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• The two subsystems operate together at hot and humid outdoor climate;
• Only the humidity control subsystem operates at cold but humid ambient condition;
• Outdoor air is directly introduced into occupied spaces after filteringwhen outdoor air is dry enough, such as 11 g/kg.
3. Performance test of the THIC system
3.1. Indoor thermal environment
Fig. 7 shows the tested results of indoor temperatures, humidity ratios and CO2 concentrations with the outdoor temperature and relative humidity of 34.9 ◦C and 61% respectively. As indicated by the figure, the THIC system could provide a comfortable indoor environment with suitable thermal condition and good indoor air quality.
Fig. 8 depicts the tested temperatures and humidity ratios along the vertical direction in the vestibule. In the occupied zone (the height within 2m), the temperature and humidity ratio were about 26 ◦C and 12 g/kg respectively which meet human thermal comfort well. Both the temperature and humidity ratio increased fast along the vertical direction, which reached about 30 ◦C and 20 g/kg respectively at the height of 10 m. And the peak temperature at upper space (over 7m) occurred at noon, due to the strong solar radiation and high ambient temperature. According to the test results, the THIC system applied in large space is effective on energy-saving, which only keeps the occupied zone in comfort condition and forms apparent stratification of indoor temperature and humidity ratio along the vertical direction. Moreover, the natural ventilation from the shutters contributed to remove the absorbed heat of decorations to outdoor environment.
3.2. Energy efficiency of the THIC system
The field test of the energy efficiency of the THIC system was conducted both under the partial load condition and very hot and humid outdoor condition. The former one is on May 27th, 2009, with ambient temperature of 29.3 ◦C and relative humidity of 79% (absolute humidity 20.3 g/kg), and the latter one is on July 16th, 2009, with ambient temperature of 34.9 ◦C and relative humidity of 61% (absolute humidity 21.6 g/kg). The measurement was pided into two parts: humidity control subsystem and temperature control subsystem.
3.2.1. Energy efficiency of humidity control subsystem
The performances of the liquid desiccant fresh air units were tested one by one on May 27th, 2009, according to the measured flow rates, air inlet and outlet parameters through the processor and input power of compressors, solution pumps and fans. The tested results of seven fresh air units are summarized in Table 2; the other two processors are neglected due to the difficult of installing the measuring sensors.The east side processor of the 2nd floor is a typical example for the fresh air unit, and its specific operation information is shown in Table 3. The fresh air flow rate was 5059m3/h, the outdoor air parameter was 29.3 ◦C and 20.3 g/kg, and the supply air parameter was 17.1 ◦C and 6.2 g/kg. So the cooling capacity (Qair), calculated by energy balance equation, was 82.6 kW. The input power of compressors and solution pumps (Pair) inside the processor was 17.8 kW, and the input power of the supply air and exhaust air fans (Pfan) was 2.2 kW. Therefore, the performance of the fresh air unit (COPair), the transport coefficient of fans (TCfan) and the performance of the entire humidity handling process (COPhum), as shown in Eqs. (1)–(3), are 4.7, 37.5 and 4.2 respectively.
As shown in Table 2, the COPair of the tested seven fresh air units are in the range of 4.4–4.9, the TCfan of the fans are 35–40, and the COPhum of the entire humidity handling processes are 4.0–4.4.
According to the test data of May 27th, 2009 and rated parameters of the fresh air units and fans, the calculated cooling capacity of the entire humidity control subsystem is 773.0kW with total inside compressors and solution pumps input power of 166.9kW and total fans input power of 20.0 kW, so the coefficient of performance of the humidity control subsystem (COPHUM), shown in Eq. (4), is 4.1.