TECHNICAL FEATURE Fundamentals at Work This article was published in ASHRAE Journal, February 2015. Copyright 2015 ASHRAE. Posted at www.ashrae.org. This article may not be copied and/or distributed electronically or in paper form without permission of ASHRAE. For more information about ASHRAE Journal, visit www.ashrae.org. Part One Radiant Heating and Cooling Systems BY KWANG WOO KIM, ARCH.D., MEMBER ASHRAE; BJARNE W. OLESEN, PH.D., FELLOW ASHRAE The use of radiant heating systems has several thousand years of history. 1,2 The early stage of radiant system application was for heating purposes, where hot air from flue gas (cooking, fires) was circulated under floors or in walls. After the introduction of plastic piping water-based radiant heating and cooling with pipes embedded in room surfaces (floor, wall, and ceiling), the application increased significantly worldwide. Earlier application of radiant heating systems was mainly for residential buildings because of its comfort and free use of floor space without any obstruction from installations. For similar reasons, as well as possible peak load reduction and energy savings, radiant systems are being widely applied in commercial and industrial buildings. This two-part article describes basic knowledge of radiant heating and cooling systems to give a principle understanding of the design and operation of this advantageous system including comfort, system load, heating/cooling capacity, installation and application of the system with examples. Embedded Radiant Heating and Cooling System Embedded radiant systems are used in all types of buildings. Due to the large surfaces needed for heat transfer, the systems work with low water temperature for heating and high water temperature for cooling. The water temperatures are operated at very close to room temperature, and, depending on the position of the piping, the system can take advantage of the thermal storage capacity of the building structure. Figure 1 shows the available types of embedded hydronic radiant systems. The embedded radiant systems, except thermally active building systems (TABS), are usually insulated from the main building structure (floor, wall and ceiling), and the actual operation mode (heating/cooling) of the systems depends on the heat transfer between the water and the space. Kwang Woo Kim, Arch.D., is a professor of architecture at Seoul National University, Seoul, South Korea, and president of Architectural Institute of Korea. Bjarne W. Olesen, Ph.D., is director, professor, International Centre for Indoor Environment and Energy, Technical University of Denmark in Lyngby, Denmark, and vice president of ASHRAE. 28 ASHRAE JOURNAL ashrae.org FEBRUARY 2015
TECHNICAL FEATURE Floor Ceiling Wall TABS FIGURE 1 Examples of water based radiant systems. 3 The radiant system is defined as a system where at least 50% of the heat transfer takes place by radiation. Figure 2 shows the total heat transfer coefficients between a heated-cooled surface and a room. The radiant heat transfer is, in all cases, 5.5 W/m² K (0.97 Btu/h ft² F). The convective heat transfer then varies between 0.5 and 5.5 W/m² K (0.09 and 0.97 Btu/h ft² F), depending on the surface type and on heating or cooling mode. This shows that the radiant heat transfer varies between 50% and 90% of the total heat transfer. The heat transfer coefficient for cold ceiling and warm floor will vary between 9 and 11 W/m² K (1.59 and 1.94 Btu/h ft² F), depending on the temperature difference between surface and room. The radiant heat transfer does not directly affect the room air temperature. The long wave radiation heats or cools the surrounding surfaces, which then indirectly heats or cools the room air. Standard for Radiant Heating and Cooling Systems As the heat transfer process between water and room is quite different from conventional air systems, an international standard on radiant heating and cooling systems has been developed based on system design and existing standards from different countries and was published in 2012. ISO 11855, Building Environment Design Design, Dimensioning, Installation and Control of the Embedded Radiant Heating And Cooling Systems, 6 11 consists of six parts: Part 1: Definition, symbols, and comfort criteria; Part 2: Determination of the design and heating and cooling capacity; Part 3: Design and dimensioning; Part 4: Dimensioning and calculation of the dynamic heating and cooling capacity of thermo active building systems; Btu/h ft 2 F 2.03 1.85 1.94 1.67 1.50 1.32 1.06 1.14 0.97 1.23 Heating Cooling Floor FIGURE 2 Heat transfer coefficients between heated/cooled surface and room. 4,5 Part 5: Installation; and Part 6: Control. Ceiling Comfort Occupants thermal comfort is the primary objective in radiantly heated or cooled space. To provide an acceptable thermal environment for the occupants, the requirements for general thermal comfort shall be taken into account by using the index of predicted mean vote (PMV) or operative temperature, t o, and local thermal comfort, e.g., surface temperature, vertical air temperature differences, radiant temperature asymmetry, draft, etc. For radiant or convective systems the comfort requirements are the same when expressed by the PMV-PPD index ( 0.5 < PMV <+0.5) or expressed as an operative temperature range corresponding to: 20 C to 24 C (68 F to 75.2 F) for heating season and 23 C to 26 C (73.4 F to 78.8 F) for cooling season in spaces with sedentary activity. 12,13 The operative temperature 14,15 is the combined influence of air temperature and mean radiant temperature. The operative temperature can be approximated with 1.94 1.41 1.41 Wall FEBRUARY 2015 ashrae.org ASHRAE JOURNAL 29
TECHNICAL FEATURE Fundamentals at Work Y 80% 60% 40% 20% 0 F 3.6 F 7.2 F 10.8 F 14.4 F 18 F Y 80% 0 F 9 F 18 F 27 F 36 F 45 F 54 F 63 F 60% 40% 1 2 20% 10% 8% 6% 4% 2% 10% 8% 6% 4% 2% 3 4 1% X = Air Temperature Difference Between Head and Feet Y = Dissatisfied FIGURE 3 Local thermal discomfort caused by vertical air temperature difference. 6 X 1% X X = Radiant Temperature Asymmetry Y = Dissatisfied 1 = Warm Ceiling 2 = Cool Wall 3 = Cool Ceiling 4 = Warm Wall FIGURE 4 Local thermal discomfort caused by radiant temperature asymmetry. 6 the simple average of air and mean radiant temperature in spaces with low air velocities (<0.2 m/s [39 fpm]), or with a small difference between mean radiant temperature and air temperature (<4K, 7 F). The operative temperature (t o ) is in spaces with low air velocities determined from the following expression: t o = 0.5(t a + t r ) Where t a = air temperature F ( C) t r = mean radiant temperature F ( C) The occupants can maintain the same comfort level with a lower air temperature in a radiantly heated space, and the same comfort level with a higher air temperature in a radiantly cooled space in comparison to convective heating and cooling systems. Therefore, reduction of the energy loss due to ventilation and infiltration is possible while maintaining the same comfort level compared with conventional heating and cooling systems. As the reference temperature for the transmission heat loss is closer to the operative temperature than to the air temperature, there will not be any significant difference of transmission heat loss between radiantly heated or cooled spaces. Interestingly enough, the difference between air- and mean radiant temperature is normally smaller in radiantly heated or cooled spaces. This is due to the fact that in winter the windows will have a lower surface temperature than the air temperature, which is compensated by a higher surface temperature of the radiant Dissatisfied (%) 100 80 0.3 0.2 60 50.4 F 0.1 0.0 40 41.4 F 0.1 20 0.2 32.4 F 0.3 0 20 30 40 50 60 70 1.0 80 Relative Humidity (%) FIGURE 5 Human satisfaction with the IAQ depending on relative humidity and air temperature. 3 system and vice-versa in summer. With air systems the colder window temperatures in winter will be compensated by a higher air temperature, which will result in an air temperature higher than the mean radiant temperature. For rooms with sedentary and/or standing occupants, the maximum permissible floor temperature for heating is 29 C (84 F), and the minimum floor temperature for cooling is 19 C (66 F). For spaces with occupants in bare feet (bathrooms, swimming pools, dressing rooms, etc.), the optimal floor temperature for comfort also depends on the floor covering material. For wall heating, a maximum surface temperature range of 35 C to 45 C (95 F to 113 F) is recommended. The maximum may depend on whether the occupants 1 Acceptability 30 ASHRAE JOURNAL ashrae.org FEBRUARY 2015
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TECHNICAL FEATURE Fundamentals at Work 1 2 3 Type A and C 1 Type B 1 Type D Type G 2 2 5 Flooring Material 3 3 4 4 4 Joist 1 = Floor Covering 2 = Weight Bearing and Thermal Diffusion Layer (Cement Screed, Anhydrite Screed, Asphalt Screed or Wood) 3 = Thermal Insulation 4 = Structural Base 5 = Heat Diffusion Device FIGURE 6 Embedded radiant system types. 7 may easily get contact with the surface or whether occupants are more sensitive persons such as children or the elderly. Wall cooling is limited by the risk of condensation and the development of a downdraft of cold air. A vertical air temperature difference between head and feet of less than 3K (5.4 F) is recommended. Most heating and cooling systems will, in modern buildings, normally have vertical air temperature differences within this limit. In high ceiling spaces it is, for energy reasons, important to avoid large vertical temperature differences. This is why floor heating is especially recommended here (atrium, foyer, industrial space, etc.) People are very sensitive to radiant temperature asymmetry from a cold window and a warm ceiling. Occupants may feel discomfort caused by a temperature asymmetry of 5K (9 F) for warm ceiling, and a temperature asymmetry of 10K (18 F) for walls or windows (Figure 4, Page 30). The critical factor at cold surfaces (windows, walls) is, however, the risk of downdraft that may cause discomfort. The radiant heating and cooling system operates with less dust transportation, as it is not a convective system, and does not require the cleaning of heat emitters or filters. With the radiant floor heating systems, carpets are not necessary. Thus, the possible allergen sources of emitting pollutants and a sink source can be eliminated. The higher mean radiant temperature in radiantly heated space means that the air temperature can be kept lower than in convectively heated space. This has the advantage that the relative humidity in winter may be a little higher. Studies show that lower air temperature and lower air humidity have a significant effect on perceived air quality (Figure 5 3 ). Due to the higher heating surface temperatures, there is less chance for condensation and mold growth. The relationship between air temperature and humidity is one of important comfort issues in radiantly cooled spaces. Where the humidity is not controlled by the air system, as in naturally ventilated spaces, radiant cooling capacity will be limited to avoid the forming of condensation on the radiant surface (see section on control in Part 2 of this article in next month s Journal). With air heating or cooling system more air has to be circulated than the amount needed for providing acceptable air quality. This may increase the noise level in a space and also increase the risk for complaints related to draft. When a part of sensible heating and/ or cooling load is taken care of by a water-based radiant system, the ventilation system may have reduced duct size and lower air velocity because it will only treat the air renewal for required IAQ and, if needed, dehumidification. In buildings with thermally active building systems (TABS) you will normally prefer to have free access to the concrete surface to increase the heat transfer with the room. This may require special solutions for the acoustics. Acoustic panels on the ceilings and suspended ceiling panels will reduce heat transfer. It will be more efficient to hang down vertical acoustical panels. 16 The application of the raised floor or the thermal/acoustic insulation in floor will decrease the upper heat flow from the TABS, which normally is much less than the heat exchange from the ceiling. Load Calculations and Heating/Cooling Capacity At a given average surface temperature and indoor temperature (operative temperature, t o ), a surface will deliver the same amount of heat flux to a space 32 ASHRAE JOURNAL ashrae.org FEBRUARY 2015
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TECHNICAL FEATURE Fundamentals at Work Structure S4 q = 0 q = 0 Material Floor Covering λ = 0.08 Btu in/h ft 2 F s = 0.6 in. Screed λ = 0.40 Btu in/h ft 2 F s = 2.4 in. Thermal Insulation λ = 0.01 Btu in/h ft 2 F s = 1.2 in. Concrete λ = 0.70 Btu in/h ft 2 F s = 7.1 in. q up =112.6 Btu/h Temperature ( F) 78.8 77 75.2 73.4 71.6 69.8 68 66.2 64.4 q down =22.3 Btu/h FIGURE 7 Temperature distribution and cooling effect up and down for a floor system calculated by FEM software for a floor cooling system with 19 C (66.2 F) water temperatures and 26 C (78.8 F) room temperature. regardless of the embedded radiant system type. Therefore, it is possible to establish basic formulas or characteristic curves of heating and cooling for the all heating and cooling surfaces independent of the embedded system types. The heat transfer between the surface and the space do, however, depend on the different surface heat transfer coefficients (Figure 2). The heat transfer between the water and surface is different for each system configuration. Therefore, the estimation of heating/cooling capacity of systems is very important for the proper system design. Two calculation methods included in ISO 11855-2 7 are simplified calculation methods depending on the type of system, and finite element method (FEM) or finite difference method (FDM). Given system types are Types A and C, Type B, Types D and G. The simplified calculation methods are specific for the given system types within boundary conditions. Based on the calculated average surface temperature at given heat transfer medium temperature and space temperature, it is possible to determine the steady state heating and cooling capacity. In case a simplified calculation method is not applicable for the considered system, either two- or three-dimensional finite element or finite difference method, or laboratory testing may be applied. The temperature distribution in floor cooling system, calculated using FEM software, is shown in Figure 7. Heat exchange coefficient is the parameter that determines the amount of heat transferred between surface and the space in relation with the system type. Acceptable surface temperature is determined based on comfort considerations and the risk of condensation. Heating/cooling capacity of the systems is: 7 Floor heating and ceiling cooling, q = 8.92 (t o t S,m ) 1.1 ; Wall heating and wall cooling, q = 8 ( t o t S,m ); Ceiling heating, q = 6 ( t o t S,m ); Floor cooling, q = 7 ( t o t S,m ). Where t o ( C) is the operative temperature in the space t S,m ( C) is the average surface temperature The ceiling has the capacity up to 100 W/m² (31.7 Btu/h ft²) for sensible cooling and 40 to 50 W/m² (12.7 to 15.9 Btu/h ft²) for heating. The floor has the capacity up to 100 W/m² (31.7 Btu/h ft²) for heating and 40 W/m² (12.7 Btu/h ft²) for sensible cooling. When direct sunlight strikes on the floor, the sensible cooling capacity of the floor may be more than 100 W/m² (31.7 Btu/h ft²). This is why floor cooling is often adopted in spaces with large window area like airports, atria and lobby halls. For the thermally active building systems (TABS), the steady-state heating/cooling capacity calculation is not sufficient, and analysis with a dynamic computational program that can predict the dynamic behavior and performance of the system together with the building 34 ASHRAE JOURNAL ashrae.org FEBRUARY 2015
TECHNICAL FEATURE is needed. 17,1 Several programs exist such Energy Plus, TRNSYS and IDA-ICE. One of the main advantages of TABS are reduced building height. For each story, you may save 500 to 600 mm (1.8 to 2 ft) of building height, which for a seven-story building amounts to an entire story and related building materials. As no suspended ceiling is needed to cover air ducts, significant saving of building materials is possible. It is also possible to operate the system at 30% to 50% lowered peak loads allowing reduced plants sizes and possible operation of heating/cooling systems with temperatures close to room temperature, allowing increased plants efficiency and use of renewable energy sources (ground heat exchanger, evaporative cooling, etc.). Thermally active building systems exploit the high thermal inertia of the slab to perform peak shaving. The peak shaving reduces the peak in the required cooling power, 7 so that it is possible to cool the structures of the building during a period in which the occupants are absent (during nighttime in office premises). This way, the cooling can be delayed and lower nighttime electricity rates can be used. At the same time, a reduction in the size of heating/cooling system components (including the chiller) is possible. During daytime, the heat is extracted from the occupied space by the ventilation system and stored in the concrete slabs. Then, during nighttime, the level of ventilation is reduced and the circulation of cool water in the slabs will remove the stored heat. For the conventional air system, the space load will be the instantaneous system load, because all the heat delivered to the space is immediately removed by the air system. For the radiant system and especially for a TABS the calculated design space load should not be used as system load. For both an air system and a TABS, it is important that the room load over a 24-hour day (Curve 1 in Figure 8) is removed within the 24 hours, else the room will get warmer and warmer day by day if the weather stays the same. The difference is that with a TABS this load is removed from the space in three ways: absorption in the concrete slab, removed by the ventilation system and removed by the water circulating in the slabs. Therefore, the load is removed by a slab system 4 1 2 1 = Heat Gain 2 = Power Needed for Conditioning the Ventilation Air 3 = Power Needed on Water Side 4 = Peak Heat Gain Reduction FIGURE 8 Example of peak-shaving (reducing the peak load) effect (time vs. cooling power). 17 Cooling Power (Btu/h) over 24 hours and by an air system normally over 8 to 10 hours. After the water circulation in the slab may have been stopped during the day and will be started again in the evening, there will be a high peak cooling load between the heated slab and the cool water; but this should not be used to size the chiller as it is a very short peak and the capacity needed will after some minutes decrease significantly. It can be somewhat complicated to calculate the needed capacity on the water side (chiller, heat pump); therefore, a dynamic building simulation is recommended. System Design Radiant system design requires determining heating/cooling surface area, type, pipe size, pipe spacing, supply temperature of the heat transfer medium, and design medium flow rate. The design steps are as follows (ISO 11855-3 8 ): 1. Calculate the design heating and sensible cooling load in accordance with a standard for heating and cooling load calculation based on operative temperature. 2. Determine the minimum supply air quantity needed for ventilation and dehumidification. In cooling application, calculate latent cooling and sensible cooling available from supply air. Determine remaining sensible cooling load to be satisfied by radiant system. Also, designate or calculate the relative humidity and dew point, because the cooling system should operate within a surface temperature range above the dew point, which shall be specified depending on the respective climate conditions in the country. By limiting supply water 3 FEBRUARY 2015 ashrae.org ASHRAE JOURNAL 35
TECHNICAL FEATURE Fundamentals at Work temperature to be maintained above dew point, the risk of condensation can be easily avoided. 3. Determine the surface area for radiant system, excluding any area covered by objects immovable or fixed to the building structure. 4. Establish a maximum permissible surface temperature and a minimum permissible surface temperature in consideration of the dew point. 5. Determine the design heat flux, including the design heat flux of peripheral area and the design heat flux of occupied area. For the design of the cooling systems, determine the room with the maximum design heat flux. 6. Determine the radiant system such as the pipe spacing and the covering type, and design heating and cooling medium differential temperature based on the maximum design heat flux and the maximum and minimum surface temperature from the field of characteristic curves. 7. If the design heat flux cannot be obtained by any pipe spacing alternatives for the room of design, it is recommended to provide supplementary heating/cooling equipment. In this case, the maximum design heat flux for the embedded system may now occur in another room. 8. Determine the thermal resistance of backside insulating layer and the design heating/cooling medium flow rate. 9. Estimate the total length of circuit. Hydronic radiant surface systems are very often coupled with an airhandling system. The air-handling system usually operates only with the amount of air needed for acceptable indoor air quality, the required IAQ standard, or amount of air needed to remove latent heat from TABLE 1 System design example for panel cooling. STEP FIND EXAMPLE 0 A = 10 m 2 (108 ft 2 ), V = 30 m 3 (1,059 ft 3 ) Room is Given to be Installed with Radiant System and HVAC System 1 Calculate Cooling Load Based on Operative Temperature Cooling Load (Latent) Cooling Load (Sensible) 2 Determine Minimum Supply Air Quantity Calculate Latent Cooling Available From Supply Air Sensible Cooling Available From Supply Air Design the Relative Humidity And Dew Point Determine Remaining Sensible Cooling Load to be Satisfied by Radiant System Q c,latent = 150 W (512 Btu/h) Q c,sensible = 1,000 W (3,416 Btu/h) V HVAC,min = 0.7 ACH = 21 CMH (12.4 CFM) Q HVAC,latent = Q c,latent = 150 W (512 Btu/h) Assuming the SHF (Sensible Heat Factor) of HVAC, SHF = (Q HVAC,sensible / Q HVAC,total ) = 0.7 Q HVAC,total = Q HVAC,latent / (1 - SHF) = Q HVAC,latent / 0.7 = 500 W (1,708 Btu/h) Then, Q HVAC,sensible = Q HVAC,total Q HVAC,latent = 500 W 150 W = 350 W (1,196 Btu/h) is available from supply air of HVAC RH = 50 %, T dew = 14.8 C (58.6 F) Remaining Q c,sensible = Q c,sensible - Q HVAC,sensible = 1,000W 350 W = 650 W (2,220 Btu/h) 3 Determine the Available Surface Area A available = 5 m 2 (53.8 ft 2 ), 50% of ceiling area is available for radiant system 4 Establish a Minimum Permissible Surface Temperature T surf,min = 17 C (62.6 F) is acceptable for cooled ceiling * (which is higher than dew point temperature) 5 Determine Maximum Design Heat Flux Q c,max = 99 W/m 2 (31.4 Btu/h ft 2 ) is allowed for cooled ceiling * 6 Determine Radiant System Pipe Spacing Covering Type Design Cooling Medium Differential Temperature Design heating capacity of radiant system, 7 Select Supplementary Cooling Equipment 8 Determine Thermal Resistance of Backside Insulating Layer Cooling Medium Flow Rate Selected Radiant System has cooling capacity of 80 W/m 2 (25.4 Btu/h ft 2 ). m T = 0.2 m (8 in.) punched aluminum sheet T m = 2 C (3.6 F) 5 m 2 = 400 W (1,366 Btu/h) Required cooling capacity of Supplementary Cooling Equipment Q out = Remaining Q c,sensible Q des = 650 W 400 W = 250 W (854 Btu/h) R cover = 0.021 m 2 K/W (0.12 h ft 2 F/Btu) m = 0.0478 kg/s (6.3 lb/min), ensured of fully developed flow in pipe If the resistance of backside insulation is high, the cooling medium flow rate could be lowered. 9 Estimate Total Length Of Circuit L cir = A available / m T = 5 m 2 / 0.2 m = 25 m (82 ft) * ISO 11855-2: Building environment design Design, dimensioning, installation and control of embedded radiant heating and cooling systems Part 2: Determination of the design heating and cooling capacity. 36 ASHRAE JOURNAL ashrae.org FEBRUARY 2015
TECHNICAL FEATURE the space and control air humidity level, while the hydronic system supplies or removes the sensible heat depending on the seasonal conditions. In the cooling mode, the air system can play a key role in avoiding surface condensation. Part 2 of this article will cover control, operation, installation and application of the system. Acknowledgments This article was supported by VELUX guest professorship, and a grant from the National Research Foundation of Korea (NRF) funded by the Korean government (MEST) (No. 2014-050381). References 1. Bean, R., Olesen, B.W., Kim, K. W. 2010. History of Radiant Heating & Cooling Systems, Part 1. ASHRAE Journal (1):40 46. 2. Bean, R., Olesen, B.W., Kim, K. W. 2010. History of Radiant Heating & Cooling Systems, Part 2. ASHRAE Journal (2):50 55. 3. REHVA. 2007. Guidebook No 7: Low Temperature Heating and High Temperature Cooling. 4. Olesen, B.W. 1997. Possibilities and limitations of radiant floor cooling. ASHRAE Transactions 103(1):42 48. 5. Olesen, B.W., Michel, E., Bonnefoi, F., De Carli, M. 2000. Heat exchange coefficient between floor surface and space by floor cooling: theory or a question of definition. ASHRAE Transactions, Part I. 6. ISO 11855-1:2012, Building environment design - Design, dimensioning, installation and control of the embedded radiant heating and cooling systems Part 1: Definition, symbols, and comfort criteria. 7. ISO 11855-2:2012, Building environment design - Design, dimensioning, installation and control of the embedded radiant heating and cooling systems Part 2: Determination of the design and heating and cooling capacity. 8. 8. ISO 11855-3:2012, Building environment design - Design, dimensioning, installation and control of the embedded radiant heating and cooling systems Part 3: Design and dimensioning. 9. ISO 11855-4:2012, Building environment design - Design, dimensioning, installation and control of the embedded radiant heating and cooling systems Part 4: Dimensioning and calculation of the dynamic heating and cooling capacity of Thermo Active Building Systems (TABS). 10. ISO 11855-5:2012, Building environment design - Design, dimensioning, installation and control of the embedded radiant heating and cooling systems Part 5: Installation. 11. ISO 11855-6:2012, Building environment design - Design, dimensioning, installation and control of the embedded radiant heating and cooling systems Part 6: Control. 12. ASHRAE Standard 55-2010, Thermal Environmental Conditions for Human Occupancy. 13. ISO EN 7730-2005, Moderate thermal environments determination of the PMV and PPD indices and specification of the conditions for thermal comfort. 14. 2012 ASHRAE Handbook HVAC Systems and Equipment. 15. ISO EN 7726-1998, Ergonomics of the thermal environment- Instruments for measuring physical quantities. 16. Weitzmann, P., Pittarello, E., Olesen, B.W. 2008. The cooling capacity of the thermo active building system combined with acoustic ceiling. Presented at Nordic Symposium on Building Physics. 17. Olesen, B.W. 2012. Thermo active building systems using building mass to heat and cool. ASHRAE Journal 54(2):44 52. Advertisement formerly in this space. Advertisement formerly in this space. FEBRUARY 2015 ashrae.org ASHRAE JOURNAL 37