Numerical comparison between energy and comfort performances of radiant heating and cooling systems versus air systems.
Introduction and purpose of the work
Hydronic radiant panels represent a successful solution to be adopted if you want to achieve high thermal comfort levels and significant energy savings simultaneously. At the beginning, applications were for heating purposes, expecially with radiation floor techonologies; later, the use of radiant systems has been extended also to cooling purposes, and different solutions of radiant cooling floors and ceilings have been developed.
Nowadays, hydronic radiant panels are a consolidated climatization technology; they are widely used in Europe, and in North America, the number of applications is continuously increasing.
As it is well known, the principles of radiant systems were already adopted in ancient times (Bean et al. 2010a), but only at the beginning of 1950s were they widely rediscovered and applied as building mechanical heating systems.
Mistakes, both in the design procedure and in the installation phase, have been done in some of the first applications (Feustel and Stetiu 1995; Bean et al. 2010b). These mistakes have generated a preliminary obstacle for the diffusion on the market of the radiant technique; on the other hand, they have stimulated the development of specific research addressed to the characterization of heat exchange mechanisms between the heating/cooling radiant surface and the surrounding indoor environment.
Moreover, progress on the material science has allowed the adoption of safer and more flexible solutions aimed at providing easier installations, with plastic (polyethylene and polypropylene) pipes, modular radiant panels, etc. (Babiak et al. 2007) being used.
At the same time, significant improvements have been carried out in the "theory of radiant climatization." Dedicated design methods have been defined (Causone et al. 2010a) and dynamic energy simulation software tools have been developed or adapted to take into account the peculiarities of radiant systems (Strand and Pedersen 2002).
The high comfort level provided in the indoor environment by radiant systems is a recognized successful feature (ASHARE 2007, 2008). In terms of thermal comfort, this is achieved by the preminent use of radiative heat exchanges between panel and occupants/walls (ASHARE 2007, 2008; Watson and Chapman 2002). To assure indoor air quality, radiant systems are typically coupled with a dedicated outdoor air system (DOAS) (Causone and Corgnati, 2011; Jeong et al. 2003). A DOAS is an important element in radiant panel design. In fact, the ventilation system plays a fundamental role, because, in addition to guarantee the desired ventilation airflow rate, it keeps the humidity under control, avoiding condensation problems on the cooled surface of the panel (Conroy and Mumma 2001). Radiant panels can be used as the unique climatization system only when their heating/cooling capacity is higher than the total heating/cooling load. As their efficiency is limited by the maximum/minimum allowed surface temperature, when loads overcome the panel efficiency, the supplied ventilation air is coupled, especially in cooling mode (Jeong and Mumma 2006).
As mentioned above, the great interest in radiant systems is due to the capability of maintaining high thermal comfort levels by using a fluid at a moderate temperature (low temperature heating and high temperature cooling; Babiak et al. 2007). This allows the exploitation of greater conversion efficiencies of primary energy systems (condensing boilers, heat pumps, liquid chilling packages, etc.). However, the actual energy savings depend strictly on the careful design and control of the radiant system, on the selection and sizing of the primary systems, on the climate, and on the other influencing boundary conditions (Olesen and Mattarolo 2009).
The main aim of this work is to is to provide an assessment of the benefits that can be achieved by a radiant system in heating and cooling mode in terms of carbon dioxide emissions and energy savings, expressed in terms of delivered energy and source energy. This is done by means of a comparison between a system based on a radiant floor/ceiling for heating and cooling and a reference conventional all-air system, analyzed in various European climates. The performance of the systems is simulated by means of a building energy simulation program.
Problem definition and discussion
Comparing the energy consumed by radiant and all-air systems for a particular building may appear quite simple from a procedural point of view, but upon deeper investigation, some problems arise affecting energy consumption assessment. The main aspects are summarized as follows:
* thermal comfort depends on air temperature and on mean radiant temperature (combined into the operative temperature), while the control of the systems is usually carried out by air temperature (Olesen 2001,2002a; Simone 2009; Berglund and Berglund 1994);
* radiant systems are not only HVAC components but also envelope components affecting the energy balance;
* radiant systems contain both radiant and convective equipment, while the HVAC reference system is only convective equipment (e.g., fan coil, variable air volume [VAV] damper and inlet, etc.) (Causone et al. 2010b; Spitler 2009); and
* the controlling equipment thermal inertia (radiant, all air) varies, while the controlled system thermal inertia (the indoor thermal environment) is constant (Olesen 2002a; Babiak et al. 2007).
Regarding the first aspect, for the same air temperature, the thermal comfort conditions guaranteed by a radiant system are better than those of a convective system, since the operative temperature is higher (or lower in summer conditions) than the one of an enclosure equipped with a convective system (Olesen 2002b). The operative temperature can be considered as a basic indicator of thermal comfort in rooms with moderate values of relative humidity and air velocity (ASHRAE 2009a). It is expressed as the weighted average of air temperature and of mean radiant temperature; it depends on both convective and radiant room heat loads. As shown in previous works by Simmonds (1994, 1996) the mean radiant temperature substantially affects thermal comfort conditions; therefore, it is not effective to contrast radiant loads with convective loads. Radiant systems contrast radiant loads with radiant loads, while air systems contrast radiant loads with convective loads; as a result, with air systems, a lower thermal comfort is obtained with the same room reference air temperature (ASHARE 2007, 2008; Watson and Chapman 2002). A correct comparison can be made only if the same operative temperature, or better thermal comfort conditions, are guaranteed in both cases (see the "Thermal Comfort Analysis" section under "Results"). To face this problem, in this numerical study, a thermostatic control based on operative temperature was selected. In the EnergyPlus simulation tool, an object called "ZoneControl:Thermostat:OperativeTemperature" was used. It allows the performance of a control based on the thermostatic operative temperature, which is a quantity calculated by the software as the average between the mean radiant temperature and the air temperature of the zone. It is thus possible to perform energy simulations assuming a thermostatic control based on the calculated operative temperature.
The second aspect is easily neutralized through an enclosure with the same thermal insulation for both systems (this is not obvious for various reasons; the radiant systems always require an insulation layer that may not be present in the reference building; the insulation is placed also in internal constructions; and the radiant system, even if well insulated, may loose heat toward the external side of the building when it represents an envelope component).
The third aspect deals with how the HVAC equipment are taken into account in heat balance calculation. Calculation methods were developed to integrate into the air heat balance method (which is the heat balance solution algorithm used by EnergyPlus and applied in the present work) a radiant heating and cooling model (Strand and Pedersen 2002); the radiant system affects the surface temperature of the floor/ceiling/wall, and it enters the radiative and the convective heat balance of the enclosure. The problem of the convective/radiative nature of the air-conditioning equipment is more important for heating and cooling load calculations than for energy requirement calculations (ASHRAE 2009b).
Regarding the last aspect, different energy requirements of the radiant system and of the reference system may be obtained as a result of equipment control and thermal inertia. This aspect is particularly important with low loads and in mid seasons, when mild and cold days may alternate, and equipment with a low thermal inertia may assure a better performance than high inertia equipments. To overcome this drawback, a suitable control strategy should be selected for the radiant system (Olesen 2002a, 2002b). Practical examples of control strategies for both heating and cooling by means of a floor radiant heating system are discussed by Simmonds (1994), while Leigh and MacCluer (1994) provided a comparative study of varius approaches for radiant floor heating system control.
Within this picture, the scope of this work is not only to carry out a comparison between the energy and environmental performances of radiant systems and a reference system, but also to provide a procedure to be followed when this comparison is performed.
The energy models of both buildings, one case study equipped with the reference system and the other equipped with the radiant system and a DOAS, were built and simulated into the energy simulation software EnergyPlus v. 3.1.0. The software is based on the air heat balance method to determine the loads of each building thermal zone and can model the most common systems and equipments.
The modeling of the low-temperature radiant heating/high-temperature radiant cooling system is performed in EnergyPlus by means of an object modeling a circuit of hot/cold fluid through pipes embedded in a floor, ceiling, or wall. The embedded heat source or sink is included into the one-dimensional heat transfer, which is calculated through the conduction transfer function method. The full description about the radiant equipment simulation can be found in Strand and Pedersen (2002) and the engineering reference of the EnergyPlus manual.
Typical weather conditions of the Rome, Milan, London, Frankfurt, Madrid, Athens, Helsinki, and Moscow locations refer to the IWEC (International Weather for Energy Calculation) database of climatic data.
The building benchmark model
One of the U.S. Department of Energy (DOE) new construction commercial building benchmarks was adopted as a reference building. The selected building is the small-sized office: a one-floor rectangular office building, with a total floor area of 511 mz (5500 ft2). It is partitioned and modeled into five conditioned zones (height 2.74 m [8.99 ft]): one central zone and four perimeter zones; moreover, there is an unconditioned attic. The internal gains, the building schedules of occupancy, electric equipment and electric lighting, and the building envelope characteristics are those of the benchmark model.
As far as building envelope characteristics, the thermal resistance of the walls, roofs, and windows is doubled when Helsinki and Moscow locations are simulated.
Reference system #1: All-air system
The reference air-conditioning system is a multizone VAV system with reheat coils in each thermal zone (Figure la). In heating mode, in each conditioned zone, a VAV damper and reheat box adjusts the flow rate and temperature of the air supplied as a consequence of the heating load. In cooling mode, the air is provided at the central air loop at 13[degrees]C (55.4[degrees]F) and then distributed into the zones; the VAV dampers adjust the flow rate. The thermostat set-point is a dual set-point with a dead band based on the values of 21[degrees]C (69.8[degrees]F) (with a set back at 15.6[degrees]C [60[degrees]F]) for heating and 25.5[degrees]C (77.9[degrees]F) (with a set back at 30[degrees]C [86[degrees]F]) for cooling.
A thermostatic control based on operative temperature was selected in order to perform the comparison between the air and radiant systems. The thermostat object controls the flow rate of the radiant system on the basis of a calculated operative temperature. This temperature is an average between air temperature and mean radiant temperature of the zone.
In addition to the thermostat, a humidistat controls the relative humidity of the core zone, providing humidification (by means of an electric steam humidifier placed in the main air loop) when relative humidity falls below 45% and dehumidification (by means of the cooling coil placed on the main air loop) when the relative humidity increases above 60%. As a consequence, the relative humidity of the zone is always between 45% and 60%.
There are two water loops, one for the hot water and one for the chilled water, set at the design temperature of 80[degrees]C and 7[degrees]C (176[degrees]F and 44.6[degrees]F), respectively. These temperatures were selected in order to limit the sizes of the heating and cooling coils and heat exchangers. The loops are fed, respectively, by a condensing boiler and an air-cooled vapor compression chiller equipped with a reciprocating compressor. The characteristics of the primary system equipment are summarized in Table 1. In this table, the rated efficiencies are listed, while the actual efficiencies depend on the part load and operation conditions (fluid temperatures, etc.) at each time step. The design air and water loops flow rates, the sizes of the coils, fans, and the other primary and secondary equipment are autosized by the simulation program EnergyPlus.
Reference system #2: Fan coil system
The performance of a fan coil reference system (Figure lb) was also compared to the performance of the radiant systems. Each conditioned zone has a forced air unit that is a four-pipe fan coil modeled separately by means of a heating coil, a cooling coil, a constant volume fan, and an outdoor air mixer. The outdoor airflow rate is set equal to the fresh air ventilation requirement, which is fixed in the benchmark specifications and equal in all the various systems that were simulated.
The thermostat set-point is, as in the previous case, a dual set-point with a dead band based on the values of 21[degrees]C (69.8[degrees]F) (with a set back at 15.6[degrees]C [60[degrees]F]) for heating and 25.5[degrees]C (77.9[degrees]F) (with a set back at 30[degrees]C [86[degrees]F]) for cooling.
Also in this case, a thermostatic control based on the operative temperature was selected in order to perform the comparison between the fan coil system and the radiant systems.
In a fan coil system, there is no possibility to perform a air humidification, so the indoor air relative humidity is not controlled during the heating season. During the cooling season, even if the indoor air relative humidity is not mechanically controlled (there is not a humidistat that senses the indoor conditions and activates the system response), the air, which is a mixture of outdoor air and recirculated indoor air, is partly dehumidified while passing through the cooling coil of the fan coil.
[FIGURE 1 OMITTED]
There are two water loops, one for the hot water, which serves the various heating coils of the forced-air units, and one for the chilled water, which serves the various cooling coils of the forced-air units. The hot water is supplied to the heating coils and to the cooling coils, respectively, at 80[degrees]C and 7[degrees]C (176[degrees]F and 44.6[degrees]F). As usual, these temperatures were selected in order to limit the sizes of the heating and cooling coils.
Similarly to the plant of the all-air system, the two loops are fed, respectively, by a condensing boiler and an air-cooled vapor compression chiller equipped with a reciprocating compressor. The characteristics of the primary system equipment are the same as in Table 1. The mean seasonal efficiency of these converters may vary as a consequence of the actual working conditions in the fan coils system. The water loops flow rates, the sizes of the coils, and fans are autosized by the simulation program EnergyPlus.
Radiant floor/ceiling system
In this case, the building is equipped with a radiant heating and cooling floor or ceiling (both cases are analyzed) system, coupled with an air system that is used only for ventilation and humidification/dehumidification purposes (Figure 1 c). The two zone equipment works together to condition the air of each zone, and the priority is assigned to the hydronic radiant system. The thermostat set-point schedules and control are the same as the VAV case.
The radiant floor is situated above an existing slab floor. It is placed over an insulation layer of 20 cm (7.87 in.). The hydronic tubing diameter is 17 mm (0.669 in.) and is contained in a gypsum mortar layer of 3.5 cm (1.38 in.). The floor finishing is a ceramic tile.
The radiant ceiling is a closed ceiling of 23 mm (0.906 in.) depth, made of a graphite layer between a steel sheet (toward the conditioned zone) and a wood plate. The hydronic tubing is made of PE and has a diameter of 10 mm (0.394 in.).
Both floor and ceiling are variable-flow low-temperature radiant systems; the inlet water temperature is fixed to 55[degrees]C (131 [degrees]F) for heating and to 18[degrees]C (64.4[degrees]F) for cooling.
The main air stream is similar to the one for the reference case, but there is only outside fresh air, which supplies the zones with constant airflow totaling 0.275 [m.sup.3]/s (582.7 cfm). In cooling mode for the air-conditioning system, the primary air may be supplied at temperatures below the indoor air temperature set-point to avoid a waste of thermal energy for post-heating after the cooling and dehumidification stages.
The reason of the use of a ventilation system coupled to the radiant systems is necessary both to:
* provide the fresh air to the building (windows are sealed, and in any case, natural ventilation will be uncontrolled);
* control the indoor air humidity ratio, which is not only important with reference to the occupants' thermal comfort but extremely important in the cooling mode for the continuous operation of radiant ceilings and floors and avoid water condensation; water vapor condensation is avoided because the air system keeps the air relative humidity below 60%; at 60% RH and 24[degrees]C (75.2[degrees]F) air temperature; the dew point air temperature is approx 15[degrees]C (59[degrees]F), and since the radiant floor/heating is fed by water at 18[degrees]C (64.4[degrees]F), condensation never occurs.
Three different primary systems were considered. The first (Table 2) is designed similarly to the baseline system: a condensing boiler and an air-cooled chiller are used to cover the radiant floor/ceiling hot water and chilled water loops. These two converters are different to the others that serve the air loop, thus exploiting the benefits of a moderate fluid working temperature for both the boiler and chiller. A second primary system (Table 3) has a high efficiency ground-source reversible heat pump that serves for radiant heating and cooling, thus taking advantage not only of moderate working temperatures of the fluids but also of constant temperature heat source and sink. Finally, a third option (Table 4) is to consider the possibility of covering the radiant system cooling energy demand by means of a ground/pond water free cooling.
Thermal comfort analysis
Even though the set-point temperature schedules and control types of the air-conditioning systems are always the same, the behavior of each system (due to the thermal inertia of the heat transfer components, flow rate control, etc.) is different and may lead to slightly different air temperatures and heating/cooling energy demand, especially in the case of a dual set-point with a dead band.
Since the final scope is contrasting the energy consumed by different systems over a long period of time, a long-term thermal comfort analysis for the Rome location based on the simulated data was developed in order to demonstrate that the level of thermal comfort guaranteed by the two systems is the same. This does not mean that the thermal comfort conditions, nor the air temperature, are always strictly equal between the three simulations (all air, radiant floor, radiant ceiling), but it allows a comparison to be made. The long-term thermal comfort analysis is based on the operative temperature of the core zone of the building and was evaluated as reported in the standard EN 15251 (CEN, 2008). Among the three categories of thermal comfort quality, category II was selected because it coincides with the satisfactory level of previous international standards (e.g., EN ISO 7730) and is adopted worldwide when thermal comfort is concerned (e.g., -0.5 < PMV < +0.5). A performance index (PI) associated with the category represents the percentage of values of operative temperatures during occupied hours that fall within the acceptability range of the category. An indoor environment is supposed to belong to a certain category when the PI is at least 90%.
The frequency of the occurrence of the operative temperatures are reported in Figures 2 to 5. In each figure, the PI for category II (20.0[degrees]C-24.0[degrees]C [68[degrees]F-75.2[degrees]F] for winter thermal comfort, 23.0[degrees]C-26.0[degrees]C [73.4[degrees]F-78.8[degrees]F] for summer thermal comfort) is indicated (winter season lasts from November 15 to March 15, summer season lasts from May 15 to October 15). From the thermal comfort analysis, it can be deduced that the thermal comfort level is equal among the various systems because the category II PI is always greater than 90%.
Provided that both radiant heating and cooling systems guarantee the same quality level of thermal comfort with respect to the all-air systems, the comparison between the reference systems and the radiant system was made in terms of:
* energy delivered to the zone, that is, the energy supplied by primary systems to the secondary system equipment; and
* energy sources fed to the primary system.
Comparison in terms of delivered energy
The heating energy, cooling energy (at various thermal levels), and electricity that are required by the demand-side components (coils, pumps, fans, etc.) of the air and water loops are reported in detail, as an example, for the Rome, London, Frankfurt, and Madrid locations in Tables 5 to 8, and are reported in a graphical comparison by means of a bar graph, where the delivered energy is divided by the conditioned floor area, in Tables 9 and 10. Compared to the all-air system, the radiant systems for Rome show lower energy requirements for air and water circulation in the loops (fans and pumps). While the energy for heating has increased, compared to the all-air system, a reduction of 11% and 15%, respectively, for radiant floor and radiant ceiling can be obtained in the delivered cooling energy for space cooling, fresh air cooling, and humidity control of the indoor environment.
[FIGURE 2 OMITTED]
[FIGURE 3 OMITTED]
[FIGURE 4 OMITTED]
[FIGURE 5 OMITTED]
The fan coil system has generally an energy requirement equal to (e.g., in the London and Frankfurt locations) or lower than the all-air system (e.g., the Rome and Madrid locations). This latter case occurs for mild climates when the cooling energy is greater than the heating energy, because the fan coils provide only a partial dehumidification of the outdoor air, contrary to the all-air system. This fact has to be taken into consideration also when comparing the delivered energy requested by the fan coil system to the delivered energy requested by the radiant systems. Again, this is particularly evident in mild climates (see for example the case of Rome), where the energy needs for space cooling and air dehumidification are greater than in cold climates. In general, for the fan coil systems, the absence of the air humidification in the winter season is one of the reasons for a generalized lower total delivered energy requirement.
The comparison between radiant floor and radiant ceiling also points out that the radiant ceiling energy requirement is smaller due to the lower thermal inertia of the terminal equipment that allows the finest thermostatic control, thus minimizing the overheating and the overcooling and the related energy requirements. This is particularly true for climates in which the cooling requirements are relevant (e.g., Madrid). For radiant floor and ceiling systems, there is also a general reduction in the electricity for air and water movement in the loops.
Comparison in terms of site energy, source energy, operation cost, and C[O.sub.2] emissions
In Table 11, the site energy consumed (natural gas and electricity), source energy, cost for energy sources purchased, and carbon dioxide emissions are reported for the Rome location for the reference system and for the radiant floor. For each indicator, the percentage reduction with respect to the all-air system is evaluated; each of the three different primary systems (#1, #2, and #3) is considered.
Primary system #2 has a high-efficiency ground-source reversible heat pump, thus taking advantage not only of the moderate working temperatures of the fluids but also of the constant-temperature heat source and sink. Primary system #3 is similar to system #2 but covers the radiant system cooling energy by means of a ground/pond water free cooling. The reductions in all three indicators are quite important (up to 50% and still greater for the radiant ceilings) not solely as an effect of the greater efficiency of the primary system installed. Looking through the results of all the locations, for primary system #2 (reported in Table 12), it should be noted that the reductions vary as a function of the weighting factors for electricity assumed from Hastings and Wall (2007). In the case of radiant ceilings, the reductions are still greater (up to 60%), because there is also a reduction in the energy requirements of the secondary system.
From this study, it can be deduced that the adoption of a radiant system coupled with a suitable primary energy system always results in a reduction of exploited energy sources of purchased energy ware costs and of carbon dioxide emissions in comparison with an all-air system. The greatest reductions in carbon dioxide emissions (up to 60%) can be achieved in climates where the energy demand for cooling is higher than the energy demand for heating. In cold climates, the reductions are smaller but are always up to 20%-30%. Similar considerations can be made by comparing radiant system performance and fan coils. Since the latter has an electricity requirement that is always lower than an all-air system (the highest reduction can be appreciated for Madrid location), the reductions in energy sources consumptions and C[O.sub.2] emissions achievable by a radiant system are slightly lower than in the previous cases but is still at least 30%.
[TABLE 9 OMITTED]
[TABLE 10 OMITTED]
This work was developed as a part of a wider research contract conducted by TEBE Research Group of Politecnico di Torino about radiant system energy performances and funded by the UPONOR Italy.
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Enrico Fabrizio, (1), * Stefano P. Corgnati, (2) Francesco Causone, (2) and Marco Filippi (2)
(1) DEIAFA--Universita degli Studi di Torino, Via Leonardo da Vinci, 44, Grugliasco, (TO) 10095, Italy
(2) TEBE Research Group, DENER--Politecnico di Torino, Corso duca Degli Abruzzi, 24, Torino, (TO) 10127, Italy
* Corresponding author e-mail: email@example.com
Received September 22, 2010; accepted March 4, 2011
Enrico Fabrizio, Phi), is Assistant Professor. Stefano P. Corgnati, PhD, Associate Member ASHRAE, is Associate Professor. Francesco Causone, PhD, was Postdoctoral Fellow. Marco Filippi, Member ASHRAE, is Full Professor.
Table 1. Reference primary system. Thermal level Rated efficiency Condensing boiler 80[degrees]C (176[degrees]F) 0.95 Air-cooled chiller 7[degrees]C (44.6[degrees]F) 3.1 Table 2. Primary system # 1. Thermal level Rated efficiency Condensing boiler 80[degrees]C (176[degrees]F) 0.95 Condensing boiler 55[degrees]C (131[degrees]F) 1 Air-cooled chiller 7[degrees]C (44.6[degrees]F) 3.1 Air-cooled chiller 18[degrees]C (64.4[degrees]F) 3.5 Table 3. Primary system #2. Thermal level Rated efficiency Condensing boiler 80[degrees]C (176[degrees]F) 0.95 Reversible heat 55[degrees]C (131[degrees]F) 4.05 pump Air-cooled chiller 7[degrees]C (44.6[degrees]F) 3.1 Reversible heat 18[degrees]C (64.4[degrees]F) 6.50 pump Table 4. Primary system #3. Thermal level Rated efficiency Condensing boiler 80[degrees]C (176[degrees]F) 0.95 Heat pump 55[degrees]C (131[degrees]F) 4.05 Air-cooled chiller 7[degrees]C (44.6[degrees]F) 3.1 Groundwater free 18[degrees]C (64.4[degrees]F) -- cooling Table 5. Rome location. All-air Fan coil Delivered energy, kWh system system Air heating energy, 80[degrees]C (176[degrees]F) 7759 7560 Radiant heating energy, 55[degrees]C 0 0 (131[degrees]F) Air cooling and dehumidification energy, 25,561 19,963 7[degrees]C (44.6[degrees]F) Radiant cooling energy, 18[degrees]C 0 0 (64.4[degrees]F) Electricity for air (fans) and water (pumps) 3748 3192 loops Electricity for humidification 1847 0 Radiant Radiant Delivered energy, kWh floor ceiling Air heating energy, 80[degrees]C (176[degrees]F) 5979 5630 Radiant heating energy, 55[degrees]C 5142 3948 (131[degrees]F) Air cooling and dehumidification energy, 8733 8345 7[degrees]C (44.6[degrees]F) Radiant cooling energy, 18[degrees]C 14,003 13,395 (64.4[degrees]F) Electricity for air (fans) and water (pumps) 2348 2206 loops Electricity for humidification 1833 2058 Table 6. London location. All-air Fan coil Delivered energy, kWh system system Air heating energy, 80[degrees]C (176[degrees]F) 23,671 24,386 Radiant heating energy, 55[degrees]C 0 0 (131[degrees]F) Air cooling and dehumidification energy, 4163 3201 7[degrees]C (44.6[degrees]F) Radiant cooling energy, 18[degrees]C 0 0 (64.4[degrees]F) Electricity for air (fans) and water (pumps) 1890 2923 loops Electricity for humidification 4109 0 Radiant Radiant Delivered energy, kWh floor ceiling Air heating energy, 80[degrees]C (176[degrees]F) 9457 8552 Radiant heating energy, 55[degrees]C 12,808 10,428 (131[degrees]F) Air cooling and dehumidification energy, 1541 1147 7[degrees]C (44.6[degrees]F) Radiant cooling energy, 18[degrees]C 3176 3780 (64.4[degrees]F) Electricity for air (fans) and water (pumps) 2059 2011 loops Electricity for humidification 2481 2901 Table 7. Frankfurt location. All-air Fan coil Delivered energy, kWh system system Air heating energy, 80[degrees]C (176[degrees]F) 28,394 30,567 Radiant heating energy, 55[degrees]C 0 0 (131[degrees]F) Air cooling and dehumidification energy, 7479 6018 7[degrees]C (44.6[degrees]F) Radiant cooling energy, 18[degrees]C 0 0 (64.4[degrees]F) Electricity for air (fans) and water (pumps) 2431 3187 loops Electricity for humidification 5757 0 Radiant Radiant Delivered energy, kWh floor ceiling Air heating energy, 80[degrees]C (176[degrees]F) 12,199 10,453 Radiant heating energy, 55[degrees]C 17,435 14,242 (131[degrees]F) Air cooling and dehumidification energy, 2599 2146 7[degrees]C (44.6[degrees]F) Radiant cooling energy, 18[degrees]C 5109 5552 (64.4[degrees]F) Electricity for air (fans) and water (pumps) 2137 2061 loops Electricity for humidification 3170 3541 Table 8. Madrid location. All-air Fan coil Delivered energy, kWh system system Air heating energy, 80[degrees]C (176[degrees]F) 12,049 11,592 Radiant heating energy, 55[degrees]C 0 0 (131[degrees]F) Air cooling and dehumidification energy, 21,728 18,344 7[degrees]C (44.6[degrees]F) Radiant cooling energy, 18[degrees]C 0 0 (64.4[degrees]F) Electricity for air (fans) and water (pumps) 5104 4555 loops Electricity for humidification 6124 0 Radiant Radiant Delivered energy, kWh floor ceiling Air heating energy, 80[degrees]C (176[degrees]F) 7599 6932 Radiant heating energy, 55[degrees]C 7700 6055 (131[degrees]F) Air cooling and dehumidification energy, 6252 5244 7[degrees]C (44.6[degrees]F) Radiant cooling energy, 18[degrees]C 13,099 12,946 (64.4[degrees]F) Electricity for air (fans) and water (pumps) 2402 2224 loops Electricity for humidification 5421 5398 Table 11. Rome location results (baseline system and radiant floor). Primary energy, System description Energy sources, kWh kWh All-air system (baseline) Natural gas 8094 8094 Electricity 24,103 52,304 Total 60,398 Radiant floor Primary system #1 Natural gas 10,816 Electricity 12,122 Total 37120 Compared to -39% all-air system Primary system #2 Natural gas 5920 Electricity 11,169 Total 30157 Compared to -50% Primary system #3 Natural gas 5920 Electricity 8709 Total 24,818 Compared to -59% all-air system Energy cost, System description Energy sources, [euro] ($) All-air system (baseline) Natural gas 566.58 (707.43) Electricity 3856.48 (4826.42) Total 4989.64 (6230.04) Radiant floor Primary system #1 Natural gas Electricity Total 2696.61 (3366.97) Compared to -46% all-air system Primary system #2 Natural gas Electricity Total 2201.50 (2749.47) Compared to -56% Primary system #3 Natural gas Electricity Total 1807.79 (2257.76) Compared to -64% all-air system C[O.sub.2] emissions, System description Energy sources, [kg.sub.CO2] All-air system (baseline) Natural gas 1627 Electricity 11,087 Total 12,714 Radiant floor Primary system #1 Natural gas Electricity Total 7750 Compared to -39% all-air system Primary system #2 Natural gas Electricity Total 6328 Compared to -50% Primary system #3 Natural gas Electricity Total 5196 Compared to -59% all-air system Table 12. Results for all the locations (with primary system #2). Rome Milan London All-air Natural gas (a) 8094 22,514 24,769 Electricity (b) 24,103 21,226 13,446 C[O.sub.2] emission (c) 12,714 14,289 12,643 Fan coil Natural gas (a) 7872 24,886 25,461 Electricity (b) 22,092 17,128 10,040 C[O.sub.2] emission (c) 11,744 12,881 10,841 Rad floor Natural gas (a) 5920 10,851 10,392 Electricity (b) 11,169 13,078 9104 C[O.sub.2] emission (c) 6328 8197 7278 Rad Natural gas (a) 5574 9683 9002 ceiling Electricity (b) 10,912 10,984 8818 C[O.sub.2] emission (c) 6140 7621 6836 Frankfurt Madrid Athens All-air Natural gas (a) 29,717 12,572 5369 Electricity (b) 18,791 33,939 34,402 C[O.sub.2] emission (c) 18,187 19,497 32,041 Fan coil Natural gas (a) 31,975 12,076 4963 Electricity (b) 13,357 27,467 30,624 C[O.sub.2] emission (c) 15,109 16,161 28,559 Rad floor Natural gas (a) 12,237 7999 5527 Electricity (b) 11,729 15,194 14,409 C[O.sub.2] emission (c) 10,083 9205 14,079 Rad Natural gas (a) 10,486 7296 5199 ceiling Electricity (b) 11,071 14,000 13,440 C[O.sub.2] emission (c) 9304 8466 13,141 Helsinki Moscow All-air Natural gas (a) 43,728 44,147 Electricity (b) 15,199 17,678 C[O.sub.2] emission (c) 16,389 22,663 Fan coil Natural gas (a) 45,153 46,535 Electricity (b) 9731 12,300 C[O.sub.2] emission (c) 13,941 18,948 Rad floor Natural gas (a) 22,539 25,511 Electricity (b) 11,161 12,600 C[O.sub.2] emission (c) 10,111 14,955 Rad Natural gas (a) 17,466 18,440 ceiling Electricity (b) 10,275 11,286 C[O.sub.2] emission (c) 8648 12,510 (a) Measured in kWh. (b) Measured in kW[.sub.e]h. (c) Measured in kg.
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|Author:||Fabrizio, Enrico; Corgnati, Stefano P.; Causone, Francesco; Filippi, Marco|
|Publication:||HVAC & R Research|
|Date:||Aug 1, 2012|
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