Indoor Thermal Environment Improvement Based on Switchable Radiation/Convection-Combined Intermittent Heating: Comparison Between Conventional Terminals and an Integrated Novel Terminal

Hongli Sun; Yifan Wu; Borong Lin; Mengfan Duan; Zixu Yang; Hengxin Zhao; Ziliang Wei; Shenfei Yu; Songjun Li; Junkang Song

doi:10.1016/j.eng.2024.08.020

Engineering ›› 2025, Vol. 53 ›› Issue (10) :58 -75. DOI: 10.1016/j.eng.2024.08.020

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Indoor Thermal Environment Improvement Based on Switchable Radiation/Convection-Combined Intermittent Heating: Comparison Between Conventional Terminals and an Integrated Novel Terminal

Hongli Sun ^a^,^e
, Yifan Wu ^b^,^d^,^*
, Borong Lin ^b^,^d^,^*
, Mengfan Duan ^c^,^d
, Zixu Yang ^b^,^d
, Hengxin Zhao ^b^,^d
, Ziliang Wei ^b^,^d
, Shenfei Yu ^a
, Songjun Li ^a
, Junkang Song ^a

Author information +

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Abstract

Intermittent heating is an energy-saving heating mode, which can save energy in terms of time, and thus is worth promoting, particularly in residential heating scenarios. Conventional radiant heating terminals, that is floor heating, and convective heating terminals, that is fan coils, cannot achieve both intermittent and thermal comfort during intermittent heating. Therefore, this study proposes a switchable convective–radiant heating regulation method for floor heating and fan coils to achieve a comfortable indoor environment with high thermal response speed. Furthermore, a novel combined radiant–convective heating terminal was proposed for a reliable and effective solution. Results showed that the proposed switchable method could increase both intermittence and thermal comfort. In addition, the heating terminal showed better heating performance than the combination of two conventional terminals at the key points of heating capacity, flexibility, and thermal response. It could initially heat up a typical residential space within 20–40 min and then stabilize the room temperature in a comfortable range of 18–22 °C, showing great potential for intermittent heating in room-scale heating conditions. This study provides a reference technique for intermittent heating with reduced system complexity and precise environmental control.

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Keywords

Intermittent heating / Radiant–convective heating terminal / Adjustable heating mode / Thermal response speed / Room-temperature distribution

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Hongli Sun, Yifan Wu, Borong Lin, Mengfan Duan, Zixu Yang, Hengxin Zhao, Ziliang Wei, Shenfei Yu, Songjun Li, Junkang Song. Indoor Thermal Environment Improvement Based on Switchable Radiation/Convection-Combined Intermittent Heating: Comparison Between Conventional Terminals and an Integrated Novel Terminal. Engineering, 2025, 53(10): 58-75 DOI:10.1016/j.eng.2024.08.020

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1. Introduction

Climate change has emerged as a global challenge affecting human survival and progress. The building sector, which accounts for 30% of the final energy demand, consumes significant energy for heating and cooling during its operational phase, particularly for heating [1], [2]. Studies on energy-saving methods for space heating applications have been conducted to reduce greenhouse gas emissions.

Factors affecting heating energy consumption include climatic conditions, building envelopes, heating methods, and human activities. Among these factors, climatic conditions have a substantial impact on heating energy consumption. In Europe, the peak heating load of a high-performance passive house in the Nordic region is approximately 50 W·m⁻² [3]. Moreci et al. [4] summarized the empirical heating demands of office buildings. The annual heating demands of Sweden [5], United Kingdom [6], Germany [7], and Belgium were in the range of 58.3–95.0 kW·h·m⁻², whereas it was only 15.0 kW·h·m⁻² in Spain [8]. The different regions of China also experience significant differences in heating load; the heating loads of empirical houses in severely cold regions such as Harbin often being approximately within 60–80 W·m⁻². By contrast, in the Yangtze River region like in Shanghai, the heating loads are significantly lower and are often concentrated at approximately 20–30 W·m⁻² [9]. In regions with cold climates, buildings require a high-performance envelope and airtightness to prevent the cold air from entering in, and centralized heating is typically the primary heating method in these regions. Currently, several studies on centralized heating scenarios in areas with high heating demand have been reported. For instance, Werner [10] shared the results from an inventory of network setups used in various low-temperature district heating projects. Finkenrath et al. [11] presented a holistic modeling and optimization approach for the thermal loads of district heating networks in Germany. However, studies on non-centralized heating areas with low heating loads have also been reported. These low-heating-load regions have large population densities and wide areas [12]. The thermal adaptability [13], living habits [14], and heating behavior [15] of the inhabitants are also extremely different from those in severely cold areas. Although the individual units in these areas have relatively small heating loads, the high population density in large regions leads to a substantially high total heating load. Therefore, adequate attention should be paid to creating a comfortable thermal environment and promoting energy savings in non-centralized heating areas.

Intermittent heating, which exhibits high thermal efficiency and comfort, can fulfill the demand for low heating loads in non-centralized heating areas [16]. However, heating terminals are essential among the various factors in creating indoor thermal environments [17]. Existing heating terminals, including convective and radiant heating, exhibit different heating performances. Recently, the ability of radiant terminals to provide thermal comfort has received increasing attention [18]. However, certain bottlenecks, such as the limited surface temperature and surface area, restrict the heating capacity of radiant heating terminals [19], resulting in low intermittency. Studies have focused on optimization to improve the performance of radiant heating terminals.

Table 1 [20], [21], [22], [23], [24], [25], [26], [27], [28], [29], [30], [31] presents a summary of recent radiant–convective terminals, detailing both radiant and convective heating methods.

As presented in Table 1, studies on the optimization of radiant terminals in terms of surface temperature uniformity and convective terminals combined with novel materials have been conducted. However, the problem of large thermal inertia persists in radiant terminals and draught and discomfort persists in convective terminals. Although the heating capacity in low-heating-load areas is relatively low, intermittent heating cannot easily be achieved using only radiant heating terminals [32]. Compared with radiant terminals, convective terminals typically achieve the desired thermal comfort environment at a faster rate. Wang et al. [33] compared the preheating times of fan coils and radiant floors and found that radiant heating terminals require longer time than convective heating terminals to achieve desired air temperature. Márquez et al. [34] also demonstrated that convective terminals have a faster thermal response during the initial period than radiant terminals. Despite the advantage of convective terminals in terms of response speed, radiant heating systems can create an environment by minimizing airflow and vertical temperature differences [35]. Moreover, during prolonged heating, the thermal environments created by the two terminals show no significant differences [36]. Reconsideration is required to complement the advantages of both types of terminals under intermittent heating conditions.

Therefore, integrating convective and radiant heating terminals to achieve intermittent heating and maintain thermal comfort is a promising strategy. However, methods for achieving the optimal heating mode combining convection and radiation under intermittent mode through a reasonable regulation strategy require further investigation [37]. This strategy involves the following four aspects: applicability of conventional radiant and convective terminals under intermittent heating; integration and refinement of the new radiant–convective terminal for intermittent heating; regulation strategy of radiant and convective heating under intermittent mode; and dynamic performance of indoor thermal environments, such as room-temperature distribution with radiant and convective modes under intermittent heating.

Considering these aspects, the intermittent heating performances of conventional radiant (floor heating) and convective (fan coil) terminals were analyzed and attempts were made to integrate the radiant and convective modes into a novel radiant–convective adjustable heating terminal. In our previous studies, we compared the heating capacity of this new terminal with that of conventional terminals in a single mode [38]. Based on previous research, this study primarily explored the applicability and advantages of this terminal operating in the combined mode. On this basis, a further regulation strategy was proposed under intermittent heating with the radiant–convective adjustable heating terminal, and a comparative experiment between the radiant–convective and conventional terminals was conducted. A specific thermal performance analysis of the radiant–convective adjustable heating terminal was conducted by performing experimental investigations and computational fluid dynamics (CFD) simulations for the indoor environment.

This study followed the structure shown in Fig. 1. Section 2 describes the experimental and simulation methods used in this study. Section 3 analyzes the heating performance and adjustability of the conventional terminal combination and the novel terminal. Section 4 compares the heating performance of the two heating methods mentioned above and provides a further analysis of thermal comfort. The results section focuses on the basic conclusions obtained from the experiment, and the discussion section presents conclusions from the analysis.

2. Methods

2.1. Experimental system

The experiment was conducted in Beijing (China) in November 2023, and two laboratory rooms were selected to represent the residential scene. The thermal environment was manually controlled in both rooms. Both the lab rooms measured 3 m × 3 m × 3 m with a window of 2 m × 2 m on the southern exterior wall, based on the actual size of rooms in small offices and residential houses. Based on the design standards for low-heating-load areas in China (hot-summer and cold-winter areas), the thermal parameters of the building envelope are presented in Table 2. In the radiant–convective terminal system, temperature and pressure sensors were installed at the inlet and outlet of the evaporator and condenser to measure the enthalpy of the refrigerant, with R134 a (one of the most widely used refrigerants) selected as the refrigerant.

The experimental system comprised conventional heating and radiant–convective heating systems. The conventional heating system includes fan coils and floor-heating devices. Their details and heating capacities are listed in Table 3, and the sum of the heating capacities of the conventional terminals is equal to that of the radiant–convective terminal.

The heat transfer of fan coils occurs mainly through forced convection, whereas floor heating occurs mainly through radiation. We did not use a radiator because its heat transfer through natural convection and radiation was inconsistent with experimental objectives.

The radiant–convective heating terminal comprised heat exchangers, serrated fins, cross-flow fans, flat heat pipes (FHPs), and connecting piping. Considering the effect of the thermal buoyancy force of heating in winter, air was expelled from the lower side of the radiant–convective terminal.

As shown in Fig. 2, the heat exchangers are connected to the outdoor unit, serving as condensers of the air-source heat pump system and transferring heat to the radiant–convective terminal.

The heat exchanger contained four branches, which can be turned ON/OFF using valves. The FHPs and serrated fins were located at the front and back of the heat exchangers accessed through Branch 4. The heat exchangers accessed via Branches 1, 2, and 3 had serrated fins on both sides. Additionally, the FHPs and serrated fins were coated with silicone grease.

In the combined mode, the four branches and the fan were opened. After heat was transferred from the heat exchangers to the fins and base of the FHPs, the radiant–convective terminal dispersed heat through FHP radiation and convection aided by cross-flow fans and serrated fins. Similarly, other modes can be realized by adjusting the switch of the branch and air volume of the fan.

The FHPs were evacuated to a 50% vacuum level and then filled with acetone as the working fluid at a filling rate of approximately 20% following engineering practices. The dimension of the radiant–convective terminal was 93 cm × 135 cm × 10 cm (length × height × width).

As shown in Fig. 3, 13 thermocouples were arranged on the front surface of the FHP in both the vertical and horizontal orientations to record the surface temperature. Two thermocouples were placed on the surface of each heat exchanger and on the top of the corresponding fins. Additionally, four thermocouples were installed at the inlet and outlet of the air passage to measure the average air temperature. A detailed description of the components used in the radiant–convective terminal is presented in Table S1 in Appendix A.

In the conventional heating terminal system, a hot water tank is used as the heat source and a distributor is used to control the operation of the floor heating and fan coil terminals in the laboratory room. Temperature sensors were installed at the inlets and outlets of both terminals. Additionally, Coriolis mass flow meters were placed in both the new and conventional systems to measure the flow rates of the refrigerant and hot water.

To evaluate the room-temperature distribution, thermocouples were arranged in both laboratories following the recommended heights (0.1, 0.6, 1.1, and 1.7 m with 2.3 m added) for the average temperatures in ANSI/ASHRAE 55–2023: Thermal Environmental Conditions for Human Occupancy [39]. The room temperature was the mean of all the values from thermocouples spread in the room for the experiment. The radiant–convective terminal was elevated by approximately 0.5 m to obtain a uniform temperature field, as shown in Fig. 4. Ten thermocouples were placed on the interior surface of the building envelope to measure surface temperature. We assumed that the occupants were sitting in the center of the room, and the mean radiant temperature (MRT) was calculated using the angle factor method in ISO 7726: Ergonomics of the Thermal Environment—Instruments for Measuring Physical Quantities [40]. As most building materials have high emissivity, the reflection can be disregarded. The equation used is:

(1)

{\bar{T}}_{r}^{4} = {T_{1}}^{4} F_{p - 1} + {T_{2}}^{4} F_{p - 2} + \dots + {T_{N}}^{4} F_{p - N}

where

{\bar{T}}_{r}

is the MRT, K;

T_{i}

is the temperature of surface i (i = 1, 2, …, N), K; and

F_{p - i}

is the angle factor between a person and surface i, estimated or calculated using a standard.

The thermocouples were connected to a computer using a data collection system to acquire data at a 10 s interval. An anemometer was used to measure the outlet wind speed at the same positions as the temperature measurement points. All thermophysical properties were obtained using REFPROP software (National Institute of Standards and Technology, NIST; USA). Considering intermittent heating, the duration of each case was set to 2 h.

2.2. Experimental conditions

In conventional heating systems, fan coil units use forced convection, whereas floor heating uses natural convection and radiation to heat rooms. Optimal heating was provided by the radiant–convective terminal through the integration of convection and radiation. Simultaneous experiments on radiant–convective and conventional heating terminals were conducted in laboratory rooms to measure their heating capacity and ambient temperature.

Considering the demand, the experimental set-points were within the ANSI/ASHRAE 55–2023 comfort zone. The experiments were conducted under the following conditions:

(1) To ensure the impartiality and accuracy of the experiment, the room temperature was maintained at (20.0 ± 1.0) °C to test the heating capability and room temperature distribution of the terminal.

The following is an evaluation of the heat transfer properties of the radiant–convective terminal using the heating capacity (Q_nt):

(2)

Q_{n t} = \frac{1}{36} G_{r} (h_{r i} - h_{r o})

where G_r is the flow rate of refrigerant, kg·h⁻¹; h_ri and h_ro are the enthalpies of the refrigerant at the inlet and outlet of the radiant–convective terminal, respectively, kJ·kg⁻¹.

The heating capacity is evaluated to assess the heat transfer characteristics of the fan coil and floor heating:

(3)

Q_{t} = \frac{1}{36} c_{w} G_{w} (T_{w i} - T_{w o})

where Q_t is the heating capacity of the conventional heating terminal, W; c_w is the heat capacity of water, kJ·kg⁻¹·K⁻¹; G_w is the hot water flow rate, kg·h⁻¹; and T_wi and T_wo are the inlet and outlet water temperatures of the two conventional terminals, respectively, °C.

(2) To ensure that the working conditions and control variables are comparable, the initial temperature of the room was set at (11 ± 1.0) °C for the experiment that tested the ability of the air conditioning system to heat the environment.

To satisfy the different stages of thermal demand during intermittent heating under low heating loads, the novel terminal alternated between the convection and radiation modes. As shown in Fig. 5, the forced convection airflow of the radiant–convective terminal can be adjusted by a rotary knob to explore three modes of airflow, namely high, medium, and low, corresponding to the outlet wind speeds of 3.8, 2.8, and 1.5 m·s⁻¹ (at corresponding air volumes of 301.5, 222.2, and 119.0 m³·h⁻¹, respectively) from the radiant–convective terminal. Furthermore, the number of openings in the heat exchangers can be adjusted. Two sets of heat exchangers (Branches 1 and 2) and three sets of heat exchangers (Branches 1, 2, and 3) were used in the experiment to provide different amounts of heat energy to satisfy different heating demands.

2.3. Experimental instruments

The experimental parameters are presented in Table 3.

3. Results

3.1. Conventional heating terminals with radiant and convective heating

The thermal performances, with the inlet water temperature of (40 ± 1) °C, of conventional heating terminals in the experimental laboratory rooms were analyzed. As shown in Fig. 6, a small variation is observed between the two at the start of the phase but subsequently becomes almost the same. Considering the small difference between the MRT and air temperature and that this study aims to investigate the characteristics of rising air temperatures under intermittent heating conditions, we focus on air temperature in the following sections. The thermal performance of fan coils with high airflow (340 m³·h⁻¹), which can produce a heating capacity of 1400–1500 W, is shown in Fig. 6(a). For the initial 30 min, the room temperature increased at an average rate of 7.8 °C·h⁻¹, and for the entire 2 h, it increased at a mean rate of 3.95 °C·h⁻¹. Fig. 6(b) shows that the floor heating supplies the heat input within the range of 600–700 W. In the first 30 min and the full 2 h, the room temperature increased at an average rate of 6.2 and 2.15 °C·h⁻¹, respectively.

3.2. radiant–convective adjustable heating terminal with both radiant and convective mode

Convective heating exhibits rapid room heating, whereas the radiation mode creates a uniform surface temperature without causing a draught during the stable period. Therefore, the radiant–convective terminal switches mode to effectively match the demand for intermittent heating.

Fig. 7 shows the switchable mode of the novel terminal. This mode initially uses the convection mode and then the radiation mode. As shown in Fig. 7(a), the novel terminal initially provided convective heating for 40–50 min until the air in the room reached the desired temperature level. Then, it switched to the radiation mode, which provided good thermal comfort. Fig. 7(b) shows the variation in room temperature during the switchable mode. Initially, forced convection was preferred to rapidly raise the room temperature to 18–20 °C, and subsequently, the radiant–convective terminal supplied approximately 700 W of radiant heating to ensure constant indoor temperature.

Compared with the radiation mode, the convection mode exhibited a higher heating capacity for meeting the room heating load. The switchable mode of the novel terminal enables the use of the radiation mode to meet the demand for high thermal comfort, thereby achieving a balance between energy savings and comfort. Notably, the heating capacity of the terminal during the starting period determines the rate at which the room warms. Therefore, in addition to actuating the switchable mode, the novel terminal combines the two modes of heating to enhance its heating capacity. Fig. 8 shows that the heating capacity in the combined mode can be further increased to 1543.2 W. Under the given situation, the room heating sped up, and the room warmed to 18 °C within 20 min under the same indoor and environmental conditions.

3.3. Surface temperature distribution of the radiant–convective terminal

Fig. 9 shows the surface temperature distribution of the radiant–convective terminal obtained by an infrared imager.

Fig. 9(a) shows the surface temperature distribution of the FHP. The temperatures at the first and second horizontally located FHPs were in the superheated state because the lower side of the FHP received heat from the refrigerant flowing in the heat exchanger. After entering the third FHP, the temperature became uniform. The temperature increase in the eighth FHP was caused by the addition of an external extrusion on top of this heat pipe to lower the thermal contact resistance between the FHP and heat exchanger section. This phenomenon is also observed in the infrared image and pictorial view shown in Fig. 9(b). The horizontal temperature distribution of the FHP was relatively uniform, and the surface temperature of the third to ninth heat pipe was in the range of 35–40 °C.

For the vertical temperature distribution in Fig. 9(a), the bottom temperature was as high as 52.0 °C, which was due to direct contact with the heat exchange, and achieved the internal evaporation in the FHP. Except for the lowest evaporator section, the vertical temperature difference of the condensing section of the FHP was extremely small, and the maximum temperature difference (MTD) of the 1 m high heat pipe was only 0.4 °C. This difference showed that the radiation heating obtained from the FHP is uniformly distributed and has good potential for thermal comfort.

3.4. Adjustability of heating capacity with radiant–convective terminal

The novel terminal also provided an adjustable heating capacity for areas with different heating loads. Two types of arrangements were tested for adjustable heating capacity: adjusting the speed of the cross-flow fans and adjusting the number of heat exchangers, as shown in Fig. 10.

The heating capacities of the novel terminal under forced convection at various air velocities are shown in Fig. 10(a). The results indicated that the novel terminal could provide heating capacities within the range of 688.6–1219.0 W. Similarly, Fig. 10(b) shows the heating capacity for different numbers of heat exchanger openings. By opening an additional heat exchanger branch, the heating capacity of the novel terminal increased from 1219.0 to 1721.3 W, indicating an increase of 41.2%. Both methods can provide an adjustable heating capacity with the novel terminal, which can be selected based on different indoor heating demands.

3.5. Thermal performance of the radiant–convective terminal

Fig. 11 shows the basic thermal performance of the radiant–convective terminal by categorizing the heating capacity into low, medium, and high heating grades. Among these, a low heating grade belongs to the radiation mode, and when only one branch of the heat exchanger was opened, a heating capacity of approximately 600 W and room temperature of approximately 16 °C were obtained. When the convection mode (medium heating grade) was activated, two branches of the heat exchanger were opened to raise the heating capacity and room temperature within the range of 650–1250 W and 17–18 °C, respectively. The high heating grade included the combined operation of both heating modes along with the activation of multi-exchanger branches mode, where the heat exchanger opened three branches. The heating capacity was further increased to 1500–1800 W, and the room was rapidly heated to a temperature of 22 °C.

To guarantee the accuracy of the experimental results, the room temperatures were maintained at (20.0 ±1.0) °C when estimating heating capacity, whereas outdoor temperatures were (10.0 ±1.0) °C. By combining and switching the modes of the conventional terminal, the novel terminal can provide variable heating effects in indoor environments for low-heating-load areas, thus fulfilling the requirements of intermittent heating.

4. Discussion

4.1. Comparison of conventional and radiant–convective heating terminals

Fig. 12 shows a comparison of the experimental data for the radiant–convective and conventional heating terminals. Two identical laboratory rooms were used to perform the experiments simultaneously. In one of the laboratory rooms, a fan coil was used initially, and floor heating was used during the stable period to provide good thermal performance for intermittent heating by the joint operation of the conventional heating terminals. However, the results indicated that using the switchable mode of the radiant–convective terminal in the other laboratory room could produce the combined effect of the conventional heating terminals. Fig. 12(a) shows that, for the most part, the novel terminal has a better heating capacity than the conventional heating terminal. However, a brief aberration occurred at the beginning of the switch mode because the novel terminal, which resembles a radiator, was less uniform than radiant floor heating. This discrepancy in temperature contributes to energy waste, which is particularly evident in the pure radiation mode. The switchable mode of the radiant–convective terminal resembled the combined effect of the conventional terminals in terms of heating capacity and rise in ambient temperature and could provide fast warming and comfortable steady-state heating.

The adjustability of the conventional terminal is limited by the inlet water temperature during use. A slight drop in outside temperature from normal was observed on the day of the experiment; consequently, the room could be stabilized only at 16 °C. However, a certain amount of low wind speed forced convection heat exchange can be flexibly supplemented by the radiant–convective terminal to provide high heating capacity. Therefore, the room temperature remains at 18 °C. In conclusion, a single type of radiant–convective terminal can serve as a good substitute for the combined use of two conventional heating terminals, thereby minimizing both initial investment and system complexity. Such a terminal is an excellent choice for intermittent heating in areas with low heat loads. Another advantage of the novel terminal is the flexible adjustment capability of the radiant–convective terminal, enabling it to address variations in the heating load.

4.2. Applicability of the switchable mode of a radiant–convective terminal for intermittent heating

An efficient method for achieving intermittent heating in low-heating-load areas is to combine radiant heating with switchable convective heating. Therefore, the thermal performance of the switchable mode in the use situation was analyzed to verify its applicability.

Fig. 13 shows the switchable mode for two different heating capacity conditions. The specific scenario for the switchable mode involves turning OFF the heating terminal when people leave the heating area. As occupants gradually return to the workplace, the forced convection mode of the radiant–convective terminal is first turned ON, and the number of heat exchanger branches can be selectively turned ON to reach the standard temperature demand within 40–50 min. Subsequently, the operating mode terminal was switched to the radiation mode to enhance thermal comfort by maintaining the basic heating capacity. The results showed that the switchable mode could first heat up rapidly and then stabilize the room temperature in a reasonable range of 18–22 °C, thus providing ideal intermittent heating under low heating load demand.

As shown in Fig. 13, a certain deviation in the initial temperature of the two groups of experimental rooms is observed because of certain fluctuations in the outdoor temperature. A repeatability experiment was conducted concurrently with the switchable modes, considering their importance for the radiant–convective terminals under intermittent heating. In the repeatability experiment, as shown in Fig. 14, the slightly high initial temperature in the original experiment was corrected. Results showed that the repeatability experiment fitted the original experiment well in both the starting and stable periods, and the difference in the initial room temperature due to the fluctuation of the outdoor temperature had less impact on the overall results of the experiment.

4.3. Indoor temperature distribution

The steady-state and dynamic processes of the indoor temperature distribution at the radiant–convective terminal were further analyzed. Considering that the combined mode operated in a constant state, we conducted a steady-state test. By contrast, the switchable mode involved switching from convection to radiation; therefore, dynamic processes were reflective of its characteristics.

As shown in Fig. 15(a), the convection, radiation, and combined modes have maximum vertical temperature differences (MVTDs) of 1.9, 4.5, and 1.7 °C, respectively. Moreover, if the 0.6 m high measuring point is removed, the MVTD of the convection mode is only 0.3 °C. The convection mode ensured a more uniform vertical temperature distribution in the room than the radiation mode.

The dynamic process of the radiation mode is shown in Fig. 15(b), where representative measuring points of 0.1, 1.1, and 1.7 m height are selected for discussion. Compared with the switchable mode, the radiation mode has a larger vertical temperature difference at the starting period, reflecting the disadvantage of the radiation mode in the initial room heating. Therefore, the switchable mode of the radiant–convective terminal was suitable for intermittent heating, reducing the high vertical temperatures caused by the radiation mode.

The space heat utilization in different modes is further discussed, as shown in Fig. 16. Using vertically located measurement points in the room (heights: 0.1, 0.6, 1.1, 1.7, and 2.3 m), the room was divided into five zones of different heights. The four areas below 2.0 m height were defined as the personnel activity area, and the area above 2.0 m height was the non-personnel activity area. The parameter y denotes the heat loss for a non-personnel activity area expressed as

(4)

y = Q_{5} / \sum_{i = 1}^{5} Q_{i} \times 100 % = c_{5} m_{5} {\bar{T}}_{5} / \sum_{i = 1}^{5} c_{i} m_{i} {\bar{T}}_{i} \times 100 %

where Q_i is the heat supplied to the i-th region, J; c_i is the heat capacity of air in the ith region, J·kg⁻¹·K⁻¹; m_i is the mass of air in the ith region, kg; and

{\bar{T}}_{i}

is the average air temperature of the measuring points in the ith region, °C.

In the different modes, if more heat accumulates in the non-personnel activity area, the less effective the heat supply from the room to the personnel. Notably, this finding is valid assuming room air distribution without considering the direct effect of radiation on people at present. The results showed that the radiation mode also had more heat loss in the non-personnel activity area than the convection and combined modes, confirming the results obtained in Fig. 16.

4.4. Thermal comfort

The aforementioned findings showed that convection and combined modes had advantages over radiation modes in terms of room-temperature distribution and heat utilization. However, the radiation mode offered benefits in terms of thermal comfort during the stable period. Fig. 17(a) shows the operative temperatures of the different modes at similar indoor temperatures. The results showed that the operative temperature in the radiation mode was 0.3 °C higher than the indoor temperature, whereas, for the convection mode, it was 1.5 °C lower, reflecting the advantage of radiation mode to some extent. Figs. 17(b) and (c) show the radiation heat transfer ratio and PMV, respectively. Considering the situation of intermittent heating and ANSI/ASHRAE 55–2023 [39], the following assumptions were made for the calculation of PMV: The metabolic rate is 60 W·m⁻² for an adult male sitting still, the garment thermal resistance is 1.37 clo (1 clo = 0.155 m²·°C·W⁻¹) under the dynamic conditions of intermittent heating, and other variables, such as the clothing surface temperature and convective heat transfer coefficient, can be calculated as per the equation given in the ISO 7730 standard [41]. The ambient temperature was 20 °C, and relative humidity was 50%. Among these, MRT and wind speed are variable parameters in different modes, and their experimental values are used to estimate the PMV as follows:

(5)$\begin{aligned}\operatorname{PMV}= & \left(0.303 \mathrm{e}^{-0.036 M}+0.0275\right) \times\{M-W-3.05[5.733-0.007(M-W) \\& \left.-P_{\mathrm{a}}\right]-0.42(M-W-58.2)-0.0173 M\left(5.867-P_{\mathrm{a}}\right) \\& -0.0014 M\left(34-T_{\mathrm{a}}\right)-3.96 \times 10^{-8} f_{\mathrm{cl}}\left[\left(T_{\mathrm{cl}}+273\right)^{4}\right. \\& \left.\left.-(M R T+273)^{4}\right]-f_{\mathrm{cl}} K_{\mathrm{c}}\left(T_{\mathrm{cl}}-T_{\mathrm{a}}\right)\right\}\end{aligned}$

where M is the metabolic rate, W·m⁻²; W is the mechanical work of the human body, W·m⁻²; P_a is the water vapor pressure, kPa; T_a is the ambient temperature, °C; f_cl is the personnel clothing level; T_cl is the exterior surface temperature of the garments, °C; and K_c is the convection heat transfer coefficient, W·m⁻²·K⁻¹.

The overall distribution of the PMV was in the range of −0.8 and 0.4. The radiation mode can also exhibit certain advantages in the PMV. With a high ratio of radiation heat transfer, the PMV approaches 0.4, and the human body feels warm, thereby enhancing thermal comfort during winter. This further establishes the comprehensive advantages of a radiant–convective terminal with a switchable mode in terms of energy savings and thermal comfort under intermittent heating.

4.5. Experimental and numerical investigation of indoor environmental fields

Fig. 18 shows the indoor environmental fields created by the radiant–convective terminal, and the temperature fields are shown at the center of the room. The temperature fields of an indoor environment under the action of the novel terminal employing convection and radiation modes are presented with different airflow velocities under steady-state conditions. Results showed that the convection mode created a uniform temperature distribution. For different wind speeds, the difference was mainly reflected by the effective distance traveled by air. As the wind speed gradually decreased, the hot air increased, resulting in an effective zone up to which the air shrank. However, for the radiation mode, owing to thermal buoyancy, even if the air outlet retained a weak air supply, a relatively high vertical temperature gradient existed in the room. In this situation, heat was concentrated in the upper-left region of the room, preventing heat flow to the personnel activity area. However, the convection mode generally had high wind speeds at a height of 0.6 m. Long-term heating with convection can also cause discomfort.

To reflect the indoor environmental characteristics under different modes intuitively, numerical simulations were performed.

Numerical simulations of the indoor environment were performed using ANSYS-Fluent. Fig. 19 shows the results of the numerical simulation of the indoor environment. This simulation was conducted for two main purposes. First, simulating an indoor environment is a convenient method for analyzing thermodynamics and flow parameters without setting up numerous measuring points in the room, as in the case of laboratory experiments. Second, the temperature distribution in the indoor environmental field assessed by an interpolation method using Matlab has interpolation errors. By contrast, ANSYS-Fluent provides a physical calculation of the indoor environmental field based on CFD, which imbibes the physical character of the system. At the same time, the simulation results can be examined by comparing them with the measured data, ensuring the accuracy of the simulation results.

Considering the situation, the following settings were used in the simulation: Boussinesq assumption was used for air. The k–ε model is suitable for fully developed turbulent conditions, and the discrete ordinates (DO) model is suitable for radiation. The mesh was generated for the air outlet, air inlet, and radiant surfaces of the novel terminal to improve accuracy. The dimensions of the room, radiant–convective terminal, and location of the radiant–convective terminal were consistent with those of the laboratory room. Among these, the measured parameters, such as the outlet temperature, air velocity, and surface temperature of the radiant–convective terminal, were used as input quantities for different modes, and the room-temperature fields under different modes were considered the output. The temperature and wind speed at each measurement point, temperature field, and error in the simulation results were compared with the measured results, as shown in Fig. 19. The final results demonstrated that the majority of the measuring points were within 10% error and that the maximum error was only 2.0 °C (10.8%), proving the accuracy of the simulation results to a satisfactory level.

The overall trend of the simulation results was consistent with the measured data, with high vertical temperature differential in the radiation mode and large relative wind speed in the convection mode. However, some variabilities were observed between the simulation results and measured data. For instance, at the air outlet of the radiant–convective terminal, considerable heat accumulation appeared at the outlet because of the simple temperature interpolation correction of the measured results. This accumulation was caused by an inaccuracy in the interpolation procedure when combined with the simulated outcomes. In addition, some differences in the increases in the supply airflow were observed. In the measured data, the airflow was primarily in the horizontal direction, whereas in the simulated results, it was simultaneously along the horizontal and inclined upward directions. This is also because the measurement was mainly interpolated between the horizontal and vertical measurement points. Therefore, a certain correction error was observed for the inclined airflow, further reflecting the necessity of the numerical simulation. In general, the results were relatively objective and accurate.

4.6. Uncertainty analysis and thermal verification

The observed parameters and their error analyses served as the basis for the uncertainty analysis of the heating capacity. The total uncertainties in the results are expressed by Eq. (6), and the uncertainties were defined using the Kline and McClintock method [42].

(6)

W_{R} = \sum_{i = 1}^{n} {(\frac{\partial R}{\partial v_{i}} w_{v_{i}})}^{\frac{1}{2}}

where W_R is the uncertainty in the result R, W_{V_i} is the uncertainty in the independent variable v_i.

As shown by Eq. (6), the final result, heating capacity, is a function of several independent variables. In conjunction with Eq. (6), the uncertainty of the heating capacity (

W_{Q}

) can be represented as:

(7)

W_{Q} = {[{(\frac{\partial Q}{\partial G} w_{G})}^{2} + {(\frac{\partial Q}{\partial T_{o}} w_{T_{o}})}^{2} + {(\frac{\partial Q}{\partial T_{i}} w_{T_{i}})}^{2}]}^{\frac{1}{2}}

where

w_{G}

is the uncertainty in flow rate of hot water;

w_{T_{o}}

and

w_{T_{i}}

are the uncertainty in the terminal’s inlet and outlet air temperature, respectively; T_i and T_o are the novel terminal’s inlet and outlet air temperature, respectively; Q is the total heating capacity; and G is the flow rate of hot water.

These findings were consistent with the observed data, indicating that the heating capacity (W_Q/Q) uncertainty ranged from 3.6% to 9.5%.

Moreover, thermal verification was performed to confirm the accuracy of the findings, as shown in Fig. 20, with an example of 15 min in a steady state.

Eqs. (2), (8), (9), (10), (11), (12), (13), (14) express the total heating capacity Q, radiation heating capacity Q_r, and convection heating capacity Q_c of the radiant–convective terminal.

(8)

Q_{r} = ε A σ_{b} (T_{s}^{4} {- M R T}^{4})

where ε represents the surface emissivity; A is the surface area of radiant panel, m²; σ_b is the Stefan–Boltzmann constant, 5.67 × 10⁻⁸ W·m⁻²·K⁻⁴; T_s is surface temperature of terminal, °C.

The sum of the forced convection heating capacity Q_f was used to compute the convection heating capacity Q_c as:

(9)

Q_{f} = c ρ V (T_{o} - T_{i})

and natural convection in an infinite space can be used to calculate the heating capacity of natural convection Q_n. The Nusselt number (Nu) and Grashof number (Gr) are expressed as:

(10)

Nu = C {(G r \times P r)}^{n} = C {Ra}^{n}

(11)

Gr = g α Δ T l^{3} / v^{2} = g α (T_{s} - T_{a}) l^{3} / v^{2}

where c represents the heat capacity of air, 1005 J·kg⁻¹·K⁻¹; ρ represents the density of air, kg·m⁻³; V is the forced convection wind speed, m³·s⁻¹; α represents the volume expansion coefficient, K⁻¹; l is the characteristic length, m;

v

represents the kinematic viscosity, m²·s⁻¹; Pr is the Prandtl number; Ra is the Rayleigh number;

C

and n are constants that depend on the flow conditions; and

g

is the acceleration of gravity, m·s⁻².

Q_n and Q_c can be expressed as:

(12)

Nu = K_{n} l / λ

(13)

Q_{n} = K_{n} A (T_{s} - T_{a})

(14)

Q_{c} = Q_{f} + Q_{n}

where K_n represents the natural convection heat transfer coefficient, W·m⁻²·K⁻¹; and λ represents the thermal conductivity of air, W·m⁻¹·K⁻¹.

As shown in Fig. 20, considering the negligible heat storage at the terminal, Q (heat source to terminal) should ideally be equal to the sum of the radiation and convection capacities (Q_r + Q_c) (terminal to room). Therefore, the accuracy of the experimental results was verified by comparing Q and Q_r + Q_c. We found that the error range was within ±10%, validating the experimental results.

4.7. Limitations and future studies

Based on the switching and combination of the radiation and convection modes, the radiant–convective terminal can replace conventional heating terminals and offer a good solution for intermittent heating in low-heating-load areas. However, this study had certain limitations. First, the contact between the FHP and heat exchanger was insufficient; therefore, a certain contact thermal resistance existed. Second, the surface temperature of the FHP was not entirely uniform because of the contact thermal resistance between the exchanger and FHP, leading to radiation inhomogeneity and thus affecting thermal comfort in the radiation mode. Additionally, although the accuracy of the instruments fulfills ISO 7726 requirements, a discrepancy between the actual and ideal accuracy of some parameters, such as air temperature (required: ±0.5 °C; desirable: ±0.2 °C) was observed. This discrepancy may affect the discussion and calculation of some parameters [43], particularly in the thermal comfort part [44].

To address these issues, future studies will include improving the contact method between the FHP and heat exchanger and further reducing the internal thermal resistance by brazing and other methods. Second, the FHP will be treated as a single piece, thereby ensuring good surface temperature uniformity. Simultaneously, a sensor with high accuracy may be replaced to reduce the experimental error caused by the measurement.

5. Conclusions

This study investigated the thermal performance of conventional heating terminals and proposed a novel radiant–convective terminal for intermittent heating. The findings showed that the intermittent heating demand could not be satisfied by convective or radiant heating. Therefore, the thermal performance of intermittent heating was significantly improved by combining radiant and convective heating. Furthermore, the advantages of each heating mode were maximized by using radiant heating during the steady period and convective heating during the initial period. The redesigned radiant–convective terminal could transmit high heating power during the warm-up stage and exhibit less thermal inertia. Moreover, the terminal, with its switchable and combined convection and radiation modes, offered a potential method for intermittent heating. The major outcomes of this study are as follows.

(1) The combined effect of conventional terminals could be accomplished by utilizing the switchable mode of the novel radiant–convective terminal. The switchable mode of the novel radiant–convective terminal could provide rapid and comfortable steady-state heating.

(2) The switchable mode of the novel radiant–convective terminal could provide ideal intermittent heating by rapidly heating the room 20–40 min and stabilizing the temperature within a suitable range 18–22 °C.

(3) Therefore, an efficient method for achieving intermittent heating in low-heating-load areas was the addition of switchable convective heating to radiant heating.

(4) Through experiment and numerical simulation, the indoor environmental characteristics of the radiant–convective terminal under different modes were analyzed. These characteristics have substantial implications for the optimal design of terminal heat transfer in the future.

Overall, the combined use of two conventional heating terminals can be replaced by a single type of radiant–convective terminal. This technique reduces cost and simplifies the system, rendering it an excellent alternative for intermittent heating in regions with low heating loads. Furthermore, the flexible adjustment capability of the radiant–convective terminal effectively addresses variations in heating loads, rendering it beneficial for intermittent heating. The results of this study can be used as a guide for low-load intermittent heating.

CRediT authorship contribution statement

Hongli Sun: Writing – review & editing, Conceptualization. Yifan Wu: Writing – original draft, Validation, Investigation, Data curation, Conceptualization. Borong Lin: Supervision, Methodology. Mengfan Duan: Writing – review & editing, Validation, Investigation. Zixu Yang: Methodology. Hengxin Zhao: Validation, Data curation. Ziliang Wei: Data curation. Shenfei Yu: Writing – review & editing, Validation, Investigation. Songjun Li: Writing – review & editing, Validation, Investigation. Junkang Song: Writing – review & editing, Validation, Investigation.

Declaration of competing interest

The authors declare that they have no known competing financial interests or personal relationships that could have appeared to influence the work reported in this paper.

Acknowledgments

This study was supported by the China National Key Research and Development Program “Integrated convection/radiation coupling terminals for local environment” (2022YFC3801502), the National Natural Science Foundation of China (52108082, 52130803, and 52394223), the China Postdoctoral Science Foundation (2023M732479), the Natural Science Foundation of Sichuan Province of China (2024NSFSC0916), and the New Cornerstone Science Foundation through the XPLORER PRIZE.

Appendix A. Supplementary data

Supplementary data to this article can be found online at https://doi.org/10.1016/j.eng.2024.08.020.

References

Publishing order | Descend order by publishing year | Descend order by cited within

[1]	United Nations Environment P.2023 global status report for buildings and construction: beyond foundations—mainstreaming sustainable solutions to cut emissions from the buildings sector.Nairobi: United Nations Environment Programme; 2023.

[2]	Building Energy Efficiency Research Center, Tsinghua University.2024 annual report on China building energy efficiency.Beijing: China Architecture and Building Press; 2024. Chinese.

[3]	Kazanci OB, Shukuya M.A theoretical study of the effects of different heating loads on the exergy performance of water-based and air-based space heating systems in buildings.Energy 2022; 238:122009.

[4]	Moreci E, Ciulla G, Lo BV.Annual heating energy requirements of office buildings in a European climate.Sustain Cities Soc 2016; 20:81-95.

[5]	Wallhagen M, Glaumann M, Malmqvist T.Basic building life cycle calculations to decrease contribution to climate change—case study on an office building in Sweden.Build Environ 2011; 46(10):1863-1871.

[6]	Carbon Trust.Energy consumption guide 19: energy use in offices.London: Carbon Trust; 2003.

[7]	Mateus T, Oliveira AC.Energy and economic analysis of an integrated solar absorption cooling and heating system in different building types and climates.Appl Energy 2009; 86(6):949-957.

[8]	Córdoba J, Macías M, Espinosa JM.Study of the potential savings on energy demand and HVAC energy consumption by using coated glazing for office buildings in Madrid.Energy Build 1998; 27(1):13-19.

[9]	Building Energy Efficiency Research Center, Tsinghua University.Annual report on China building energy efficiency.Beijing: China Architecture and Building Press; 2020. Chinese.

[10]	Werner S.Network configurations for implemented low-temperature district heating.Energy 2022; 254:124091.

[11]	Finkenrath M, Faber T, Behrens F, Leiprecht S.Holistic modelling and optimisation of thermal load forecasting, heat generation, and plant dispatch for a district heating network.Energy 2022; 250:123666.

[12]	Hu S, Yan D, Cui Y, Guo S.Urban residential heating in hot summer and cold winter zones of China—status, modeling, and scenarios to 2030.Energy Policy 2016; 92:158-170.

[13]	Chen S, Wang X, Lun I, Chen Y, Wu J, Ge J.Effect of inhabitant behavioral responses on adaptive thermal comfort under hot summer and cold winter climate in China.Build Environ 2020; 168:106492.

[14]	Jiang H, Yao R, Han S, Du C, Yu W, Chen S, et al.How do urban residents use energy for winter heating at home? A large-scale survey in the hot summer and cold winter climate zone in the Yangtze River region.Energy Build 2020; 223:110131.

[15]	Wang Z, Zhao Z, Lin B, Zhu Y, Ouyang Q.Residential heating energy consumption modeling through a bottom–up approach for China’s Hot Summer–Cold Winter climatic region.Energy Build 2015; 109:65-74.

[16]	Wang X, Fang Y, Cai W, Ding C, Xie Y.Heating demand with heterogeneity in residential households in the hot summer and cold winter climate zone in China—a quantile regression approach.Energy 2022; 247:123462.

[17]	Kazanci OB, Olesen BW.IEA EBC Annex 59—possibilities, limitations and capacities of indoor terminal units.Energy Procedia 2015; 78:2427-2432.

[18]	Hooshmand SM, Zhang H, Javidanfar H, Zhai Y, Wagner A.A review of local radiant heating systems and their effects on thermal comfort and sensation.Energy Build 2023; 296:113331.

[19]	Shinoda J, Kazanci OB, Tanabe S, Olesen BW.A review of the surface heat transfer coefficients of radiant heating and cooling systems.Build Environ 2019; 159:106156.

[20]	Chae YT, Strand RK.Thermal performance evaluation of hybrid heat source radiant system using a concentrate tube heat exchanger.Energy Build 2014; 70:246-257.

[21]	Li T, Liu Y, Chen Y, Wang D, Wang Y.Experimental study of the thermal performance of combined floor and Kang heating terminal based on differentiated thermal demands.Energy Build 2018; 171:196-208.

[22]	Sun H, Lin B, Lin Z, Zhu Y.Experimental study on a novel-convective flat-heat-pipe heating system integrated with phase change material and thermoelectric unit.Energy 2019; 189:116181.

[23]	Li Z, Zhang D, Li C.Experimental evaluation of indoor thermal environment with modularity radiant heating in low energy buildings.Int J Refrig 2021; 123:159-168.

[24]	Shao S, Zhang H, You S, Zheng W, Jiang L.Thermal performance analysis of a new refrigerant-heated radiator coupled with air-source heat pump heating system.Appl Energy 2019; 247:78-88.

[25]	Shao S, Zhang H, Fan X, You S, Wang Y, Wei S.Thermodynamic and economic analysis of the air source heat pump system with direct-condensation radiant heating panel.Energy 2021; 225:120195.

[26]	Sun H, Duan M, Wu Y, Lin B, Yang Z, Zhao H.Thermal performance investigation of a novel-convective heating terminal integrated with flat heat pipe and heat transfer enhancement.Energy 2021; 236:121411.

[27]	Dong J, Zhang L, Deng S, Yang B, Huang S.An experimental study on a radiant–convective heating system based on air source heat pump.Energy Build 2018; 158:812-821.

[28]	Dong J, Zheng W, Ran Z, Zhang B.Experimental investigation on heating performance of a novel-convective radiant–convective heating terminal.Renew Energy 2021; 164:804-814.

[29]	Hu B, Wang R, Xiao B, He L, Zhang W, Zhang S.Performance evaluation of different heating terminals used in air source heat pump system.Int J Refrig 2019; 98:274-282.

[30]	Wu Y, Sun H, Duan M, Lin B, Zhao H, Liu C.Radiant–convective radiation–adjustable heating terminal based on flat heat pipe combined with air source heat pump.Engineering 2023; 20:192-207.

[31]	Lu Y, Dong J, Lu H, Niu D, Zhang S, Fang Z, et al.Stratum-air-distributed radiant–convective room air conditioner for heating.Energy Build 2022; 271:112311.

[32]	Duan M, Wu Y, Sun H, Yang Z, Shi W, Lin B.Intermittent heating performance of different terminals in hot summer and cold winter zone in China based on field test.J Build Eng 2021; 43:102546.

[33]	Wang Z, Luo M, Geng Y, Lin B, Zhu Y.A model to compare convective and radiant heating systems for intermittent space heating.Appl Energy 2018; 215:211-226.

[34]	Márquez AA, López JMC, Hernández FF, Muñoz FD, Andr ACés.A comparison of heating terminal units: fan-coil versus radiant floor, and the combination of both.Energy Build 2017; 138:621-629.

[35]	Imanari T, Omori T, Bogaki K.Thermal comfort and energy consumption of the radiant ceiling panel system: comparison with the conventional all-air system.Energy Build 1999; 30(2):167-175.

[36]	Sun H, Yang Z, Lin B, Shi W, Zhu Y, Zhao H.Comparison of thermal comfort between convective heating and radiant heating terminals in a winter thermal environment: a field and experimental study.Energy Build 2020; 224:110239.

[37]	Benakopoulos T, Vergo W, Tunzi M, Salenbien R, Kolarik J, Svendsen S.Energy and cost savings with continuous low temperature heating versus intermittent heating of an office building with district heating.Energy 2022; 252:124071.

[38]	Sun H, Duan M, Yang Z, Ding P, Wu Y, Lin B.Evaluation of the intermittent performance of heating terminals based on exergy analysis: discriminate the impacts of heat and electricity input.Appl Energy 2023; 346:121331.

[39]	ANS I/ASHRA E 55–2023: Thermal environmental conditions for human occupancy.American Standard.New York City: American National Standards Institute; 2023.

[40]	IS O 7726: Ergonomics of the thermal environment—Instruments for measuring physical quantities.ISO standard.Geneva: International Standardization Organization; 2002.

[41]	IS O 7730: Ergonomics of the thermal environment—Analytical determination and interpretation of thermal comfort using calculation of the PMV and PPD indices and local thermal comfort.ISO standard.Geneva: International Standardization Organization; 2005.

[42]	Kline SJ, McKlintock FA.Describing uncertainties in single sample experiments.Mech Eng 1953; 75:3-8.

[43]	D FR’Ambrosio Alfano, Olesen BW, Palella BI, Riccio G.Thermal comfort: design and assessment for energy saving.Energy Build 2014; 81:326-336.

[44]	D FR’Ambrosio Alfano, Palella BI, Riccio G.The role of measurement accuracy on the thermal environment assessment by means of PMV index.Build Environ 2011; 46(7):1361-1369.