Thermal performance evaluation of hybrid heat source radiant system using a concentrate tube heat exchanger

Accepted Manuscript Title: Thermal Performance Evaluation of Hybrid Heat Source Radiant System Using a Concentrate Tube Heat Exchanger Author: Young Tae Chae Richard K. Strand PII: DOI: Reference:

S0378-7788(13)00795-0 http://dx.doi.org/doi:10.1016/j.enbuild.2013.11.078 ENB 4679

To appear in:

ENB

Received date: Revised date: Accepted date:

9-6-2013 15-11-2013 23-11-2013

Please cite this article as: Y.T. Chae, R.K. Strand, Thermal Performance Evaluation of Hybrid Heat Source Radiant System Using a Concentrate Tube Heat Exchanger, Energy and Buildings (2013), http://dx.doi.org/10.1016/j.enbuild.2013.11.078 This is a PDF file of an unedited manuscript that has been accepted for publication. As a service to our customers we are providing this early version of the manuscript. The manuscript will undergo copyediting, typesetting, and review of the resulting proof before it is published in its final form. Please note that during the production process errors may be discovered which could affect the content, and all legal disclaimers that apply to the journal pertain.

*Manuscript

Thermal Performance Evaluation of Hybrid Heat Source Radiant System Using a Concentrate Tube Heat Exchanger

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Young Tae Chae1 ([email protected]), Richard K Strand2 ([email protected]) IBM T.J. Watson Research Center, Yorktown Heights, NY, United States

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School of Architecture, University of Illinois at Urbana-Champaign, Champaign, IL, United States

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ABSTRACT

This study developed a new radiant system concept having a concentric tube heat exchanger

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embedded in a radiant panel and evaluated the system characteristics. The concentric tube heat

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exchanger allows two fluids, air and water, to flow in the same direction. The outdoor air for the space ventilation requirement passes through an inner tube. The primary heat transfer medium,

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water, flows through the outer tube exchanging heat with both radiant panel and air tube. Two

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fluids would have an identical temperature condition by the characteristics of the heat exchanger. The system configuration enables simultaneous satisfaction of the space sensible loads and ventilation requirement. It has also flexibility to use outdoor air directly at night to reduce any

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heat built up during the day in the thermal mass.

A numerical analysis model based on

computational fluid dynamics (CFD) has evaluated this conceptual system. Comparing this new system with the performance of a typical system incorporated with a dedicated outdoor air system (DOAS), the proposed system is shown to provide a more closed the room set temperature condition and a better local thermal environment. In addition, the developed system can improve the performance of the night ventilation operation using nocturnal outdoor air.

Keywords : Hydronic radiant panel system, Hybrid system, Concentric tube heat exchanger, Night ventilation, Computational Fluid Dynamics 1 Page 1 of 39

1. Introduction Radiant panel systems use controlled temperature surfaces such as the floor, walls or the ceiling to maintain thermally comfortable conditions within spaces. These surface temperatures

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are maintained by circulating heat transfer media like water or air or by using electric resistance

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[1]. The systems have been reported to have advantages in improving occupants’ thermal comfort and reducing heating and cooling energy consumption [2-3]. It was also the main focus

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of International Energy Agency (IEA) ECBCS Annex 37: low Exergy systems for heating and cooling buildings [4]. Although the conventional radiant system has advantages for the

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occupants’ thermal comfort and energy saving, it has also drawbacks to be overcome such as the

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start-up issue due to high thermal mass, sensible load only control, potential system interrupt due to surface condensation in cooling mode, and satisfaction of ventilation requirements.

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Many researches have pointed out that conventional radiant systems need to be incorporated

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in conjunction with a convective forced air system to minimize the shortcomings. McDonell [5] claims that radiant heating and cooling systems alone may not serve as a perfect space conditioning system because there are three human comfort parameters: radiation comfort (40%

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to 50% of the human comfort equation), air movement (30% of the human comfort equation) and indoor air humidity (10% to 15% of the human comfort equation). Therefore, the radiant system must be integrated with a forced-air system to improve occupant’s thermal comfort and minimize the possibility of condensation (radiant cooling systems only). Additionally, the author presented some practical data for the radiant heating and cooling capacity and concluded that the radiant cooling system can be designed to work in nearly all applications and most climate zone without fear of condensation when combined with a conventional air system that deals with outdoor air requirements. Oxizidis and Papadopoulos [6] also pointed out that a hybrid system, radiant floor

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panel incorporated with forced air system, provides a better thermal environment with relatively low energy consumption than other systems in cooling period. There have been various attempts to decouple the outside air load due to ventilation

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requirements from the space conditioning load to avoid oversizing the HVAC system [7-9]. According to the accepted definition of this type of load-sharing or hybrid system, a radiant

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system is combined with a conventional forced air system to provide the conditioning needs of

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the space. In this case, the radiant system takes over most or all of the space sensible cooling while the air system controls the latent load and outdoor air requirements. One convective forced

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air system designed to control the ventilation requirement is the dedicated outdoor air system (DOAS). It is 100% outdoor air constant volume system designed to deliver the volumetric flow

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rate of ventilation air to each conditioned space in the building. Mumma, et al., [10-11]

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developed the system concept and has been implemented them in buildings since the late 1980’s.

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The authors proposed and evaluated DOASs in conjunction with a cooling radiant ceiling panel (CRCP) system. This system combined an outside air control system and metal CRCP. According to steady-state simulation results comparing this combination system with a

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conventional VAV system, the DOAS with CRCP system can reduce the annual total chiller energy consumption by 25%. Although the authors showed remarkable potential for the DOAS with CRCP system as an energy conserving alternative, it can been argued that implementation of this system requires at least eight different components to make the system work properly. This increases the initial cost and makes system control and maintenance more difficult. In addition, the cooling or heating demand might not be offset using potentially less expensive offpeak energy because the CRCP is composed of a metal panel so that the system has little to no thermal mass effect.

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Kilkis [12] developed an analytical model for a hybrid heating and cooling terminal unit composed of a perforated metal deck and raised floor air chamber. Outside air or room return air passes through the plenum space of the raised floor and diffuses through the perforated holes of

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the metal deck. The deck acts as a typical radiant system but its convective heat transfer is enhanced by air passing through the perforated holes. Based on this system design, the author

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suggested the optimal design parameters for this system for the renovation of an old library

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building. Although this seems like an effective method to improve the capacity of an existing HVAC system for an historic building, the author also claimed that the system has several

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challenges such as concerns about condensation which require surface temperature control of the metal deck, mold-growth in the porous carpet, chemical and dust effect of the carpet, air flow

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direction, and handling water from condensation in the air plenum.

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The benefits of the radiant system in conjunction with a forced-air system have been

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evaluated by several field measurement studies. Baskin tested the heating and cooling energy performance of a residential house with both a radiant and a convective air system. The precooling operation of the radiant slab was effective in reducing the cooling demand and shifting it

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from the on-peak time to early morning (off-peak). If the radiant slab was pre-cooled with the night ventilation operation of the convective air system, the energy demand of the daytime would be shifted to the nighttime [13-16].

This strategy is helpful in locations where the local power company supplies electrical power at a reduced rate during off-peak/nighttime periods. Scheatzle [17] monitored a residential building with a radiant system and heat recovery unit with dehumidifier for four years in a hot and arid climate. According to the author’s findings, the thermally massive radiant heating and

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cooling system combined with convective dehumidification/ventilation can be an effective system option for a residential building. Based on previous work, the performance of a radiant system can be enhanced when it is

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integrated with a convective forced air system for outside air control. However, it requires additional work to control the two different system types and additional space for the air

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handling unit. Furthermore, its inherent complexity makes it a more difficult HVAC system to

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design and size when the two systems use different heat transfer media.

The purpose of this study is to design a new radiant system that can satisfy the ventilation

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requirement and space thermal load simultaneously. It verifies the thermal performance of the proposed system by comparing a convective forced air system integrated with a conventional

2.1 Background

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2. Concept development

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hydronic radiant system based on numerical analysis.

Even though a radiant panel heating and cooling system can be categorized by the heat

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transfer media types, the thermal characteristic of building components used as the radiant panel, and the construction layout, a hydronic radiant system with water tubes embedded in a concrete panel is the primary focus of this study. There are two considerations in the design and construction of a hydronic radiant system. First, the tubes of typical embedded hydronic radiant systems are constructed as continuous tubes. The spacing of the tubes ranges from 150mm to 450mm on centers [1]. The total tube length will range from 2m to 6m per unit area (m2) of a radiant panel. If the tube, which is embedded into the slab, has a diameter of 15mm, the contact surface area is in-between 0.094(m2) and 0.28(m2)

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inside of the panel. This provides a relatively large heat exchange area between the water loop and floor slab. Second, the water temperature of a hydronic radiant system is restricted to avoid excessively

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hot or cold surface temperatures. It is common to design for an 11 °C temperature drop for heating across a given system grid and a 3 °C rise for cooling. However, the surface temperature

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for a radiant floor heating system in an occupied space is recommended to be between a lower

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limit of 19 °C for cooling and an upper limit of 29 °C for heating to avoid thermal comfort problems associated with too hot or too cold surfaces [18]. This means that the supply and return

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water temperature are often around 35 °C and 29 °C, respectively, for heating and 20 °C and 19 °C, respectively, for cooling.

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From these design guidelines for hydronic radiant systems, it is possible to conclude that the

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temperature of the heat transfer medium (water) of a radiant system is lower for heating and

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higher for cooling than a conventional forced air system. It can also be seen that the radiant system has a relatively large surface area to exchange heat between the tube and panel. Applying these guidelines to the concentric tube radiant system with the water exchanging heat with both

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the slab and the air loop simultaneously and assuming that the air can exchange heat completely with the water tube, we can anticipate that the outlet temperature of the air will be approximately 29 °C for heating and 19 °C for cooling. Outlet air temperatures within this range would not only provide for conditioning of outdoor air to meet the ventilation requirement, but it would also reduce the space thermal loads by convective heat transfer when delivered into the space. These arguments and the relative simplicity of the concentric tube radiant system when compared to the combined radiant/DOAS system provide the basis for the conceptual development described in the sections that follow.

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2.2 System configuration This study introduces a concentric tube arrangement with air flowing through the inner tube and water through the outer tube, the outdoor air for the ventilation requirements is conditioned

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within the panel as illustrated in Fig. 1. This, then, allows the proposed radiant system to satisfy the sensible loads of the space and provide the outdoor air to meet the ventilation requirements

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without resorting to the combination of more complex all-air system like a DOAS with a radiant

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system. In addition, this configuration may allow using ventilation through the concentric tube system to provide free night-time cooling and shift day-time thermal loads to night-time.

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The proposed system has a configuration that is similar to a typical hydronic radiant system with respect to system installation. It has a hot and chilled water circulation system and a water

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channel embedded into a radiant floor in a space. Additionally, it adds an air circulation

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component and an air tube placed inside of the water tube as illustrated in Fig. 2.

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The air circulation system introduces outdoor air into the air tube and then conditioned by passing through the air path parallel to the water tube. When the concentric tubes reach the end of the radiant panel, the air tube separates and the conditioned air is delivered into the space

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directly and the water is re-circulated to the plant equipment. It has named Hybrid Heat Source Radiant (HHSR) system because it has two heat transfer media, air and water, and enables the system to use each medium simultaneously or individually.

3. Model description Although the proposed system has the theoretical potential to satisfy the space thermal and ventilation loads and utilize outside air directly as a heat sink/source as discussed above, it is necessary to investigate whether the proposed system can be operated as intended. This study has 7 Page 7 of 39

applied a unit slab and a room scale computational fluid dynamics (CFD) model to evaluate the thermal characteristics of the proposed system.

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Numerical simulations for a real scale model were performed to evaluate the system performance of three different types of radiant systems with a CFD tool, ANSYS CFX 11.0. This

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code is based on the finite-volume method, and it solves the Reynold-Average Naviar-Strokes

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equations of the fluid for each finite volume [25]. 3.1 Unit slab model

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A plain slab model embedded the concentrated tube heat exchanger was generated. The concentric tube heat exchanger was located in the middle of the slab. Air flows through the inner

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tube and water flows in the outer shell which is in contact with the panel directly. Both fluids flow parallel to each other in the heat exchanger. To investigate the thermal response of the panel

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to different loads in heating and cooling operation, various heat fluxes have been applied

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uniformly at the entire surface of the slab, while bottom surface has adiabatic condition and other sides have symmetry boundaries.

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Although it is a simple planar slab that is 0.15m wide and 15m long, the modeling of this simple slab can be effective in the investigation of the interactive heat exchange not only from the panel to the heat exchanger, but also between the water and the air tube simultaneously. The thermo-physical properties of solid materials and fluid boundary conditions are illustrated in Table 1.

3.2 Room model A room scale CFD model was also generated to evaluate the indoor thermal conditions and plant load under both winter and summer design day conditions for O’Hare international airport at Chicago, USA. To compare the thermal performance, a conventional hydronic radiant system 8 Page 8 of 39

with a dedicated outdoor air system (DOAS+RAD system). The DOAS conditions outside air to be the same temperature as the room design temperature condition and radiant panel manages the

before being supplied into the radiant panel as described in Fig. 3.

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rest of the room sensible load. Hot and chilled water flow to the heating or cooling coil of DOAS

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A simple room in Fig. 4 with one exterior wall and one exterior window is considered for this evaluation process. The main characteristics of the sample room are summarized in Table 2. All

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the variables for this test case are based on the recommended data [1], and the required

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ventilation rate calculated by ASHRAE Standard 62 [19]. In the case of HHSR system, the outside diameter of water tube is larger than the other cases due to the internal air tube, but the

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hydraulic diameter of the water tube is identical to the conventional radiant system with the dedicated outdoor air system.

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Due to the asymmetry of the physical domain, the computational mesh model has two

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different air domains (the space itself and the inner tube of concentric tube heat exchanger), two solid domains (the radiant panel in the space and the tube material), and a water domain for the

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outer shell of the heat exchanger. There are various parameters that affect the reliability of the CFD simulation such as the grid spacing, the fluid model for turbulence and buoyancy, and the radiation model. The effect of grid spacing on the computed results was minimized by increasing the total number of elements to about 7,600,000. This study selected the standard k   model as the turbulence solution model for the fluid domains such as the room, the water tube, and the air tube. It is also derived using a high Reynolds number hypothesis and treatment of conditions near the walls are based on the application of wall functions [20].

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For buoyant flows within the space where the density variation is assumed to be driven only by small temperature variations, the Boussinesq model is used to characterize natural convection [21]. The model assumes that the change in air density over the expected range of condition is

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relatively small, so it assumed the constant fluid density through the computational domain. However, it applies a local gravitational body force throughout the fluid which is a linear

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function of fluid thermal expansivity(  ) and the local temperature difference.

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For space thermal radiation modeling, it applied the discrete transfer method [22]. The method assumes that air is a participating media and that surfaces of the space are “gray”. The

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diffusivity of all surfaces in the room including the radiant floor is assumed to be 0.9, which is

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typical for building components that are non-reflective [23]. The method is based on tracing the domain using multiple rays leaving from the bounding surfaces.

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At the interfaces between the solid domain and the space air or water, there will be additional

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heat transfer to/from the surface besides radiation so heat transfer to/from the surface by convection and conduction must also be taken into account [24]. To solve this conjugate heat transfer problem, the CFD program calculates both fluid-side and solid-side temperatures based

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on heat flux conservation. These values are representative of the temperature within the halfcontrol volumes around the vertices on the interface. The root mean square (RMS) of the residuals for each equation such as energy of fluids, radiation, and total energy are calculated and convergence was taken to be attained at a RMS value of less than 1.0E−4, which from the CFD program viewpoint is considered to be very tight [25-26]. To solve the coupled system of partial differential equations, a high-resolution scheme using the first order backward Euler method was adopted. This scheme can provide second or higher

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order spatial accuracy in smooth portions of the solution and the solutions are free from spurious oscillations or wiggles. The high resolution scheme during most of the simulation maintains second order accuracy. A total of 8 different simulation cases based on the different radiant

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systems and ventilation methods were simulated as shown in the next section.

4.1 Slab with the concentric tube heat exchanger

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4. Results and discussions

Fig. 5 and 6 illustrate the averaged fluid temperature distributions of water and air along the

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concentric tube heat exchanger length in the unit slab model. Under both the heating and cooling

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conditions, the temperature distributions are similar to a typical concentric tube heat exchanger of parallel fluid flow. Although the temperatures of the water and the air are not exactly identical

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at the outlet under an adiabatic condition, the difference of the average fluid temperature at the

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outlets (X=15m) are less than 0.07°C for heating and 0.017 °C for cooling, respectively. When the slab has a heating and cooling flux on the top of the surface, the balance point, where both fluids have an identical outlet temperature, depends on the load condition. As

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illustrated in Fig. 5, the lower load condition requires more heat exchanger length for the two fluids to come to the same temperature. For example, this point is reached at X=13.39 m with a 50 W/m2 heating load on the surface while it is reached at X=8.57m with a 400 W/m2 heating load. Similar results are obtained in cooling mode as shown in Fig. 6. Once the temperature has been identical, air temperature distribution follows on the water temperature dominated by thermal load on the slab. Due to the assumption that the water tube directly contacts on the slab and thermal resistance of the internal tube between water and air fluid, the average sectional water temperature is slightly higher for cooling and lower for heating

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than the air temperature near the water outlet. However, the temperature difference at the outlets is negligible except the extreme load conditions. The air outlet temperature ranges from 27.89 °C ( qh = 500 W/m2) to 29.09 °C ( qh =100 W/m2)

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for heating, and from 20.45 °C ( qc =100 W/m2) to 21.69 °C ( qc =400 W/m2) for cooling,

supplies the air into a space as described in the previous section.

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respectively. This temperature condition may a feasible convective heat source when the system

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From these studies, it can be seen that the concentric tube heat exchanger embedded into a

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radiant panel can result in outdoor air that is conditioned to a temperature that is close to the

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water outlet temperature under various thermal load conditions.

4.2 Radiant panel temperature distributions

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Fig. 7 illustrates the temperature distributions on the radiant panel surface for a heating design

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day condition in the example room. Using the DOAS+RAD system, the average panel surface temperature is 25.77 °C (  =0.69) with a ceiling diffuser and 25.80 °C (  =0.70) with a wall outlet. The proposed system controls the average surface temperature to 26.7 °C (  =0.65) with

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a ceiling diffuser and 26.90 °C (  =0.62) for a sidewall diffuser. The location of air diffuser affects the surface temperature variations. Compared with the ceiling diffuser cases, the overall surface temperature of the radiant panel of each system using a wall supply diffuser is slightly higher. This is explained by the fact that when an air inlet is placed near the radiant panel the surface convection heat transfer is enhanced. Fig. 8 shows the temperature distributions on the panel surface in cooling design day condition. In the cases with the ceiling diffuser, the average surface temperature of the DOAS+RAD system is 21.53 °C (  =0.30) and the temperature of the HHSR system is 21.45 °C 12 Page 12 of 39

(  =0.33). When the tempered outdoor air is supplied through a wall diffuser, the average temperature is 21.52 °C (  =0.09) for the DOAS+RAD system and 21.37 °C (  =0.33) for the

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than the other system when subjected to cooling design day conditions.

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HHSR system, respectively. Thus, the surface temperature of HHSR system is slightly lower

4.3 Room air temperature distributions

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Panel surface temperature and supply air condition may affect the air temperature distribution of a space. Fig. 9-10 provide comparisons of vertical room air temperature distributions in

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heating and cooling mode at the center plane cross-section (X=1.0m) for each case. Under heating design day conditions, the DOAS+RAD system maintained the average room

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temperature at 20.31 °C (  =0.71) with the wall diffuser and 20.22 °C (  =0.81) with the ceiling

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air outlet. The HHSR system controlled the average room temperature at 21.50 °C (  =0.76) and

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21.51 °C (  =1.17) with the wall and ceiling diffusers, respectively. Although both systems are not fully successfully at satisfying the set-point condition for the air temperature (22 °C) due to the thermal characteristics of the radiant system, the proposed HHSR system controlled the

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average temperature much closed to the set-point temperature condition and uniformed room air temperature.

Under cooling design day conditions, the average room air temperature of DOAS+RAD system is 25.90 °C (  =0.68) and 25.60 °C (  =1.14), for ceiling and wall diffuser configurations, respectively. For the HHSR system in cooling operation, the room temperature is maintained at 25.41 °C (  =0.58) and 25.38 °C (  =1.74), for supply air through the ceiling and wall diffusers, respectively. When the air delivered thorough wall of the proposed system, it increase stratifications of room air temperatures due to the buoyancy by temperature difference

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between room and supplied air, as described in the previous study [27-28]. Based on this design condition and this room configuration for these two systems, there is not as clear a difference in the performance of the systems as in heating operation but the proposed system produced

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conditions closer to the set point temperature.

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4.4 Global thermal comfort index

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Predictive mean vote (PMV) and Predicted percentage of dissatisfied (PPD) are useful indices to estimate an occupant’s thermal sensation in a space. Table 3 shows the PMV and PPD values

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from the simulation results for each case. Using the ISO guideline [18] for these cases, the metabolic rate was defined as 1.2 met and the clothing level is set at 1.0 clo for heating and 0.5

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clo for cooling, respectively. As shown in the table, the PMV and PPD of both systems under

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design day conditions were close to the thermally neutral condition with less than 10% PPD.

4.5 Heating and cooling load of primary system side

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Based on the outlet temperature of water, the main heat transfer medium, the plant load for each operation mode and system configuration was estimated as shown in Fig. 10. The estimated plant load of the HHSR system shows an increase of 9% over the load of the DOAS+RAD system due to the additional heat transfer to the ventilation air in the HHSR system. For the cooling operation, the chilled water plant load of the proposed system requires almost 20% more chilled water load than the DOAS+RAD system. While this appears to show that the proposed system may not be as efficient as the DOAS+RAD system, the HHSR system does provide better thermal conditions within the space and that load comparisons can only be made when the level of comfort is truly equivalent between the systems. 14 Page 14 of 39

From another perspective, the result also shows the limitation of the steady state calculation. For example, the HHSR system has higher panel and room temperature in heating mode than the DOAS+RAD system. This could be an energy advantage under varying conditions because it

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should require less heat to maintain a comfortable indoor thermal environment. Therefore, to gain a full understanding of whether or not the HHSR system will provide better comfort and

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better energy performance, it is necessary to simulate these cases under more realistic operational

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4.5 Precooling application using outdoor air

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conditions and investigate transient effects.

As described in the previous section, the configuration of the proposed system may also allow

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the utilization of outdoor air as a heat transfer medium when the temperature difference between

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the inside and outside of the building is favorable during cooling or transient season, especially

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during the night-time. This strategy can reduce the temperature variation of the room during daytime and peak thermal demand [29-30] by storing cooling potential in the mass of the building component like a slab.

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To investigate the night ventilation performance of the proposed system, the previous CFD mesh of the space was employed and modified to assume an interior space to avoid any heat exchange through the exterior wall. The night ventilation with outdoor air has been started when the space was vacant (8 PM) and the outdoor air temperature (25.54 °C) was close to the condition of the space design temperature (24 °C). The space air and overall slab temperature at initial conditions were assumed to 30 °C after a typical weekday operation. To approximate the period from 8PM to 7AM, 12 steady-state simulations were conducted. Although this has various limitations such as the assumption that the outdoor air and indoor

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thermal conditions are not changing in any given hour and that one hour was sufficient time to arrive at a steady-state condition, the results will be helpful to evaluate the system performance and consider the viability of the proposed system.

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A conventional night ventilation model (CNV) was also considered to compare with the performance of the HHSR system. The conventional model introduces outdoor air into the space

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directly through the ceiling diffuser during the unoccupied period. Whilst the outdoor air for the

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HHSR system is introduced into the internal tube and circulated through the slab before being supplied to the space directly at the same air mass flow rate (1ACH). During the precooling

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process, the water in the outer tube is a stationary medium and the convection heat transfer of the fluids is not taken account. Fig. 12 shows the hourly outdoor air temperature profile as well as

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the averaged overall panel temperature and room air temperature for both HHSR and CNV for a

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cooling design day.

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In the case of the CNV, both the space and slab temperature follow the decline pattern of the outdoor air temperature. The space temperature is lower than the averaged overall slab temperature. The lowest slab and space temperatures are 24.41 °C and 23.89 °C at 6 AM,

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respectively. In contrast to the typical operation, the average panel temperature of the HHSR system in the period is lower than the space temperature. This demonstrates that the forced airflow in the tube embedded in thermal mass of the floor enhances the convective heat transfer coefficient and the outdoor air can contact a larger heat transfer area compared to natural convection alone [14]. Consequently, the average space air temperature is higher for the HHSR than the conventional nighttime ventilation configuration when the night ventilation began, 8PM12PM, since it is still extracting heat from the panel. However, the space temperature declines rapidly as the slab is cooled down using the outdoor air and the temperature difference between

16 Page 16 of 39

the slab and the space become closer as time elapses. The minimum slab and space temperatures for the HHSR system in the period are 21.77 °C at 5AM and 22.13 °C at 6AM, respectively. The averaged room air and slab temperatures are almost 2.10 °C lower than those resulting from the

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conventional night ventilation strategy by the time the space becomes occupied.

5. Conclusions

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This study proposed a new type of radiant system and evaluated the concept with respect to

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its thermal characteristics. The proposed system has a concentric tube heat exchanger allowing it to use two different heat transfer medium, water and air, in the radiant panel. The outdoor air for

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ventilation requirement exchanges heat with the water tube and then the conditioned outdoor air is delivered into the room. Using this configuration, the system can satisfy both the room

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sensible thermal load and the ventilation requirement simultaneously.

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To evaluate thermal characteristics of the proposed model, two CFD models under steady state conditions were generated. From the result of a plain slab model with the concentrate tube

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heat exchanger in the slab, both fluids have almost identical temperature at the outlets under various load conditions. In the case of the room model, the proposed system, HHSR system, in a particular condition, was able to control the average space temperature closer to the set-point temperature condition than the typical hybrid system in both heating and cooling operation, since the HHSR system supplied the outdoor air at the water outlet temperature. This improved thermal environment in the HHSR system resulted in about 10% more heating plant load and 20% more of cooling plant load than the typical hybrid system but this should be verified through transient simulation. In the perspective of occupants’ thermal sensation, both systems with two different air supply locations control PMV and PPD within the acceptable range. This study also 17 Page 17 of 39

investigated the thermal characteristics using the passive cooling strategy of night ventilation during cooling season. Comparing this method to the conventional approach, the space and slab temperatures of the proposed system reached more favorable temperature condition before the

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day-time operation.

Although there is certainly room for investigation on the practical application under actual

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scale experiments and evaluation of general performance of the system by using transient whole

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building simulation model, this study has been shown that the proposed system can provide an acceptable indoor thermal environment and the system can be used to take advantage of the

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lower nocturnal temperature of the outdoor air more effectively. Therefore, the system can be an alternative type of radiant system in terms of reducing system complexity, without additional coil

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and air duct work for conditioning outdoor air, and operation flexibility to allow using outdoor

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air as heat sink in cooling and transitional seasons.

Nomenclature

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1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24 25 26 27 28 29 30 31 32 33 34 35 36 37 38 39 40 41 42 43 44 45 46 47 48 49 50 51 52 53 54 55 56 57 58 59 60 61 62 63 64 65

 = density, kg/ m

3

k = thermal conductivity, W/m·K

c = sepcific heat, J/kg·K m = mass flow rate, kg/s

slab heating load, W/m2 2 qc = slab cooling load, W/m  = standard deviation U = overall heat trasnfer coefficient, W/ m2·K ACH = Air change per hour qh =

Acknowledgement This work was supported by a Dissertation Completion Grant from the Graduate College of the University of Illinois at Urbana-Champaign. 18 Page 18 of 39

References [1] ASHRAE, ASHRAE Handbook—HVAC Systems and equipment, ASHRAE Handbook,

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Atlanta: American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc (2012). [2] Imanari, Takehito, Toshiaki Omori, and Kazuaki Bogaki,Thermal comfort and energy

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consumption of the radiant ceiling panel system.: Comparison with the conventional all-air

us

system, Energy and buildings 30.2 (1999): 167-175.

[3] Olesen, Bjarne W., Radiant floor heating in theory and practice, ASHRAE journal 44.7

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(2002): 19-26.

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[4] Ala-Juusela, Mia, ed., Heating and Cooling with Focus on Increased Energy Efficiency and Improved Comfort: Guidebook to IEA ECBCS Annex 37 Low Exergy Systems for Heating and

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Cooling of Buildings, VTT, 2004.

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[5] McDonell, G., Is radiant cooling an option?, ASHRAE Journal 50.4 (2008):46-52. [6] Oxizidis, Simos, and Agis M. Papadopoulos, Performance of radiant cooling surfaces with

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respect to energy consumption and thermal comfort, Energy and Buildings (57) (2012) 199-209. [7] Kilkis, Ibrahim B., M. Sapci, and S. R. Suntur, Hybrid HVAC systems, ASHRAE journal 37.12 (1995):23-28.

[8] Conroy, Christopher L., and Stanley A. Mumma, Ceiling radiant cooling panels as a viable distributed parallel sensible cooling technology integrated with dedicated outdoor air systems, TRANSACTIONS-AMERICAN SOCIETY OF HEATING REFRIGERATING AND AIR CONDITIONING ENGINEERS 107.1 (2001): 578-588.

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[9] Khattar, Mukesh K., and Michael J. Brandemuehl, Separating the V in HVAC: A dual-path approach, ASHRAE journal 44.5 (2002): 37-43.

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[10] Mumma, Stanley A., and Brian W. Lee., Extension of the multiple spaces concept of ASHRAE Standard 62 to include infiltration, exhaust/exfiltration, interzonal transfer, and

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additional short-circuit paths, TRANSACTIONS-AMERICAN SOCIETY OF HEATING REFRIGERATING AND AIR CONDITIONING ENGINEERS 104 (1998): 1232-1244.

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[11] Mumma, Stanley A., and Jae-Weon Jeong, Direct digital temperature, humidity, and

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condensate control for a dedicated outdoor air-ceiling radiant cooling panel system, ASHRAE Transactions 111.1 (2005): 547-558.

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[12] Kilkis, Birol I., SYMPOSIUM PAPERS-HI-02-06 Radiant Panel Heating and Cooling: Recent Developments and Applications-Modeling of a Hybrid HVAC Panel for Library

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Engin 108.2 (2002): 693-697.

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Buildings, ASHRAE Transactions-American Society of Heating Refrigerating Airconditioning

[13] Baskin, Evelyn, SYMPOSIUM PAPERS-OR-05-03-Controlling Hydronic Radiant Heating

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and Cooling Systems-Evaluation of Hydronic Forced-Air and Radiant Slab Heating and Cooling Systems, ASHRAE Transactions-American Society of Heating Refrigerating Airconditioning Engin 111.1 (2005): 525-534.

[14] Ren, M. J., and J. A. Wright, A ventilated slab thermal storage system model, Building and Environment 33.1 (1998): 43-52. [15] Russell, M. B., and P. N. Surendran, Influence of active heat sinks on fabric thermal storage in building mass, Applied energy 70.1 (2001): 17-33.

20 Page 20 of 39

[16] Inard, Christian, Jens Pfafferott, and Christian Ghiaus, Free-running temperature and potential for free cooling by ventilation: A case study, Energy and Buildings 43.10 (2011): 2705-

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2711. [17] Scheatzle, David, Combining radiant and convective systems with thermal mass for a more

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comfortable home, ASHRAE transactions (2006): 253-268.

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[18] Olesen, Bjarne W., and K. C. Parsons, Introduction to thermal comfort standards and to the proposed new version of EN ISO 7730, Energy and buildings 34.6 (2002): 537-548.

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[19] American Society of Heating, Refrigerating and Air-Conditioning Engineers. 62.1 User's

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Manual: ANSI/AHRAE Standard 62.1-2004, Ventilation for Acceptable Indoor Air Quality. American Society of Heating, Refrigerating and Air-Conditioning Engineers, 2004.

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[20] Hirsch, Charles, Numerical computation of internal and external flows, Computational

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methods for inviscid and viscous flows 2 (1990).

[21] Tritton, David J. , Physical fluid dynamics, Oxford, Clarendon Press, 1988, 536 p. 1 (1988).

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1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24 25 26 27 28 29 30 31 32 33 34 35 36 37 38 39 40 41 42 43 44 45 46 47 48 49 50 51 52 53 54 55 56 57 58 59 60 61 62 63 64 65

[22] Cumber, P. S., Improvements to the discrete transfer method of calculating radiative heat transfer, International Journal of Heat and Mass Transfer 38.12 (1995): 2251-2258. [23] Wang, Suya, et al., Evaluating the low exergy of chilled water in a radiant cooling system, Energy and Buildings 40.10 (2008): 1856-1865. [24] Incropera, Frank P., Adrienne S. Lavine, and David P. DeWitt. Fundamentals of heat and mass transfer. John Wiley & Sons Incorporated, 2011. [25] Ansys, C. F. X. 11.0 User's Guide, Ansys CFX-Sol. ver Theory Guide, Ansys Inc (2007).

21 Page 21 of 39

[26] Varma, Mahesh N., and A. Kannan, CFD studies on natural convective heating of canned food in conical and cylindrical containers, Journal of Food Engineering 77.4 (2006): 1024-1036.

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[27] Kong, Qiongxiang, and Bingfeng Yu, Numerical study on temperature stratification in a room with underfloor air distribution system, Energy and Buildings 40.4 (2008): 495-502.

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[28] Causone, Francesco, et al., Floor heating and cooling combined with displacement

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ventilation: Possibilities and limitations, Energy and Buildings 42.12 (2010): 2338-2352. [29] Henze, Gregor P., et al., Primary energy and comfort performance of ventilation assisted

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thermo-active building systems in continental climates, Energy and Buildings 40.2 (2008): 99-

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111.

[30] Shaw, M. R., K. W. Treadaway, and S. T. P. Willis, Effective use of building mass,

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Renewable energy 5.5 (1994): 1028-1038.

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22 Page 22 of 39

YTC_Figure List.docx

Figure List Fig. 1. Diagram of the proposed radiant system embedded a concentric tube heat exchanger

Fig. 3. System Configuration of the typical hydronic radiant system with DOAS

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Fig. 4. Building geometry mesh condition for the numerical (CFD) evaluation

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Fig. 2. System configuration of HHSR system

Fig. 5. Temperature distributions of fluids in the concentric tube heat exchanger in the plain slab in heating mode

Fig. 7. Slab surface temperature distributions for heating mode

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Fig. 8. Slab surface temperature distributions for cooling mode

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Fig. 6. Temperature distributions of fluids in the concentric tube heat exchanger in the plain slab in cooling mode

Fig. 9. Vertical room air temperature distributions for heating mode

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Fig. 10. Vertical room air temperature distributions for cooling mode

Fig. 11. Plant load estimation for DOAS+RAD and HHSR during heating and cooling mode

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Fig. 12. Comparison of averaged room air and slab temperature profiles in night ventilation.

Page 23 of 39

YTC_Tables.docx

Tables Table 1. Thermal properties of materials Component

Properties

Slab (concerte)

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Thickness : 0.06 m  =2240

k = 1.4 Thickness : 0.002 m  = 95

k = 0.042

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PolyEthylene cross(X) linked tube(PEX)

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c p = 880

c p = 1550 m  0.004125 Air

c p = 1004.4

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m  0.06 Water

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c p = 4187.7

Table 2. Design conditions of the numerical model

Cooling

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Design Condition

Outside Air Temp. Room Air Temp. Outside Air Temp.

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Heating

Room Air Temp.

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Dimensions

Room Condition

Slab and tube configuration

-17.3 °C 22 °C 31.5 °C 24 °C 2 m * 2 m * 2.8m (D*W*H)

Volume

11.2 m3

Exterior Wall area

4.25 m2

Exterior Wall U value

0.3 W/ m2K

Exterior Fenestration Area

1.05 m2

Exterior Fenestration U-Value

1.8 W/ m2K

Infiltration rate Ventilation rate Slab Thickness

0.5 ACH 1.0 ACH 0.06 m

Tube Spacing(center-center)

0.3 m

Tube length

13.2 m

Tube Material

Polyethylene cross linked(PEX) tube

Tube diameter (Water/Air)

0.0001767 m2

Page 24 of 39

Table 3. PMV and PPD value of each case

Heating Mode

DOAS+R AD HHSR DOAS+R AD

PMV [-] -0.3 -0.1 -0.5 -0.4 0 -0.1 0 0

PPD [%] 6.9 5.2 10.2 8.3 5 5.2 5 5

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Cooling Mode

HHSR

C.D. S.D. C.D. S.D. C.D. S.D. C.D. S.D.

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Type

Page 25 of 39

ed

Space

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i

Figure1

Ac

ce pt

Radiant Slab

Air Water

Fluid Direction Heat Exchange Page 26 of 39

M an

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cr

i

Figure2

Air Tube

ce pt

ed

Water Tube

Fan Box

Plant

Filter

Ac

OA

Storage Tank (Opt.)

Window

Radiant Panel

Enlarged

Room Air Outlet

Circulation Pump

Page 27 of 39

Dedicated Outdoor Air system

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Figure3

Room Air Outlet

ce pt

Radiant Panel

Ac

Plant

Window

ed

Required Filter Outside Air

Circulation Pump

Page 28 of 39

M an

Room Air outlet

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Figure4

ed

Wall

Infiltration inlet

ce pt

Window

Water(Air) Outlet

Ac

Ventilation Inlet

Water(Air) Inlet

Page 29 of 39

34

32

32

30

30

us

34

28

Fluid temperature [Deg.C]

26 24 22 20

Air (Adiabatic) Water (Adiabatic)

10 5 0 -5

26 24 22

M an

Fluid temperature [Deg.C]

28

20

Air Water

10 5 0

-5

-10

-10

-15

-20 -2

0

2

4

6

8

10

Heat exchanger length [m]

12

14

32 30 28 26

Ac

24 22 20 10

-2

0

2

4

6

8

10

12

14

16

14

16

Heat exchanger length [m]

(b) qh=100 W/m2 34 32 30 28

Fluid temperature [Deg.C]

34

-20

16

ce pt

(a) qh=0 W/m2 (adiabatic)

ed

-15

Fluid temperature [Deg.C]

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i

Figure5

Air Water

5 0 -5

26 24 22 20 10 5 0 -5

-10

-10

-15

-15

-20

Air Water

-20 -2

0

2

4

6

8

10

Heat exchanger length [m]

(c) qh=300 W/m2

12

14

16

-2

0

2

4

6

8

10

12

Heat exchanger length [m]

(d) qh=500 W/m2

Page 30 of 39

30

30

Air (Adiabatic) Water (Adiabatic)

22 21

28

Air Water

26 22 21 20

19 -2

0

2

4

6

8

10

ed

20

12

Heat exchanger length [m]

14

30

Ac

28 26 22

-2

0

2

4

6

8

10

12

14

16

Heat exchanger length [m]

(b) qc=100 W/m2 32 30

Fluid temperature [Deg.C]

32

19

16

ce pt

(a) qc=0 W/m2 (adiabatic)

Fluid temperature [Deg.C]

Fluid temperature [Deg.C]

28 26

cr

32

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32

M an

Fluid temperature [Deg.C]

i

Figure6

Air Water

21 20

28 Air Water

26 22 21 20

19

19 -2

0

2

4

6

8

10

Heat exchanger length [m]

(c) qc=300

W/m2

12

14

16

-2

0

2

4

6

8

10

12

14

16

Heat exchanger length [m]

(d) qc=500 W/m2

Page 31 of 39

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HHSR

ed

M an

us

DOAS+RAD

i

Figure7

Ac

ce pt

DOAS+RAD

(a) Ceiling diffuser HHSR

Temp. [°C] 19.95 20.95

(b) Wall diffuser 21.95

22.95

23.95

24.95

25.95

26.95

27.95

28.95

Water Flow direction

Page 32 of 39

cr

HHSR

ed

M an

us

DOAS+RAD

i

Figure8

Ac

ce pt

DOAS+RAD

(a) Ceiling diffuser HHSR

Temp. [°C]

19.95

20.55

(b) Wall diffuser 21.15

21.65

22.25

22.75

23.35

23.85

24.45

24.95

Water Flow direction

Page 33 of 39

(a) Ceiling diffuser HHSR

Ac

ce pt

DOAS+RAD

ed

M an

us

cr

HHSR

DOAS+RAD

i

Figure9

(b) Wall diffuser

Temp. [°C]

14.95

15.75

16.55

17.35

18.15

18.65

19.65

20.45

21.25

21.95

Air flow direction

Page 34 of 39

ed

M an

us

cr

HHSR

DOAS+RAD

i

Figure10

(a) Ceiling diffuser HHSR

Ac

ce pt

DOAS+RAD

Temp. [°C]

23.95

24.55

(b) Wall diffuser 25.15

25.65

26.25

26.75

27.35

27.85

28.45

28.95

Air flow direction

Page 35 of 39

us

cr

i

Figure11

339.08

350

321.75

309.38

200

334.13

ce pt

150

100 50

0

HHSR

ed

250

Ac

Plant load [W]

300

DOAS+RAD

M an

400

Ceiling diffuser

106.43

103.95 84.15

Sidewall diffuser

Heating

84.15

Ceiling diffuser

Sidewall diffuser Cooling

Operation mode

Page 36 of 39

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i

Figure12

34 Mean Space Air (CNV)

32

Night ventilation

Mean Space Air (HHSR)

Mean Slab (HHSR)

ed

28

ce pt

26

24

22

20 20

21

Ac

Temperature [oC]

30

Mean Slab (CNV)

M an

Outdoor air

22

23

0

1

Daytime operation

2

3

4

5

6

7 8 9 Time [Hour]

10

11

12

13

14

15

16

17

18

19

Page 37 of 39

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Graphical Abstract (for review)

M an

Concentric tube heat exchanger

Air Tube

ce pt

ed

Water Tube

Fan Box

Plant

Filter

Ac

OA

Storage Tank (Opt.)

Window

Radiant Panel

Enlarged

Room Air Outlet

Circulation Pump

Page 38 of 39

*Highlights

Highlights We developed a new concept of radiant system using concentrate tube heat exchanger.



The system enables to satisfy room sensible load and ventilation requirement.



We evaluate the system performance using CFD model.



The system provides a more closed room design condition both heating and cooling mode.



It improves also the performance of thermal storage with night ventilation.

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

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