 26  C for the summer and 23  C for the winter season. The supply air temperatures for the winter case represent mean temperatures for the winter season. The supply and exhaust air velocities were in the range of 0.5e2.5 m/s in the exchanger channel and the diffusivity of water in the membrane was set at 8  106 m2/s [23]. The basic core dimensions were taken from [16,23] and are given in Table 1. Table 2 shows the summary of the simulation matrix for the parametric study. From the temperature and concentration of moisture fields obtained from the CFD simulations, integration over the boundaries

W. Yaïci et al. / Applied Thermal Engineering 51 (2013) 770e780

775

Table 2 Summary of the simulation matrix. Parameters effects

Figures

HRV Vs/Ve (m/s) Ts/RH ( C/%)

Sect. 4.1 Fig. 3a Fig. 4a Fig. 4b Sect. 4.2 Fig. 5a Fig. 5b Fig. 6a Fig. 6b Fig. 7a Fig. 7b Fig. 8a Fig. 8b Fig. 9a Fig. 9b Fig. 10a Fig. 10b

ERV Vs/Ve (m/s) Ve (m/s) Ts/RH ( C/%) D (mm)

d (mm) D (m2/s)

Cocurrent flow

Countercurrent flow

Summer

Winter

V (m/s)

X

X X

X X

X

X

X X X X X

X X X X X X

X X X X X X X X X X X X

X

¼

ms ðTsi  Tso Þ þ me ðTeo  Tei Þ 2ðmÞmin ðTsi  Tei Þ

X X X X

¼

ms ðWsi  Wso Þ þ me ðWeo  Wei Þ 2ðmÞmin ðWsi  Wei Þ

¼

ms ðhsi  Hso Þ þ me ðheo  hei Þ 2ðmÞmin ðhsi  hei Þ

X X

X X

X X

d (mm)

X X

X X

X

(35)

(36)

where h is the specific enthalpy. The sensible, latent and enthalpy energy transfer effectivenesses are also calculated using the following simplified equations for comparison with Eqs. (34)e(36) at constant flow properties and equal supply and exhaust flow rates [1,15,23]:

3S

¼

Tsi  Tso Tsi  Tei

(37)

3L

¼

Wsi  Wso Wsi  Wei

(38)

3H

¼

hsi  hso hsi  hei

(39)

The enthalpy energy transfer effectiveness as a function of 3 S, 3 L, and H* is calculated by [11,14]:

D (m2/s)

X X

X

(34)

d (mm)

X X

X

where W is the humidity ratio.The enthalpy heat transfer effectiveness is:

3H

X X

X

where the subscripts e, i, s, o refer to the exhaust, inlet, supply and outlet, m is the airflow rate, T is the temperature, and (m)min is equal to either ms or me, whichever is smaller.The latent heat transfer effectiveness is:

3L

RH (%)

X X X X

X

is performed to calculate average temperatures and concentrations of the supply air and exhaust air at the outlet of the HRV/ERV, which are used for the sensible and latent effectiveness calculations. The effectiveness is calculated as follows [10,16]:The sensible heat transfer effectiveness is:

3S

Ts ( C)

3H

¼

H* ¼

þ 3 L H* 1 þ H*

3S

DW 2501ðWsi  Wei Þ z2501 DT cpa ðTsi  Tei Þ

(40)

(41)

As can be seen in Eq. (41), H* is basically a ratio of latent to sensible energy differences between the inlets of two airstreams flowing through the ERV exchanger. It can in theory vary from N to þN, but varies typically from 6 to þ6 for total recovery in HVAC applications. Eq. (40) shows that the total energy transfer effectiveness is not a simple algebraic average of sensible and latent effectivenesses [11].

4. Results and discussion In this Section, the performance results of the numerical simulations for the HRV and ERV are presented in some details for the Canadian summer and winter season conditions shown in Tables 1 and 2, followed by the discussion of results.

4.1. HRV results 4.1.1. Effect of air velocity Fig. 3 illustrates effect of airstream flow velocity (Vs ¼ Ve) on HRV performance and fan power consumption in the summer and winter seasons for two flow configurations (co-current and counter-current). The results show that the effectiveness reduced by 34.2% and 20.0% for co-current and counter-current flows, respectively in the winter season due to increase the flow velocity from 0.5 to 2.5 m/s as the flow rate increases. In addition, at low flow velocity (V < 1 m/s), the season has no significant effect on the effectiveness. Whereas, the effectiveness decreased by 6.2% and 7.5% at flow velocity of 2.5 m/s for co-current and countercurrent, respectively due to the seasonal change from winter to summer. Furthermore, the effectiveness decreases gradually as the flow velocity further increases in both winter and summer seasons. On the other hand, the fan power consumption for both supply and exhaust streams is calculated using Eq. (42); where, DP is the pressure drop across the modelling domain (250 mm) for a single channel (n ¼ 1) of the supply and exhaust passes.

W. Yaïci et al. / Applied Thermal Engineering 51 (2013) 770e780 90%

70%

a

a

Effectiveness_cocurrent_summer Effectiveness_countercurrent_summer

80%

0.10

Summer 0.09

65%

Effectiveness_cocurrent_winter Effectiveness_countercurrent_winter

0.08

70%

0.07

ε

ε

60%

60%

55%

50%

50%

40%

0.06

45%

Effectiveness_RH = 40% Effectiveness_RH = 70%

0.05

Effectiveness_RH = 90% Fan power_RH = 40%

0.04

Fan power_RH = 70%

0.03

Fan power_RH = 90% 30% 1.0

1.5

2.0

40%

2.5

0.02 26

28

Air velocity (m/s) 0.12

b

34

36

38

70%

b

Fan power_winter

0.10

Winter 0.09

65%

0.08

0.08

60%

0.07

ε

Fan power (Watt)

32

Supply air inlet temperature ( C)

Fan power_summer

0.10

30

0.06

55%

0.06 Effectiveness_RH = 40%

0.04

0.05

Effectiveness_RH = 70%

50%

Effectiveness_RH = 90%

0.04

Fan power_RH = 40%

0.02

45%

Fan power_RH = 70%

0.03

Fan power_RH = 90%

0.00 0.0

0.5

1.0

1.5

2.0

2.5

3.0

Air velocity (m/s)





nV DP

 s

   þ nV DP

40%

0.02 -10

-8

-6

-4

-2

0

2

4

6

Supply air inlet temperature ( C)

Fig. 3. Variations of effectiveness and fan power consumption with air velocity for different seasons and flow directions of HRV.

E ¼

Fan power (Watt)

0.5

Fan power (Watt)

776

(42)

e

The results show that the fan power consumption increases appreciably as the flow velocity further increases and it increases about 32 times in the winter season due to increase the flow velocity from 0.5 to 2.5 m/s, as shown in Fig. 3. The flow direction has no significant effect on the fan power consumption. In addition, at low flow velocity (V < 1 m/s), the season has no significant effect on the fan power consumption. However, the fan consumed 18.6% higher power in the summer compared to the winter season at flow velocity of 2.5 m/s. 4.1.2. Effects of outdoor air temperature and relative humidity Fig. 4a and b demonstrate performance of HRV system for counter flow direction in winter and summer seasons representative of four typical Canadian cities (Calgary, Halifax, Ottawa and Vancouver) at different supply temperatures and relative humidities (Table 3). The membrane spacing and thickness have value of 2 and 0.1 mm, respectively and 250 mm is used for channel length. The supply and exhaust stream velocities have value of 2 m/s. HRV performance is presented in term of effectiveness and fan power consumption based on a single simulated channel for each supply and exhaust passes. The results show that there is no significant effect of increase in the supply temperature and relative humidity on the HRV effectiveness and the fan power consumption in the winter and summer seasons. On the other hand, in the winter weather, the effectiveness and fan power consumption increase by 7% and 22%, respectively compared to summer season.

Fig. 4. Variations of effectiveness and fan power consumption with supply air inlet temperature for different relative humidities on counter flow of HRV: (a) Summer; (b) winter.

4.2. ERV results 4.2.1. Effect of air velocity Fig. 5a and b present comparisons between the counter and cocurrent flow arrangements of the variations of sensible and latent energy transfer effectivenesses when the ERV is operated with equal supply and exhaust air velocities in the exchanger for the summer and winter conditions previously described. The air velocity in the exchanger is set at 0.5, 1.0, 1.5, 2.0 and 2.5 m/s. The dimension of membrane is set as follows: thickness at d ¼ 0.1 mm and the channel height (membrane spacing) at d ¼ 2 mm. The results show that the effectiveness decreases with the increase of air velocity, which is attributed to the air residence time within the ERV core. When the air velocity is low, the residence time is high, allowing a higher amount of heat and moisture transferred per kg of airflow; the larger the residence time, the higher is the effectiveness. The results also show that the latent effectiveness is lower than the sensible effectiveness, which is due to the additional resistance of the membrane. In addition, the counter flow configuration provides a higher effectiveness. This is because the temperature, concentration and specific enthalpy gradients are higher for the counter flow. For the counter flow configuration and summer season, the sensible and latent effectivenesses decrease from 78.4% to 47.4%, and 77.4% to 44.9%, when the air velocity increases from 0.5 to 2.5 m/s, respectively. For the winter season, the corresponding values are 71.9e43.7%, and 71.3e38.5%, respectively. On the other hand, for the co-current flow arrangement and summer season,

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Table 3 Ambient air conditions for the outdoor temperature and humidity study. Summer case TDB ( C) RH (%) AH (g/kg) Winter case TDB ( C) RH (%) AH (g/kg)

26 40 8.40

26 70 14.85

26 90 19.22

29 40 10.04

29 70 17.18

29 90 23.05

32 40 11.95

32 70 21.23

32 90 27.56

35 40 14.19

35 70 25.27

35 90 32.87

10 40 0.64

10 70 1.12

10 90 1.45

5 40 0.99

5 70 1.74

5 90 2.24

0 40 1.51

0 70 2.65

0 90 3.41

5 40 2.16

5 70 3.78

5 90 4.87

the sensible and latent effectivenesses decrease from 50.0% to 37.4%, and 50.0% to 37.1%, respectively. For the winter season, the corresponding values are 46.9e33.9%, and 47.3e33.7%, respectively. The effectiveness of the co-current flow is in average 38% and 21% lower than for the counter flow and air velocity in the range 0.5e2.5 m/s. Further, for this typical Canadian case studied, the ERV is more effective during the summer season. A comparison of the effectiveness values using Eqs. (34) and (35) with those obtained using Eqs. (37) and (38) shows the maximum deviations for all the effectiveness predictions are below 0.5% and 5% for the summer and winter cases, respectively. The pressure profiles for the co-current and counter flows, summer and winter seasons are obtained for all the cases studied. The pressure drop through the exchanger increases with the flow velocity; a higher flow resistance results to a higher pressure drop. For the summer season and for a single channel, when the airflow velocity increases from 0.5 to 2.5 m/s, the maximum pressure in the

a

80%

38 40 16.79

38 70 29.99

38 90 39.10

ERV increases from 1.8 to 11.2 Pa. For the winter season, the corresponding values are 1.5 and 9.5 Pa. 4.2.2. Effect of exhaust air velocity variation Fig. 6a and b present the comparison between the counter and co-current flow arrangements of the variations of sensible and latent effectivenesses for a constant supply air velocity of 2 m/s and when the exhaust air velocities in the ERV exchanger are set at 0.5, 1.0, 1.5, 2.0 and 2.5 m/s for the summer and winter conditions, respectively. The same trend as that reported in previous analysis is observed. However, the ERV is more effective when it is operated with exhaust airflow lower than the supply airflow. For example, at supply air velocity of 2 m/s, the latent effectiveness for the counter flow configuration increased in average by 25% and 43% and by 13% and 20% at exhaust air velocity of 0.5 and 1.5 m/s for the summer and winter seasons, respectively. This is again due to a higher

100%

Summer

Sensible_countercurrent

Summer

70%

Sensible_countercurrent

90%

Sensible_cocurrent Latent_countercurrent

Sensible_cocurrent Latent_countercurrent

80%

Latent_cocurrent

Latent_cocurrent

70%

ε

ε

60%

60% 50%

50% 40%

40%

a

30%

30% 0.5

1.0

1.5

2.0

0.5

2.5

1.0

Air velocity (m/s) 80%

b

Winter

70%

1.5

2.0

2.5

Exhaust air velocity (m/s) 100%

Sensible_countercurrent Sensible_cocurrent Latent_countercurrent Latent_cocurrent

Winter

Sensible_countercurrent Sensible_cocurrent Latent_countercurrent Latent_cocurrent

90% 80%

60%

ε

ε

70% 60%

50%

50% 40%

40%

b

30%

30% 0.5

1.0

1.5

2.0

2.5

Air velocity (m/s) Fig. 5. Variations of sensible and latent effectivenesses with air velocity (Ve ¼ Vs) for counter and co-current flows of ERV: (a) Summer; (b) winter.

0.5

1.0

1.5

2.0

2.5

Exhaust air velocity (m/s) Fig. 6. Variations of sensible and latent effectivenesses with exhaust air velocity for Vs ¼ 2 m/s, counter and co-current flows of ERV: (a) Summer; (b) winter.

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W. Yaïci et al. / Applied Thermal Engineering 51 (2013) 770e780

temperature and concentration gradients resulting in a higher heat and mass transfer rates or higher effectiveness. 4.2.3. Effects of outdoor air temperature and relative humidity As for the HRV, effects of the outdoor air temperature and relative humidity on the membrane-based ERV performance on counter flow arrangement for both typical Canadian summer and winter seasons are investigated. Fig. 7a and b show the variations of sensible and latent effectivenesses of ERV system at different supply temperature and relative humidity for summer and winter seasons, respectively. As can be seen, results are similar to that of the ERV exchanger; the outdoor air temperature and relative humidity have practically no effect on sensible effectiveness. The latent effectiveness increases slightly as the supply air inlet relative humidity increases. Furthermore, in the winter case, the effectiveness decreases initially with the increase of supply temperature, and then increases slowly as for the summer season conditions, with the minimum occurring at around 5  C, for the case considered. This is due to the fact that the driving force through the membrane increases with the increase of relative humidity in the supply air. 4.2.4. Effect of membrane spacing The effect of the membrane spacing (exchanger channel height) on the ERV performance is performed as for the previous cases for an equal supply and exhaust air velocities of 2 m/s, the channel height is varied between 1.58 and 2.90 mm. The number of channels can be calculated according to the following equation [16,25]:

l ¼ 2nðd þ dÞ 60%

a

(43)

Fig. 8a and b depict the comparison between the counter and cocurrent flow arrangements and for the summer and winter seasons of the variations of sensible and latent effectivenesses. The figures show that for a given supply air velocity, as the membrane spacing increases, the effectiveness decreases. The latent effectiveness in summer season decreases in the ranges of 58e37% and 45e31%, when the membrane spacing increases in the range 1.58e2.90 mm for the counter and co-current flows, respectively. For the winter case, when the membrane spacing increases, the latent effectiveness decreases gradually in the ranges of 59e38% and 47e29%, until a constant value is reached. This is because as the membrane spacing increases, the airflow rate increases but the total surface area decreases, resulting globally in a reduction of heat and mass transfer rates. When the membrane thickness is fixed, increasing the membrane spacing will decrease the number of channels, and therefore the contact area and NTU. As a result, the smaller the membrane spacing, the more effective is the membrane-based ERV. 4.2.5. Effect of membrane thickness The effect of the membrane thickness on the ERV performance is investigated as for the membrane spacing case. The membrane thickness is varied between 0.05 and 0.21 mm for which corresponds the number of channels that can be calculated according to the Eq. (43). Fig. 9a and b illustrate the comparison between the counter and co-current flow arrangements and for the summer and winter seasons of the variations of sensible and latent effectivenesses. As expected, the sensible effectiveness is higher than the latent effectiveness, because the heat transfer resistance is usually smaller

70%

Sensible_RH = 40% Sensible_RH = 70% Sensible_RH = 90% Latent_RH = 40% Latent_RH = 70% Latent_RH = 90%

Summer

55%

a

Summer

Sensible_countercurrent Sensible_cocurrent Latent_countercurrent Latent_cocurrent

60%

ε

ε

50%

40% 50%

30%

20%

45% 26

28

30

32

34

36

1.5

38

o

Supply air inlet temperature ( C) 60%

b

Winter

70%

Sensible_RH = 40% Sensible_RH = 70% Sensible_RH = 90% Latent_RH = 40% Latent_RH = 70% Latent_RH = 90%

ε

55%

2.0

2.5

3.0

Membrane spacing (mm)

b

Winter

Sensible_countercurrent Sensible_cocurrent Latent_countercurrent Latent_cocurrent

60%

50%

ε

50%

40% 45%

30%

20%

40% -10

-8

-6

-4

-2

0

2

4

6

Supply air inlet temperature (oC) Fig. 7. Variations of sensible and latent effectivenesses with supply air inlet temperature for different relative humidities on counter flow of ERV: (a) Summer; (b) winter.

1.5

2.0

2.5

3.0

Membrane spacing (mm) Fig. 8. Variations of sensible and latent effectivenesses with membrane spacing for Ve ¼ Vs ¼ 2 m/s, counter and co-current flows of ERV: (a) Summer; (b) winter.

W. Yaïci et al. / Applied Thermal Engineering 51 (2013) 770e780

4.2.6. Effect of membrane diffusivity The effect of the diffusivity of water in the membrane on the ERV performance is analysed as for the membrane spacing and thickness cases. The diffusivity is varied between 2.5  109 and 8  106 m2/s. Fig. 10a and b show the comparison between the counter and cocurrent flow arrangements and the variation of latent effectivenesses 70%

a

Summer

50%

a

Summer

Latent_countercurrent Latent_cocurrent

40%

ε

30%

20%

10%

0% 1E-09

1E-08

1E-07

1E-06

1E-05 2

Diffusivity of water in membrane (m /s) 50%

b

40%

Winter

Latent_countercurrent Latent_cocurrent

30%

ε

than the mass transfer resistance though the membrane. The figure shows that for a given supply air velocity, as the membrane thickness increases, the latent effectiveness decreases. The sensible effectiveness follows the same trend as for the latent effectiveness, but the decrease is very slow with the increase of membrane thickness, especially for the co-current flow. For the case studied, for the summer season, on co-current flow configuration, the sensible effectiveness is not influenced by the membrane thickness. For the winter case, the sensible effectiveness decreases slightly with the membrane thickness in the range 0.05e0.10 mm and then becomes not sensitive to the membrane thickness. The latent effectiveness in summer season decreases in the ranges of 51e45% and 41e38%, when the membrane thickness increases in the range 0.05e0.21 mm for the counter and cocurrent flows, respectively. Similarly, for the winter case, the latent effectiveness decreases slowly in the ranges of 46e41% and 38e32%, respectively. This is because the thermal and mass resistances increase when the membrane thickness increases, leading then to a reduced sensible and latent heat transfer rates. When the membrane spacing is fixed, increasing the membrane thickness will decrease the number of channels, and therefore the contact area and NTU. As a result, the thinner the membrane, the more effective is the membrane-based ERV.

779

20%

10%

0% 1E-09

1E-08

1E-07

1E-06

1E-05 2

Diffusivity of water in membrane (m /s) Fig. 10. Variation of latent effectiveness with diffusivity of water in membrane for Ve ¼ Vs ¼ 2 m/s, counter and co-current flows of ERV: (a) Summer; (b) winter.

Sensible_countercurrent Sensible_cocurrent Latent_countercurrent

60%

ε

Latent_cocurrent

50%

40%

30% 0.00

0.05

0.10

0.15

0.20

0.25

Membrane thickness (mm) 70%

b

Sensible_countercurrent

Winter

Sensible_cocurrent

5. Conclusions

Latent_countercurrent

60%

ε

Latent_cocurrent

50%

40%

30% 0.00

for the summer and winter seasons. It can be seen that for a given supply air velocity, as the diffusivity of water in the membrane increases, the latent effectiveness increases rapidly until a value of 106 m2/s, and then increases slowly until 8  106 m2/s. The latent effectiveness in summer season increases in the ranges of 0.2e49% and 0.2e30%, when the diffusivity increases between 2.5  109 and 8  106 m2/s for the counter and co-current flows, respectively. On the other hand, for the winter season, the latent effectiveness increases in the ranges of 0.2e45% and 0.2e38%, respectively. Moreover, the sensible and total effectivenesses of the ERV are slightly affected by the variation of the diffusivity. Therefore, for a membrane-based ERV with a given airflow rate and based on the above results, a good performance can be obtained by using a thin membrane with a high diffusivity and a suitable channel height.

0.05

0.10

0.15

0.20

0.25

Membrane thickness (mm) Fig. 9. Variations of sensible and latent effectivenesses with membrane thickness for Ve ¼ Vs ¼ 2 m/s, counter and co-current flows of ERV: (a) Summer; (b) winter.

A numerical study was performed to assess the effects of the supply and exhaust air velocity variations, the membrane spacing and thickness, the water diffusivity through the membrane, the outdoor air temperature and humidity on the performance of an HRV and a membrane-based ERV used in HVAC systems. A CFD model was developed and validated with data from the literature for this purpose. Performance predictions were then obtained in terms of sensible and latent effectivenesses. Critical exchanger parameters that can have a significant impact on the exchanger performance, such as airstream velocity, temperature, pressure, moisture concentration, and enthalpy were simulated. The results of the present investigation confirm those of previous research concerning the strong influence of design

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W. Yaïci et al. / Applied Thermal Engineering 51 (2013) 770e780

arrangement and operating conditions on HRV/ERV performance in terms of higher effectiveness of counter-flow arrangements. They also show a decrease in the ERV effectiveness with further increase in the supply and the exhaust airflow velocities. For a given air velocity, the effectiveness decreases noticeably with increase in membrane spacing. However, the effectiveness decreases slowly with increase in membrane thickness. The ERV effectiveness in the summer is higher in comparison to the winter under Canadian climatic conditions. The latent effectiveness increases significantly with the increase in the water diffusivity in the membrane up to a value of 106 m2/s. For the HRV, the results show that at equal supply and exhaust velocity, increasing flow velocity 5 times from 0.5 m/s reduced the HRV effectiveness about 34% and 20% for co-current and counter flow configurations, respectively. In addition, the season has a significant effect on the HRV performance and fan power consumption at an airstream velocity higher than 1 m/s. It is found for the case investigated that the total fan power consumption is about 18.6% higher in the summer at high flow velocity (V ¼ 2.5 m/ s) in comparison to winter. Moreover, the HRV and membrane-based ERV results show that the outdoor air temperature and relative humidity have no significant effect on sensible effectiveness and fan power consumption in both winter and summer. However, for the membrane-based ERV, the latent effectiveness is slighted affected by the outdoor conditions. The effectiveness increases to some extent with increasing relative humidity in both summer and winter cases investigated, but very slightly with the increase of outdoor air temperature. Further work will include effects of membrane configuration and material properties on the effectiveness, optimisation of the HRV/ERV performances, as well as validation of the numerical predictions with experimental data from real HRV and membranebased ERV systems. Through these sample results, CFD has proven to be a practical and an effective design tool that can help HVAC engineers and manufacturers design HRV/ERV systems under different operating conditions, flow configurations, climates and seasons, and membrane material properties. Qualitative and quantitative analyses on the operation and performance can be performed for optimal design operation. Detailed behaviours and average features of all the relevant parameters can be easily predicted and the effects of individual parameters on the HRV or ERV system performance can be analysed. Acknowledgements The authors would like to express their appreciation and to acknowledge the financial support received for this work from the Canadian Federal Government’s Programme on Energy Research and Development (PERD). References [1] L.Z. Zhang, Total Heat Recovery, Heat and Moisture Recovery from Ventilation Air, Nova Science Publishers, Inc., New York, 2008.

[2] P. Talukdar, S.O. Olutmayin, O.F. Osanyintola, C.J. Simonson, An experimental data set for benchmarking 1-D, transient heat and moisture transfer models of hygroscopic building materials. Part I. Experimental and material property data, International Journal of Heat and Mass Transfer 53 (2007) 4527e4539. [3] L.Z. Zhang, J.L. Niu, Energy requirements for conditioning fresh air and the long-term savings with membrane-based energy recovery ventilator in Hong Kong, Energy 26 (2) (2001) 119e135. [4] Natural Resources Canada, Office of Energy Efficiency, Energy Use Data Handbook Tables. Canada (2008), http://oee.nrcan.gc.ca/corporate/statistics/ neud/dpa/handbook_tables.cfm. [5] L.Z. Zhang, Progress on heat and moisture recovery with membranes: from fundamentals to engineering applications, Energy Conversion and Management 63 (2012) (2012) 173e195. [6] L.Z. Zhang, Y. Jiang, Heat and mass transfer in a membrane-based enthalpy recovery ventilator, Journal of Membrane Science 163 (1999) 29e38. [7] Y.P. Zhang, Y. Jiang, L.Z. Zhang, Y.C. Deng, Z.F. Jin, Analysis of thermal performance and energy savings of membrane based heat recovery ventilator, Energy 2 (25) (2000) 515e527. [8] ASHRAE, ASHRAE Handbook e HVAC Systems and Equipment, American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc., Atlanta, 2000 (Chapter 44.7). [9] A. Mardiana-Idayu, S.B. Riffat, Review on heat recovery technologies for building applications, Renewable and Sustainable Energy Reviews 16 (2012) 1241e1255. [10] M. Nassif, R. Al-Waked, G. Morrison, M. Behnia, Membrane heat exchanger in HVAC energy recovery systems, systems energy analysis, Energy and Buildings 42 (2010) 1833e1840. [11] C.J. Simonson, R.W. Besant, Energy wheel effectiveness: part I - development of dimensionless groups, International Journal of Heat and Mass Transfer 42 (1999) 2161e2170. [12] C.J. Simonson, R.W. Besant, Energy wheel effectiveness: part II e correlations, International Journal of Heat and Mass Transfer 42 (1999) 2171e2185. [13] J.L. Niu, L.Z. Zhang, Membrane-based enthalpy exchanger: material considerations and clarification of moisture resistance, Journal of Membrane Science 189 (2001) 179e191. [14] L.Z. Zhang, J.L. Niu, Effectiveness correlations for heat and moisture transfer processes in an enthalpy exchanger with membrane cores, ASME Journal of Heat Transfer 124 (2002) 922e929. [15] L.Z. Zhang, C. Liang, L. Pei, Heat and moisture transfer in application scale parallel-plates enthalpy exchangers with novel membrane materials, Journal of Membrane Science 325 (2008) 672e682. [16] J. Min, M. Su, Performance analysis of a membrane-based energy recovery ventilator: effects of membrane spacing and thickness on the ventilator performance, Applied Thermal Engineering 30 (2010) 991e997. [17] J. Min, M. Su, Performance analysis of a membrane-based energy recovery ventilator: effects of outdoor air state, Applied Thermal Engineering 31 (2011) 4036e4043. [18] L.Z. Zhang, Energy performance of independent air deshumidification systems with energy recovery measures, Energy 31 (2006) 1228e1242. [19] L.Z. Zhang, D.S. Zhu, X.H. Deng, B. Hua, Thermodynamic modelling of a novel air deshumidification system, Energy and Buildings 37 (3) (2005) 279e286. [20] L.Z. Zhang, Convective mass transport in cross-corrugated membrane exchanger, Journal of Membrane Science 260 (2005) 75e83. [21] L.Z. Zhang, F. Xiao, Simulation of heat and moisture transfer trough a composite supported liquid membrane, International Journal of Heat and Mass Transfer 51 (2008) 2179e2189. [22] L.Z. Zhang, Simulation of heat and mass transfer in plate-fin enthalpy exchangers with different plate and fin materials, International Journal of Heat and Mass Transfer 52 (2009) 2704e2713. [23] L.Z. Zhang, Heat and mass transfer in a quasi-counter flow membrane-based total heat exchanger, International Journal of Heat and Mass Transfer 53 (2010) 5478e5486. [24] COMSOL, Multiphysics and CFD Modules, Version 4.2: User’s Manuals, COMSOL, Inc., May 2011. [25] F.P. Incropera, D.P. DeWitt, Fundamentals of Heat and Mass Transfer, fourth ed., John Wiley & Sons, New York, 1996. [26] W.M. Kays, M.E. Crawford, Convective Heat and Mass Transfer, fourth ed., McGraw-Hill, New York, 2005. [27] M. Nassif, G.L. Morrison, M. Behnia, Membrane based enthalpy heat exchanger performance in HVAC system, Journal of Applied Membrane Science and Technology 2 (2005) 31e46.