Anticipating an efficient relative humidity in a room under direct solar radiation and equipped by radiant cooling panel system

Accepted Manuscript

Anticipating an efficient relative humidity in a room under direct solar radiation and equipped by radiant cooling panel system Nima Najafi Ziarani , Amin Haghighi Poshtiri PII: DOI: Reference:

S0140-7007(18)30406-7 https://doi.org/10.1016/j.ijrefrig.2018.10.018 JIJR 4143

To appear in:

International Journal of Refrigeration

Received date: Revised date: Accepted date:

23 May 2018 13 October 2018 21 October 2018

Please cite this article as: Nima Najafi Ziarani , Amin Haghighi Poshtiri , Anticipating an efficient relative humidity in a room under direct solar radiation and equipped by radiant cooling panel system, International Journal of Refrigeration (2018), doi: https://doi.org/10.1016/j.ijrefrig.2018.10.018

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ACCEPTED MANUSCRIPT Highlights  In the room with western external wall ceiling temperature should be 9°C to 10°C higher than the room with south wall. 

By using coated glass maximum ceiling temperature could be raised up to 6°C.



Undesirable temperature rises of the room surfaces due to solar radiation could be damped by using coated glass. The maximum limit of relative humidity in the room is almost 37% by using clear glass and

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

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55% to 61% by using coated glass.

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ACCEPTED MANUSCRIPT

Anticipating an efficient relative humidity in a room under direct solar radiation and equipped by radiant cooling panel system

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Nima Najafi Ziarani, Amin Haghighi Poshtiri *

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* Department of Mechanical Engineering, University of Guilan, P.O. Box 3756, Rasht, Iran

* Corresponding author: Amin Haghighi Poshtiri Assistant Professor

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Mech. Eng. Department University of Guilan

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Tel: +98-13-33690274-6058, Fax: +98-13-33691084, Mob: +98-9125857902

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Email: [email protected]

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ACCEPTED MANUSCRIPT Abstract In the present study the performance of Radiant Cooling Ceiling System in a room under direct solar radiation is investigated by combination of numerical and analytical solutions. Assessments are implemented in four distinct climates in Iran which include Tabriz (dry and cold), Tehran (mild and dry), Rasht (mild and humid) and Yazd (dry and hot) in the hottest

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hour of the hottest day of the year 2017. Efficient surface temperatures of the ceiling are determined to supply thermal comfort conditions. Turbulent flow, radiation, solar radiation are modeled by Low-Reynolds k-ε model, discrete ordinate model and ray-tracing model, respectively, by FLUENT 6.3 software. To avoid condensation, critical RH method is defined

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to determine suitable RH accurately. Thermal comfort is investigated by PMV index. The energy saving effect of coated glass is also illustrated. Results demonstrate that by using coated glass, energy usage can be reduced 49 to 58 percent. Furthermore, changing room‟s orientation from west to south can boost the energy saving. Accordingly, the higher ceiling

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temperature which provides thermal comfort for occupants means the lower condensation probability; hence, radiant cooling ceiling systems could be utilized even in the mild and

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humid climate without dehumidifying necessity. Keywords: Radiant cooling, Cooling ceiling, Solar irradiation, Thermal comfort,

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Condensation, Relative humidity

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ACCEPTED MANUSCRIPT 1. Introduction Nowadays buildings are the most energy consumer portion in comparison with other economic parts; 30-45% of the global energy usage is consumed in buildings (Pout, MacKenzie, & Bettle, 2002),(Asimakopoulos et al., 2012). There are different methods for cooling a building, one of which is convective systems., However, there is another type of

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cooling systems which have radiant interaction with room‟s surfaces and also occupants in addition to convective interaction; Radiant systems are gaining more popularity because of high thermal comfort, reduce energy consumption, quiet operation, space saving and etc (Rhee & Kim, 2015).

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Radiant cooling panels are in fact panels with controllable surface temperature embedded in the ceiling, floor or walls. The specific temperature of the surfaces could be determined by certain temperature of water flow through the embedded pipes attached on the panel. The surface which its temperature is governed, would be dubbed “Radiant Panel” if the portion of

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the radiation heat transfer with other surfaces in the room is 50 percent or more (ASHRAE, 2000).

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A significant goal in the HVAC systems is providing thermal comfort conditions for occupants. Should the series of conditions do tolerable for at least 80% of people, thermal

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comfort is supplied; in another word, whenever the human‟s body does not feel hot and cold, it is satisfied by thermal conditions, consequently it is in thermal comfort conditions (CIBSE,

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2006). Catalina et al. (Catalina, Virgone, & Kuznik, 2009) assessed thermal comfort in a

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room equipped by cooling ceiling. Results demonstrate that the temperature gradient on the vertical line at the room center is lower than 1°C. Tye-Gingras and Gosselin (Tye-Gingras & Gosselin, 2012) carried out an investigation of thermal comfort and energy consumption in ceiling and wall hydronic radiant heating systems. Results indicate that the inlet water temperature is the more important parameter than the way of arraying and locating the panels. Gao et al. (Gao et al., 2017) investigated the specifications of the radiant cooling by means of studying the heat exchange between human body and the room surfaces. It is

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ACCEPTED MANUSCRIPT suggested that instead of appraising complex environmental factors, human thermal balance can be used so as to evaluate thermal comfort. Both thermal comfort parameters and system‟s energy consumption should be investigated by determining heat transfer coefficient of room‟s surfaces. Karadag (Karadag, 2009) provided the correlation between radiation and convection heat transfer coefficients on the

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radiant cooling ceiling panels. It is demonstrated that the ratio of radiation heat transfer to convection heat transfer is in the range of 0.7 to 2.2 rely on the difference between air and ceiling surface temperature. Cholewa et al. (Cholewa, Anasiewicz, Siuta-Olcha, &

Skwarczynski, 2017) assessed heat transfer coefficients for a heated/cooled radiant ceiling.

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They concluded that heat transfer coefficients on heated radiant ceiling were overestimated. They also found out that the higher ceiling surface emissivity was, the higher radiant heat flux would be resulted..

The condensation phenomenon is one critical aspect of using cooled surface in a room. Yen

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et al. (Yin, Wang, Zhai, & Ishugah, 2014) investigated weak performance in heat transfer and condensation phenomenon in radiant cooling panels. Results show that the heat

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transfer rate could be enhanced by increasing water velocity through the pipes up to a particular range; and heat transfer rate would be higher due to condensation. Ning et al.

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(Ning, Chen, Liu, & Zhang, 2016) proposed cooling radiant ceiling panel with thin air layer by which the condensation phenomenon in hot and humid climate could be controlled. The

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resulted increment in cooling capacity is 43-46% compared to common cooling radiant

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ceiling systems.

Radiant cooling systems can also be utilized to improve all-air systems. Karmann et al. (Karmann, Schiavon, & Bauman, 2017) reviewed several literatures so as to compare the application of radiant systems and all-air systems; it is claimed that radiant systems could provide equal and even better thermal conditions than all-air systems. Corgnati et al. (Corgnati, Perino, Fracastoro, & Nielsen, 2009) investigated draft risk in the all-air systems with outlet air near the ceiling and radiant cooling ceiling systems. More longitudinal throw and less vertical drop of outlet air flow and also decline in draft risk are reported. Khan et al. 5

ACCEPTED MANUSCRIPT (Khan, Khare, Mathur, & Bhandari, 2015) investigated the performance of combination of radiant cooling system (both ceiling and floor) and fan coil system compared with all-air system. Uniform vertical temperature gradient and lower energy consumption is resulted by radiant cooling system. The effect of direct sun radiation can cause huge changes in prediction of cooling load.

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Olesen (Olesen, 1997) assessed cooling capacity of the radiant cooling floor and it is resulted that the maximum cooling capacity for this system is about 50 w/m2; whenever the space is imposed by direct solar radiation, the maximum cooling capacity could be higher, around 100 to 150 W/m2. Causone et al (Causone, Corgnati, Filippi, & Olesen, 2010)

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calculated the cooling load in radiant systems. Considering condensation phenomenon, it is resulted that in the air temperature of 28°C and relative humidity of 58 percent, the maximum cooling load which can be supplied by radiant cooling panel is 117 w/m2. Feng et al.(Feng, Schiavon, & Bauman, 2016) provided new method for the design of radiant floor cooling

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system.. Results indicate that the actual cooling capacities are 1.2 times higher than the values computed with the ASHREA method, and 1.44 times higher than the ISO 11855.

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Energy consumption could be one of the most important parameters in design, if not the most. Li et al. (Li, Yoshidomi, Ooka, & Olesen, 2015) provided experimental investigation of

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the performance of radiant heating/cooling ceiling systems. According to the results, the maximum required thermal loads are measured 50.8 W/m2 and 35.4 W/m2 for heating and

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cooling respectively. Results moreover showed that the output heat flux from the upside of

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the panel devote 30 to 40 percent of total output heat flux to itself in both cases. Su et al. (Su, Li, Zhang, Sun, & Qian, 2015) investigated cooling and heat transfer characteristics of the concrete ceiling radiant cooling panels. The results indicate that the rate of declining the surface temperature of the concrete panels at the startup of the system is less than metal panels due to the higher thermal inertia of the concrete panels. The average cooling capacity of this system will be in the range 40 W/m2 to 50 W/m2. Despite all of various assessments of radiant cooling systems, the effect of direct solar radiation on these systems by providing CFD solution to attain suitable relative humidity so 6

ACCEPTED MANUSCRIPT as to avoid condensation have not yet been investigated. The mentioned evaluations are grossly crucial, specifically in the administrative and commercial buildings which are included a large area of glass surfaces. The main objectives of the present study are calculating maximum surface temperature of the radiant cooling ceiling based on thermal comfort necessities numerically, and determining the corresponding critical relative humidity in the

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test room to avoid condensation on the cold surface of the ceiling analytically; Maximum cooling loads in four different direction of same room is obtained to find the critical condition and using coated glass is proposed to damp undesirable parameters like direct sun radiation to the room and make it possible that the temperature of the ceiling to be higher and the risk

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of condensation lower. The investigations are done in four cities of different climates of the Iran. Accordingly, the objectives of present study are as follows:

1- Three dimensional numerical modeling of a room including a glass window under direct solar irradiation and equipped by radiant cooling ceiling system.

thermal comfort conditions.

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2- Obtaining the sufficient surface temperature of the radiant panels based on providing

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3- Define a new method so as to consider the suitable relative humidity of the air. 4- Determining maximum allowable relative humidity in a room in which the thermal

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comfort conditions are established in different cases so as to avoid condensation. 5- Investigating the effect of using coated glass instead of clear glass on saving energy

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in the radiant cooling ceiling systems.

2. Schematic and Numerical modeling 2.1. Schematic According to Error! Reference source not found., a 3m×3m×3m cubical room including an external wall has been considered. There is a 1m×3m glass window at the height of 1.5m from the floor is set up on the single external wall. Actually, it is not an imaginary room; it is a room which is between two other rooms in a building. Although it could have been 7

ACCEPTED MANUSCRIPT considered as a room with three walls near to ventilated space and an out-wall, we considered the other three walls as adiabatic so as to investigate the direct effect of sun radiation and room direction on radiant cooling ceiling temperature and thermal comfort parameters; in this way other variables would not helpful. Consequently, the amount of heat will come through the out-wall in conduction and radiation ways of heat transfer and will be

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absorbed by radiant cooling panels in convection and radiation ways of heat transfer.

2.2. Numerical Solution

Based on comparisons between CFD model and experiments which is done by Mikeska et

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al. (Mikeska, Fan, & Svendsen, 2017); they show that the air temperature in the room can be predicted well by CFD model; on the other hand, it is said that there are some differences between CFD air speed calculation and measured air speed, which could consider as an ignorant in the evaluation of the indoor comfort category by CFD.

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To obtain velocity and temperature distribution in the room, CFD method is used. Numerical solution is carried out by FLUENT 6.3 software to solve continuity, momentum, energy and

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radiation equations for steady state, incompressible, three dimensional and turbulent flow. Whenever the air density varies in the room due to temperature differences, in the presence

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of gravity, buoyant forces could cause natural convection flow. The intensity of natural convection flow due to buoyant forces is obtainable by Rayleigh number.

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The present study is a natural convection problem that the Rayleigh number is approximately 1012; hence, the air flow is turbulent flow caused by buoyancy forces.

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Hashimoto (Y, 2005) demonstrates that Standard k-ε could be exploited in the fully turbulent flows, whereas it is not applicable in the flows forming by buoyant forces. In mentioned study, flow field in an enclosure is simulated by Low-Reynolds k-ε model with AKN damping function. Awbi (Awbi, 1998) illustrates that Low-Reynolds k-ε model gives more accurate results of convection heat transfer coefficient in natural convection flow on the room‟s surfaces in comparison with Standard k-ε.

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ACCEPTED MANUSCRIPT In present study modeling of the turbulent flow is done by Low-Reynolds k-ε model with AKN damping function (equations 1)(ABE, Kondoh, & Nagano, 1994): ⁄

̅

⁄

⁄

{(

⁄

)

⁄

}

⁄

̅

⁄

⁄

{(

⁄ )

⁄

}

( ̅⁄ (

̅⁄

)

(1-a)

⁄ ̅̅̅̅̅ ̅ ⁄

⁄

(1-b)

( ⁄ )

(1-c)

⁄ )

(1-d)

In the equations outlined above,

and

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̅̅̅̅̅

̅̅̅̅̅ ̅ ⁄

provided in Error! Reference source not found.

are damping functions of the model. Damping functions apply the near-wall effect and lowReynolds effects.

In Error! Reference source not found. the model‟s constants (

,

,

and

) are

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,

provided.

Discrete Ordinate radiation model is utilized so as to radiation modeling in the domain. The discrete ordinates method is based on a discrete representation of the directional variation of

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the radiative intensity. A solution to the transport problem is found by solving the equation of

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transfer for a set of discrete directions spanning the total solid angle range of 4 . As such, the discrete ordinates method is simply a finite differencing of the directional dependence of

̂

(

̂)

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the equation of transfer(F.Modest, 2013) ( ) ( )

( ) (

̂)

( )

∫

(

̂) ( ̂

̂)

(2)

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FLUENT software provide the capability of solar load modeling; So that, it could be exploited to simulate the effect of solar irradiation in any geographical location, month, day and time of

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the day; in this way, in the present study Ray-Tracing model is used. This model adds the cooling load associated with direct solar irradiation to energy equation as a source term. Thus, the cooling load caused by direct solar irradiation is applied as boundary condition, and its impact is added into energy equation. The emissivity coefficients of the surfaces and the glass surface are determined 0.9 (materials)and 0.84 (LBNL) respectively. Control volume method and SIMPLE algorithm are used in the numerical solution. The discretization of Pressure-Velocity coupling and 9

ACCEPTED MANUSCRIPT Momentum and energy equations are done by Body Force Weighted (BFW) and Power Low respectively; other equations are discretized by Second Order Upwind.

2.3. Boundary Condition In a room illustrated in Error! Reference source not found., it is assumed that the whole

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ceiling area is covered by radiant cooling panels. Tye- Gingras and Gossilen (Tye-Gingras & Gosselin, 2012) demonstrated that temperature gradient on the radiant panel in a 2m length panel is approximately 0.7°C which is negligible; accordingly, in the present study the

boundary condition of the ceiling is considered as constant temperature. It is also assumed

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that inner walls and the floor are insulated; therefore the zero heat flux boundary condition is applied to the mentioned boundaries. Above and below the glass window is covered with 2m2K/W thermal resistance wall which is dubbed „out-wall‟.

In order to apply out-door temperature effect on out-wall and consequently on the room,

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correlation (Palyvos, 2008):

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convection heat transfer coefficient of outside of out-wall is determined by following

(3)

In the present study the wind velocity is considered 3m/s (which is the average wind velocity

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during the studying days); in order to avoid the deceptive effects of weather events on the results. In this regard, based on equation 3, convection heat transfer coefficient at the

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outside of out-wall is obtained 19.4 w/m2K.

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To apply solar irradiation on the glass window, the boundary condition is considered as semi-transparent object. The thermo-physical properties of the glass are provided in the Error! Reference source not found. (LBNL). In present study two kinds of glasses are simulated, clear glass and coated glass. Optical properties of these two are provided in Table 4 (LBNL).

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ACCEPTED MANUSCRIPT 3. Mesh Independency and Validation 3.1. Mesh Independency In order to evaluate mesh independency, outlet heat flux from the ceiling is calculated and compared for some different structured meshes in Table 5. Calculations are done for the same geometry (Figure 1) in Tehran at 16 O‟clock in the hottest day of summer. With

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respect to 0.318 percent difference in output heat flux between 50×50×50 mesh and that of 60×60×60 mesh and due to the necessity of declining the calculation time, the 50×50×50 mesh is chosen as efficient mesh. Errors in Error! Reference source not found. are

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relative errors between each mesh and its precedent.

3.2. Validation

Because of the inherent errors of numerical solutions, it is necessary to examine the accuracy of the adopted methods. The main purpose of validations in this study is finding out

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that the turbulent method which is used in numerical solution gives true results; because, all of these methods are formed based on experimental background and there is no guarantee

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to solve every turbulent flows truly. In the flows which are natural convection in an enclosure with high Rayleigh numbers the vortexes are mainly formed in the vicinity of the walls and

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these vortexes are the cause of turbulent flow in the room. Thus, the results of parameters like heat transfer coefficient near the walls have major significance. It should be noted that

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the grids form in the domain is very dense mesh near the wall because of the treatment of

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the flow near it in such problems. Additionally, the validity of the numerical solution is checked by temperature distribution validation through a line at the center of the room.

3. 2. 1. Validation of natural convection flow modeling To validate natural convection simulation, the data resulting from present study is compared with reference (Karadag, 2009). The mentioned simulation is done in a 3m×3m×3m room. Ceiling temperatures ranging from 0°C to 25°C with 5°C steps; Walls temperature ranging from 28°C to 36°C with 2°C steps; and the floor is insulated. This comparison is illustrated in 11

ACCEPTED MANUSCRIPT Error! Reference source not found.. The average error in all comparison points is less than 6 percent, which is revealed validation of the calculated convection heat transfer coefficients. The mentioned correlation is provided in equation 4. (

)

(4)

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3. 2. 2. Validation of radiation In order to validate Discrete Ordinate (DO) radiation model, data resulting from present study is compared with that of obtaining from the correlation in reference (Karadag, 2009). The mentioned simulation is done in a 3m×3m×3m room with different boundary conditions.

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Ceiling temperature ranging from 0°C to 25°C with 5°C steps; Walls temperature are

constant (36°C); and the floor is insulated. The results are illustrated in Error! Reference source not found..

The average error in all comparison points is less than 6 percent, which is revealed that the

(

)

(-5)

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⁄

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calculated ratio (equation5) is valid.

3. 2. 3. Validation of air temperature distribution

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The geometry investigated by reference (Xamán, Arce, Álvarez, & Chávez, 2008) is used to validate the room temperature distribution. This study evaluated natural convection in a

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cavity including a glass wall, numerically. A 3.24m×3.24m cavity including a glass wall, a constant temperature wall in front of the glass wall, with insulating ceiling and floor; 750

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W/m2 radiation flux is imposed on the glass wall; the temperature of the another wall is constant (21°C); the Thermo-physical properties of the glass wall and its optical properties are as same as the properties in Error! Reference source not found. and Error! Reference source not found., respectively. The result of this comparison along a horizontal line at the middle of the cavity is illustrated in Figure 4. The average error in all comparison points is less than 3.5 percent, which is revealed that the calculated ratio is valid.

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ACCEPTED MANUSCRIPT 4. Thermal Comfort Fanger thermal comfort model (Fanger, 1972) is used to assess the thermal comfort conditions in the room under the mentioned conditions. (equations 6). (

)

, (

-

)

( ,(

(6-a) )

,(

{

| |

√ √

|

|

)

) }

}

(6-b)

(6-c)

(6-d)

√

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{

( ̅

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|

|

) -

)

)

(

{

( ̅

)

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(

)

(

{

-

) )

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,

( ,( ( )

In the present study it is assumed that an unknown orientation occupant stands at the center are considered as 0.5 met and 1.2 clo,

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of the room, wears summer clothing; so M and

2007). ̅

̅

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respectively (ISO, 2005). Mean radiant temperature ̅ is calculated by equation 7 (ISO,

̅

̅

(7)

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View factors are calculated based on the algorithm which is illustrated in equations 8 providing by Reference (Cannistraro, Franzitta, Giaconia, & Rizzo, 1992). .

⁄

/

(

⁄

)

{

(8-a)

(8-b)

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ACCEPTED MANUSCRIPT It is assumed that the center point for the standing object is at the height of 1.1m from the floor. The correction method based on equation 9 is used so as to modify the error in view factors calculations. ⁄∑

(9) and ∑

In equation 9,

are modified view factors and summation of view factors

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before modification, respectively. The view factors which are calculated and modified by mentioned method are presented in Error! Reference source not found.. Thus, based on equation 7, mean radiant temperature is calculated.

Admittedly, because of the existence of a surface which would be much colder than the

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others in the room, by and large, condensation phenomena would be possible. Therefore the lower ceiling surface temperature is, the higher risk of condensation would be. Radically, low temperatures of the ceiling are inevitable due to thermal comfort requirements. Though, relative humidity is the only parameter that could be governed by dehumidifying the supply

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air, in present study, the numerical solution is not affected by relative humidity. Thereby, in order to avoid condensation and supply thermal comfort conditions, the highest possible

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level of relative humidity is determined. For this purpose a specific procedure is introduced so as to determine the sufficient relative humidity. The maximum allowable relative humidity

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which would prevent condensation is defined as „Critical Relative Humidity‟. After computing air temperature numerically for a certain ceiling surface temperature, the

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relative humidity corresponding to these mentioned parameters would be attained. The

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relative humidity is calculated based on August-Roche-Magnus approximation (Alduchov & Eskridge, 1996) which is presented in equation 10. (

(

⁄

)⁄

(

⁄

))

(10) The calculated relative humidity would be tapped into the code written based on equations 6, so as to compute PMV and PPD values. According to references (Wolkoff & Kjærgaard, 2007) and (Balaras, Dascalaki, & Gaglia, 2007) to meet comfort conditions relative humidity

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ACCEPTED MANUSCRIPT should be between 30 to 60 percent; any relative humidity out of the mentioned range could cause thermal discomfort even if thermal conditions are approved by PMV index. Hence, in present study, the critical relative humidity is calculated; if the resulting value is in the range of 30 to 60 percent, it will be tapped into the PMV code; if not, for the values lower than 30 percent and higher than 60 percent the considered values for relative humidity will be 30 and

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60 percent, respectively (it is assumed that in these cases there is a humidifier which can set the unsuitable relative humidity). Therefore, the air relative humidity has always to be in the range of 30 to 60 percent and lower than critical relative humidity at the same time to avoid condensation.

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As it is demonstrated in Error! Reference source not found., in order to characterize the efficient and also sufficient temperature of the radiant panels the following process must be accomplished; Firstly, the air temperature and velocity and room surfaces temperatures should be derived by simulating the air flow numerically for obtaining an arbitrary ceiling

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temperature (steps 1 and 2); Secondly, mean radiant temperature should be calculated by equation 7 using the resulted room surfaces temperatures from the former step (step 4); and

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based on the selected ceiling temperature and air volume average temperature, critical relative humidity should also be computed by means of August-Roche-Magnus

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approximation (steps 5 and 6); Thirdly, the required parameters have to tap into the PMV code which is written based on equations 5; and so the PMV and PPD values associated

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with arbitrary ceiling temperature should be derived (step 7). According to Fanger model, to

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attain comfort conditions the PMV values have to be in the range of -0.5 to 0.5; in present study this range is narrowed due to the capability of comparing the results accurately. Accordingly, a range of 0.3 to 0.5 is selected from the upper bound of PMV range. The mentioned range is chosen due to achieve to the maximum possible temperature of the ceiling. Therefore if the calculated PMV were in this range, the considered ceiling temperature would be efficient (step 9); otherwise, the ceiling temperature would be changed to a centigrade degree upper or lower and the process would be repeated (step 10).

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ACCEPTED MANUSCRIPT 5. Results and Discussion All of the investigations carried out in present study are done in order to determine the efficient radiant panels‟ temperature and the sufficient relative humidity. In this section, firstly, the ceiling surface temperatures supplying thermal comfort conditions for occupants are provided. Then, the effects of room orientation and using coated glass on the ceiling

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surface temperatures are determined. At last, thermal comfort parameters and critical relative humidity are assessed.

Evaluations are done in four different climates in Iran including Tabriz (dry and cold), Tehran (mild and dry), Rasht (mild and humid) and Yazd (dry and hot) in the hottest hour of the

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hottest day of the year. Outdoor air temperature and relative humidity are presented in Error! Reference source not found. (Weather-Underground).

Assessments are done by applying the geographical properties of each city in a particular day and time of the day in order to calculate an exact sun radiation and its cooling loads.

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Solar direction vectors and solar irradiation values are provided in Error! Reference source

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not found..

5.1. The effect of external wall orientation

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Results show that the suitable ceiling temperature which satisfies thermal comfort conditions in the room with western out-wall in Tehran, Tabriz and Yazd is 10°C and 11°C in Rasht.

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While, this figure in Tehran, Tabriz and Yazd is 19°C and 21°C for Rasht, when the room

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has southern out-wall. Clearly, the lower ceiling temperature is, the higher output heat flux from the ceiling and higher energy consumption will be resulted. For instance, the output heat flux in the southern out-wall room in Tabriz is 61.4

⁄

, however, it is 186 ⁄

in

the western out-wall room. The temperature distribution contours of these two rooms in Tabriz are demonstrated in Figure 6 and Figure 7. As it is obvious in these contours, the surfaces which have remarkably high temperatures in comparison to other surfaces of the room, are those one which are imposed by direct solar radiation. Clearly, this surface encompasses larger area in Error! Reference source not found. in comparison to Error! 16

ACCEPTED MANUSCRIPT Reference source not found.; this is a witness to the significance of the extent of the exposed surface by direct solar radiation in cooling load. Ultimately, it can be resulted that the room with a western external wall is the critical orientation for radiant cooling panels; consequently, investigations are done for this room.

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5.2. The effect of using coated glass Evaluating the significance of utilizing coated glass when the room is imposed by direct solar radiation is one of the assessments in present study.

In Table 9 the efficient ceiling temperatures in two cases of rooms with clear glass and

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coated glass are provided. Clearly, due to the unwanted effect of solar irradiation is damped by coated glass, suitable ceiling surface temperature in the room with coated glass is obtained about 6°C to 7°C higher than the room with clear glass.

The coated glass damps the unfavorable factor of solar radiation and this lead to decline the

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output heat flux from the ceiling about 58 to 49 percent, which can be considered as energy

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saving.

5.3. Thermal comfort parameters

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The average air temperature, velocity and mean radiant temperature used in PMV calculations are also provided in Error! Reference source not found.. It is clear that in two

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cases of using clear glass and coated glass, the air temperatures are approximately close to

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each other in the range of 25°C to 26°C; but, higher air volume average velocity and mean radiant temperature values are attained in the case which has lower ceiling temperature.Because mean radiant temperature is rely on the room surfaces temperature and the partial high temperature rise on the floor occurred in a room with clear glass (Error! Reference source not found.) is damped by coated glass (Error! Reference source not found.), it caused decrement in this parameter‟s value.

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ACCEPTED MANUSCRIPT Using coated glass, the average air velocity would be decreased and this could be assessed by natural convection in this room which is formed due to the variations in the room‟s air density. These temperature differences are mainly because of heat exchanging between the air which is flowing near the hot or cold surfaces; the ceiling which is colder than other surfaces and would be cause of downward air flow; being hotter than other surfaces, an

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external wall and the irradiated floor, upward flow would be formed. The velocity distribution contour and streamlines diagram on a central plane which is perpendicular to the external wall and Z direction is depicted in Error! Reference source not found.. It is clear that the external wall and also the irradiated area cause an upward flow; these two main flows have

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generally made two vortexes in the air flow.

The velocity distribution contour and streamlines diagram in a room with coated glass on the mentioned symmetry plane is shown in Error! Reference source not found.. It is obvious that the main cause of flow is external wall and there is an extensive vortex made by

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external wall; and the vertex made due to the upward flow nearest the irradiated floor is shrunk.

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Another important parameter that affects the PMV calculations and the efficiency of radiant panel system is relative humidity which is significant parameter to avoid condensation. In the

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present study, the ultimate possible limit of the relative humidity is determined.. As it is provided in Table 10, in the certain air temperature and different ceiling surface

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temperatures, condensation would not occur if the relative humidity is equal or less than the

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critical values in it. By comparing the outdoor relative humidity of Rasht (RH=49%) and its critical relative humidity when using clear glass (RH=38.85%), it is necessary to dehumidify the supply air in order to avoid condensation. Such problem is vanished by coated glass (RH=60%). In other cities critical relative humidity is much higher than outside relative humidity which is shown better performance of the system.

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ACCEPTED MANUSCRIPT 6. Conclusion In the present study, the performance of the radiant cooling panels when the room is under direct solar irradiation is evaluated in terms of condensation restrictions and the amount of output heat flux from the ceiling to provide comfort conditions. At the same conditions, the effect of using coated glass on the amount of cooling load is assessed. All investigations are

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carried out in four cities of different climates of Iran. To compare the effect of orientation, assessments are done in two cases of rooms with

southern and western external wall; It is concluded that the room with western external wall is critical orientation; in the room with a southern window in comparison to western window,

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ceiling temperature could be 9°C to 10°C higher; subsequently the output heat flux from the ceiling could be virtually 67 to 75 percent lower so as to provide comfort conditions, which means significant energy saving.

By coated glass, the maximum ceiling temperature could be raised up to 6°C and

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subsequently the output heat flux from the ceiling could be decreased approximately 49 to 58 percent.

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The temperature distribution contours show that by using coated glass, solar radiation would damp and this could cause lower surface temperature.

PT

Streamlines on the symmetry plane in the case of using clear glass show that the irradiated surface on the floor could cause an upward flow in addition to the upward flow due to

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external wall; and this could cause two vortexes in the flow. On the other hand, using coated

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glass would damp the undesirable temperature rises on the floor, an upward flow and the resulting vortex; therefore, the velocity distribution would be more uniform. By defining critical relative humidity based on air temperature and ceiling surface temperatures, the maximum limit of relative humidity to avoid condensation is determined. By using clear glass this limit is almost 37%, whilst it is approximately in the range of 55% to 61% by using coated glass.

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ACCEPTED MANUSCRIPT References

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ABE, K., Kondoh, T., & Nagano, Y. (1994). A new turbulence model for predicting fluid flow and heat transfer in separating and reattaching flows - 1. Flow field calculations. Int J Heat Mass Transfer, 37, 139-151. Alduchov, O. A., & Eskridge, R. E. (1996). Improved Magnus form approximation of saturation vapor pressure. J. Appl. Meteor., 35, 601-609. ASHRAE. (2000). Panel heating and cooling ASHRAE Handbook, HVAC systems and equipment (SI). USA: American Society of Heating. Asimakopoulos, D. A., Santamouris, M., Farrou, I., Laskari, M., Saliari, M., Zanis, G., . . . Giannakopoulos, C. (2012). Modelling the energy demand projection of the building sector in Greece in the 21st century. Energy and Buildings, 49, 10. Awbi, H. B. (1998). Calculation of convective heat transfer coefficients of room surfaces for natural convection. Energy and Buildings, 28, 219-227. Balaras, C. A., Dascalaki, E., & Gaglia, A. (2007). HVAC and indoor thermal conditions in hospital operating rooms. Energy and Buildings, 39, 454-470. Cannistraro, G., Franzitta, G., Giaconia, C., & Rizzo, G. (1992). Algorithms for the calculation of the view factors between human body and rectangular surfaces in parallelepiped environments. Energy and Buildings, 19, 51-60. Catalina, T., Virgone, J., & Kuznik, F. (2009). Evaluation of thermal comfort using combined CFD and experimentation study in a test room equipped with a cooling ceiling. Building and Environment, 44, 1740-1750. Causone, F., Corgnati, S. P., Filippi, M., & Olesen, B. W. (2010). Solar radiation and cooling load calculation for radiant systems: Definition and evaluation of the Direct Solar Load. Energy and Buildings, 42, 305-314. Cholewa, T., Anasiewicz, R., Siuta-Olcha, A., & Skwarczynski, M. A. (2017). On the heat transfer coefficeients between heated/cooled radiant ceiling and room. Applied Thermal Engineering, 117, 76-84. CIBSE. (2006). CIBSE KS06: Comfort, the Chartered Institution of Building Services Engineers CIBSE. Corgnati, S. P., Perino, M., Fracastoro, G. V., & Nielsen, P. V. (2009). Experimental and numerical analysis of air and radiant cooling systems in offices. Building and Environment, 44, 801-806. F.Modest, M. (2013). The method of discrete ordinate Radiative Heat Transfer (pp. 541-542). USA: Elsevier. Fanger, P. O. (1972). Thermal comfort. New York: Mc Graw-Hill. Feng, J. D., Schiavon, S., & Bauman, F. (2016). New method for the design of radiant floor cooling systems with solarradiation. Energy and Buildings, 125, 9-18. Gao, S., Wang, Y. A., Zhang, S. m., Zhao, M., Meng, X. Z., Zhang, L. Y., . . . Jin, L. W. (2017). Numerical investigation on the relationship between human thermal comfort and thermal balance under radiant cooling system. Energy Procedia, 105, 2879-2884. ISO. (2005). International Standard 7730. Ergonomics of the thermal environment- Analytical determination and interpretation of thermal comfort using calculation of the PMV and PPD Indices and local thermal comfort criteria. Geneva: International Organization for Standardization. ISO. (2007). ISO 7726:1998, Ergonomics of the thermal environment - Instruments for measuring physical quantities. Geneva: International Standard Organization. Karadag, R. (2009). The investigation of relation between radiative and convective heat transfer coefficient at the ceiling in a cooled ceiling room. Energy Conservation and Management, 50, 1-5. Karmann, C., Schiavon, S., & Bauman, F. (2017). Thermal comfort in buildings using radiant vs. all-air systems: A critical literature review. Building and Environment, 111, 9.

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Khan, Y., Khare, V. R., Mathur, J., & Bhandari, M. (2015). Performance evaluation of radiant cooling system integrated with airsystem under different operational strategies. Energy and Buildings, 97, 118-128. LBNL. Windows and Daylighting. Retrieved from http://windows.lbl.gov Li, R., Yoshidomi, T., Ooka, R., & Olesen, B. W. (2015). Field evaluation of performance of radiant heating/cooling ceilingpanel system. Energy and Buildings, 86, 58-65. materials, E. v. f. c. http://www.infrared-thermography.com/material-1.htm. Mikeska, T., Fan, J., & Svendsen, S. (2017). Full scale measurements and CFD investigations of a wall radiant cooling system integrated in thin concrete walls. Energy and Buildings, 139, 5. Ning, B., Chen, Y., Liu, H., & Zhang, S. (2016). Cooling capacity improvement for a radiant ceiling panel with uniform surface temperature distribution. Building and Environment, 102, 64-72. Olesen, B. (1997). Possibilities and limitations of radiant cooling. ASHRAE Trans., 103. Palyvos, J. A. (2008). A survey of wind convection coefficient correlations for building envelope energy systems’ modeling. Applied Thermal Engineering, 28, 801-808. Pout, C. H., MacKenzie, F., & Bettle, R. (2002). Carbon dioxide emissions from non-domestic buildings: 2000 and beyond: BRE Energy Technology Centre, Watford. Rhee, K.-N., & Kim, K. W. (2015). A 50 year review of basic and applied research in radiant heating and cooling systems for the built environment. Building and Environment, 1-25. Roshan, G., Farrokhzad, M., & Attia, S. (2017). Defining thermal comfort boundaries for heating and cooling demand estimation in Iran's urban settlements. Building and Environment, 121, 35. Su, L., Li, N., Zhang, X., Sun, Y., & Qian, J. (2015). Heat transfer and cooling characteristics of concrete ceiling radiant cooling panel. Applied Thermal Engineering, 84, 170-179. Tye-Gingras, M., & Gosselin, L. (2012). Comfort and energy consumption of hydronic heating radiant ceilings and walls based on CFD analysis. Building and Environment, 54, 1-13. Weather-Underground. Weather History & Data Archive. Retrieved from http://www.wunderground.com Wolkoff, P., & Kjærgaard, S. K. (2007). The dichotomy of relative humidity on indoor air quality. Environment International, 33, 850-857. Xamán, J., Arce, J., Álvarez, G., & Chávez, Y. (2008). Laminar and turbulent natural convection combined with surface thermal radiation in a square cavity with a glass wall. International Journal of Thermal Sciences, 47, 1630-1638. Y, H. (2005). Numerical study on airflow in an office room with a displacement ventilation system. Building Simulation 2005 Proceedings, 15-18. Yin, Y. L., Wang, R. Z., Zhai, X. Q., & Ishugah, T. F. (2014). Experimental investigation on the heat transfer performance and water condensation phenomenon of radiant cooling panels. Building and Environment, 71, 15-23. Nomenclature

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Width of the surface ( ) length of the surface ( ) Perpendicular distance between surface and an object ( ) Specific heat capacity ( ⁄

,

)

, Model constants of

model

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ACCEPTED MANUSCRIPT Clothing surface area factor ,

Model functions of low-Reynolds number

model

view factor between an object and th surface Convection heat transfer coefficient ( ⁄

)

Convection heat transfer coefficient on cooling ceiling ( ⁄

Clothing insulation (

⁄ )

Blackbody intensity (Planck function) Intensity of radiation

Metabolic rate ( ⁄

)

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Thermal conductivity ( ⁄

)

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Radiation heat transfer coefficient on cooling ceiling ( ⁄

)

)

Water vapor partial pressure ( Predicted Mean Vote

)

r

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Position vector (m)

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Predicted Percentage of Dissatisfied

Turbulent Reynolds number

⁄

̂

PT

Relative humidity

Unit vector into a given direction

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Time ( )

Air temperature ( )

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Clothing surface temperature ( )

̅

Mean radiant temperature ( ) Dimensionless temperature ((Tair – Tcold wall temp.)/(Tmax glass temp.- Tcold wall temp.)) Volume average room temperature ( ) ( ) Ceiling surface temperature ( ) Ceiling surface temperature ( )

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ACCEPTED MANUSCRIPT Dew point temperature ( ) Air temperature ( ) ̅

Temperature of the th surface ( ) Mean radiant temperature ( ) Cool surface temperature ( )

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Turbulent fluctuation in i-direction Wind velocity ( ⁄ ) ̅

Mean velocity ( ⁄ ) Relative air velocity ( ⁄ )

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Air velocity ( ⁄ ) Effective mechanical power ( ⁄

)

Cartesian coordinate in streamwise direction with x=0 at step location Dimensionless coordinates, (x/room height)

⁄ )

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Non-dimensional length from wall surface ( Greek letters

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Extinction coefficient Kronecker delta

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Turbulent kinetic energy dissipation rate (

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Emissivity coefficient Turbulence kinetic energy (

⁄ )

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Wavelength ( ) Constant Kinematic viscosity ( Eddy viscosity ( Density(

⁄

⁄ )

⁄ )

)

Absorption coefficient

23

⁄ )

ACCEPTED MANUSCRIPT ,

Model constants in turbulent diffusion terms of

model

Scattering coeficient constant Solid angle Subscript

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Air critical Dew point

1,2 and 3 denote x-, y- and z-directions, respectively

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Denote the parameter value between an object and ith or Ath surface Surface

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Corrected

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Table 1. Damping functions using in Low-Re- k-ε. , ( ⁄ )- 0 . ⁄ ,

(

⁄

)- ,

⁄

1.5

⁄

1.9

) +-

1.4

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0.09

* (

) +1

⁄

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Table 2. Low-Re- k-ε model constants.

* (

/

1.4

Table 3. Thermo-physical properties of glass.(LBNL)

(

( ) 0.006

)

2500

(⁄

)

750

( ⁄

)

1.4

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Glass

⁄

Wavelength Band

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Glass

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Table 4. Optical properties of clear glass and coated glass.

Clear Glass

Transmission

Absorption

Reflection

Coefficient

Coefficient

Coefficient

0.771

0.159

0.07

0.884

0.036

0.08

0.234

0.402

0.364

0.199

0.296

0.532

Direct Visible

(

)

(103)

Direct IR (

)

Direct Visible Coated Glass (

)

(11389) Direct IR

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ACCEPTED MANUSCRIPT (

)

-191.1

45×45×45

-193.7

50×50×50

-197.9

60×60×60

-197.2

-

1.4 2.2 0.3

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40×40×40

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Table 5. Output heat flux and corresponding relative error associated with different mesh sizes. Relative Error Ceiling Output Heat Flux Mesh Size (%) ⁄

0.045845921

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Table 6. Calculated View factors based on a subject standing at the center of the room. View Factor Surface Glass Window Inner Walls

0.116624644

Out Wall

0.222459825

Floor

0.127657918

Ceiling

CE

PT

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3×0.162470564

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Table 7. Air temperature and relative humidity of air in Tabriz, Tehran, Rasht and Yazd.(Weather-Underground) City

Date/Time

Outdoor Air Temperature (°C)

Outdoor RH (%)

Tabriz

Aug 22/16:30

40

6

Tehran

Jun 30/16:00

41

7

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ACCEPTED MANUSCRIPT Rasht

Aug 3/16:00

35

49

Yazd

Jun 30/16:00

42

5

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Table 8. Sun direction vector, geographical position and direct solar irradiation in hottest hour of hottest day of the year 2015 in Tabriz, Tehran, Rasht and Yazd. City

Solar direction vector

Longitude

Latitude

Solar irradiation ( ⁄

)

i

j

k

Tabriz

-0.0997785

0.535173

0.838 829

Tehran

0.0362937

0.626634

0.778468

51.42

35.69

781.475

Rasht

-0.0890039

0.611077

0.786552

49.58

37.28

790.611

Yazd

0.0977984

0.587068

0.803609

54.37

31.90

764.329

38.07

764.547

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46.30

Table 9. Radiant cooling ceiling surface and room air temperature, ceiling out put heat flux,

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room air velocity and mean radiant temperature in thermal comfort condition in the two

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rooms with an external wall with clear and coated glass. Ceiling

Ceiling surface

Glass

AC

City

Room air

Room air

Mean Radiant

temperature

velocity

Temperature

(℃)

(℃)

(℃)

output heat

temperature flux (℃) ( ⁄

)

Clear 10.0

186.0

25.54

0.14

27.23

16.0

91.7

25.07

0.08

25.84

Glass Tabriz Coated Glass

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ACCEPTED MANUSCRIPT Clear 10.0

195.2

25.99

0.14

27.80

16.0

93.4

25.65

0.10

25.57

11.0

179.4

26.09

0.17

27.65

17.0

74.0

25.03

10.0

190.0

25.69

16.0

97.6

Glass Tehran Coated Glass Clear Glass Coated Glass Clear Glass Yazd

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Coated

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Rasht

25.54

25.53

0.14

26.96

0.10

25.67

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Glass

0.08

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Table 10. Radiant cooling ceiling surface and room air temperature, critical and appropriate relative humidity in thermal comfort condition in the two room with an external wall with clear

Ceiling surface

Critical Room air

Glass

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City

CE

PT

and coated glass.

Sufficient Relative Relative

temperature (℃)

Humidity ( )

temperature (℃) Humidity ( )

Clear Glass

10.0

25.54

37.55

37.55

Coated Glass

16.0

25.07

57.15

57.15

Clear Glass

10.0

25.99

36.56

36.56

Coated Glass

16.0

25.65

55.22

55.22

Clear Glass

11.0

26.09

38.85

38.85

Coated Glass

17.0

25.03

61.05

60.00

Tabriz

Tehran

Rasht

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ACCEPTED MANUSCRIPT Clear Glass

10.0

25.69

37.21

37.21

Coated Glass

16.0

25.54

55.58

55.58

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Yazd

Figure 1. 3(m) ×3(m) ×3(m) cubic room including an out-wall and glass window and

PT

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equipped by radiant cooling ceiling system.

5

AC

hcc (w/m2)

6

CE

7

4

Present Study

3

[Karadag, 2009]

2 0

5

10

15

(Ti-Tc) (°C)

29

20

25

30

ACCEPTED MANUSCRIPT Figure 2. Graph of convective heat transfer coefficient at temperature difference between

AC

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room air temperature and ceiling surface temperature.

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3 Present study

[Karadag, 2009]

1

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hcr/hcc

2

0 0

5

10

15

20

25

(Ti-Tc) (°C)

30

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Figure 3. Graph of ratio of radiation heat transfer coefficient to convective heat transfer coefficient at temperature difference between room air temperature and ceiling surface

M

temperature.

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1

PT

0.8

CE

T*

0.6 0.4

[Xaman et al, 2008] Present Study

AC

0.2 0

0

0.2

0.4

0.6

0.8

x* Figure 4. Non-dimensional horizontal air temp profile at non-dimensional room length.

31

1

ACCEPTED MANUSCRIPT

Start 1

Input an arbitrary ceiling surface temperature (Tcc) 2

Solving governing equations in a room with ceiling temperature of Tcc in FLUENT

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3

Extracting volume average room air temperature (Tair), velocity (vair), Surface temperatures (Ts), Output heat flux from ceiling

5

Calculating critical relative humidity (RHcr) by entering Tair and vair into August-Roche-Magnus approximation

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4

Calculating mean radiant temperature (Tmrt) by entering room surface temperatures and calculated view factors associated with room geometry into the MRT calculation code

Yes

6

M

No

RHcr>60 Calculating PMV and PPD by entering Tair, vair, RH, Tmtr into PMV calculation code

ED

30≤ RHcr≤60

No

Yes

8

PT

RH= RHcr

7

RH= 60

CE

10

AC

0.3≤ PMV≤0.5

Yes

No

Considered Tcc is unapt

Considered Tcc is apt 9 END

Figure 5. Numerical and analytical solution process to determine appropriate radiant cooling ceiling temperature.

32

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Figure 6. Temperature distribution contours on inner surfaces of the room with a south

AC

CE

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external wall and clear glass in Tabriz.

Figure 7. Temperature distribution contours on inner surfaces of the room with a western external wall and clear glass in Tabriz.

33

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Figure 8. Temperature distribution contours on inner surfaces of the room with a western external wall and coated glass in Tabriz.

ED

3 2.5

Y

1.5

PT

2

CE

0.38 0.36 0.34 0.32 0.30 0.28 0.26 0.24 0.22 0.20 0.18 0.16 0.14 0.12 0.10 0.08 0.06 0.04 0.02

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velocity-magnitude

1

AC

0.5 0

0

0.5

1

1.5

2

2.5

3

X

Figure 9. Velocity magnitude and streamlines at the center plane of the room perpendicular to the external western wall and clear glass.

34

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velocity-magnitude

3 2.5 2 1.5 1 0.5

0

0.5

1

1.5

2

2.5

X

3

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0

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Y

0.38 0.36 0.34 0.32 0.30 0.28 0.26 0.24 0.22 0.20 0.18 0.16 0.14 0.12 0.10 0.08 0.06 0.04 0.02

Figure 10. Velocity magnitude and streamlines at the center plane of the room perpendicular

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to the external western wall and coated glass.

35