ASHRAE 4770-2005 Energy Performances of Prototype VAV and CAV Systems under Simulated Humid Tropical Climates《在模拟潮湿热带气候下的原型变风量和鸡贫血病病毒系统能源表现》.pdf

《ASHRAE 4770-2005 Energy Performances of Prototype VAV and CAV Systems under Simulated Humid Tropical Climates《在模拟潮湿热带气候下的原型变风量和鸡贫血病病毒系统能源表现》.pdf》由会员分享，可在线阅读，更多相关《ASHRAE 4770-2005 Energy Performances of Prototype VAV and CAV Systems under Simulated Humid Tropical Climates《在模拟潮湿热带气候下的原型变风量和鸡贫血病病毒系统能源表现》.pdf（8页珍藏版）》请在麦多课文档分享上搜索。

1、4770 Energy Performances of Prototype VAV and CAV Systems under Simulated Humid Tropical Climates Ahmadul Ameen Member ASHRAE ABSTRACT This paper reports the relative energy consumption of prototype VAVand CAVair-conditioning systems under various load conditions simulating those of a typical ofice

2、building in humid tropical climates such as Singapore s. Tests were conducted inside two adjoining environmental chambers where space loading was variedfrom 0.21 7 to 0.295 kW/m2 (68.81 to 93.54 Btuh fi) with corresponding sensible heat ratio (SHR) ranging from O. 65 to 0.9. The VAVsystem consistent

3、ly consumed less energy compared to the CAVsystem under constant as well as variable load conditions. The percentage energy saving rangedj?om 21.97%at0.21 7kW/m2 (68.81 Btu/h f?) to 12.42% at 0.295 kW/m2 (93.54 Btuh f?) under constant loading, and it decreased with increased space loading. The exper

4、iments measured moderate eneqy savings under the experimental conditions that are assumed to be the typical load condition in Singapore. Because the implementation costs are sign$cant, it appears that a VAV system has marginal economic benejt in Singapore. INTRODUCTION The basic principle of the var

5、iable-air-volume (VAV) air-conditioning system is to control the volume of supply air into the space to be air-conditioned in response to the space load. In constant-air-volume (CAV) systems, however, a constant volume of air is supplied to the conditioned space and the supply air temperature is var

6、ied with variation of space load. The advantages generally attributed to VAV systems are their versatility in individual zone controlled space and their ability to dispense with reheating, and, generally, the systems are considered energy-efficient by design analysis. Further- more, it is claimed th

7、at the first cost of a VAV system is one of Khizir Mahmud Associate Member ASHRAE the lowest of any system of comparable quality in the USA (Gupta et al. 1987). It has also been claimed that qualitatively VAV systems perform better with respect to temperature fluc- tuations (Sekhar and Chung 1998).

8、There are a number of disadvantages of such systems though, e.g., uneven tempera- ture and air distribution, causing lack of air motion. Among other disadvantages is the lack of fresh air at part-load opera- tion, complexity of design, poor air balance, and energy savings not meeting expectations. H

9、owever, there has been significant improvement of VAV systems over the years, resulting in marked improvements with regard to noise, pres- sure imbalance, and control (Gupta et al. 1987; Gardner 1988; Geake 1980). Prior to the large-scale use of VAV systems in the 1970s, most of the air-conditioning

10、 systems installed in commercial premises, particularly in office buildings, were CAV systems using various types of reheat: zone, multizone, and dual duct reheat. In the aftermath of the oil price boom in the early 1970s, VAV systems gained popularity as the potential waste- fulness of reheat syste

11、ms was realized. Another factor for the popularity of VAV systems in air-conditioning applications in the USA is the ease with which conventional systems can be retrofitted with VAV appa- ratus. Heating, ventilating, and air-conditioning (HVAC) systems in many US buildings have been converted to VAV

12、 systems and the performances of many such systems have been reported over the years. Johnson (1984) reported one system where the retrofit yielded an annual energy savings of 46.5% relative to constant volume operation. Kloostra (1979) reported that after the retrofitting of an existing CAV system

13、with VAV air distribution assemblies at a Houston building, energy usage was reduced by 38%. Based on their simulation Ahmadul Ameen is an associate professor and Khizir Mahmud is a graduate student in the School of Mechanical Engineering, Universiti Sains Malaysia, Pulau Pinang, Malaysia. 320 02005

14、 ASHRAE. study, Ardehali and Smith (1996) reported a reduction in energy consumption by VAV systems in comparison with that of CAV reheat systems in excess of 50%. Similar trends of using VAV systems in preference to CAV systems can be found in Singapore over the last several years. However, the rel

15、ative performance of VAV systems over CAV systems in hot and humid tropical climates is not known. Unlike the USA, climatic conditions in humid tropical countries such as Singapore are fairly uniform throughout the year and the diur- nal temperature variation is also very moderate. Furthermore, rehe

16、ating is hardly practiced in commercial premises where load variation is relatively moderate during office hours. Rela- tive humidity is hardly controlled in such premises and is allowed to fluctuate. The load diversity can also be considered to be negligible for such premises. Sekhar (1 997) carrie

17、d out energy simulation studies for a number of types of buildings in Singapore using DOE-2 with a view to comparing the energy performances of VAV and CAV systems. He reported that the savings in total energy consumption of VAV systems ranged between 11.5% and 25.7%. Another important factor to be

18、kept in view is that, unlike in the USA, VAV systems are generally 20%-30% more expensive in terms of first cost in Singapore. The feedback from industry indicates scepticism with respect to the advantages of VAV systems. In view of the above factors, it was planned to undertake experimental investi

19、gations to establish the relative energy consumption of the two systems under simulated climatic conditions. With the availability of such data, life-cycle cost analysis can be carried out for the two systems for any specific project to determine their relative superiority. Research data will also b

20、e useful to determine the feasibility of conversion of CAV systems into VAV systems through retrofitting. Since the comparative tests are not practical in actual building installa- tions, it was considered necessary to construct twin environ- mental chambers, where prototype air-conditioning systems

21、 would be installed and operated under identical simulated load conditions. ENVIRONMENTAL TEST FACILITY Although the immediate objective of developing the envi- ronmental chambers was to carry out comparative perform- ance tests on VAV and CAV systems, additional instruments were incorporated to acc

22、ommodate diverse types of future research activities. Hence, standards and features of similar facilities built elsewhere were reviewed prior to the develop- ment of the facility (ASHRAE 1999; Fahrni 1986; Mueller et al. 1981; Standard Association of Australia 1976). The 8 m x 5 m x 3 m (26.24 ft x

23、16.4 ft x 9.84 fi) envi- ronmental chamber has been constructed with demountable clip-lock type insulated panels in the Thermodynamics Labo- ratory at the Nanyang Technological University, Singapore. The i O0 mm (3.93 in.) insulated panels are of zinc-aluminium alloy coated steel sheets, laminated t

24、o an insulation core of polyurethane. Figure 1 shows an external view of the environ- mental chamber. The chamber was partitioned in the center to Figure 1 An external view of the environmental chamber: create two chambers of equal size for comparative tests (Ameen and Soon 1988). Two air distributi

25、on systems of the same capacity, one of the CAV type and the other of the VAV type, were installed in the two adjoining chambers, as shown in Figure 2. There are two identical air-handling units (AHU) supplying chilled air to the two duct systems in the chambers. Two air-cooled chillers were located

26、 outside, while the AHUs were located inside the laboratory. The nominal cooling capacity of the chillers is 10.5 kW (3TR) (35826 Btuh). The CAV duct system comprises a supply air duct, two diffusers, and a return air duct. The VAV duct system comprises a supply air duct, two VAV boxes, two diffuser

27、s, and a return air duct. The indoor air temperature is maintained by room ther- mostats located in the two chambers. In the CAV system, in which the supply air volume is constant, the variation of load is taken care of by varying the supply air temperature. This temperature is controlled by variati

28、on of chilled water flow through the three-way bypass valve controlled by return air temperature. In the VAV system, however, the room thermo- stat controls the damper inside the VAV box, and the supply air volume varies according to the cooling load inside the condi- tioned space. With reduced load

29、 there is a buildup of air pres- sure inside the supply duct. A static pressure controller, in turn, reduces the fan speed controlled by a static frequency converter. There is a three-way valve in the chilled water return line, which is controlled by return air temperature. Figure 3 shows the contro

30、l schematics of the CAV and VAV air-condi- tioning systems. The above-mentioned instruments and controls are moni- tored by a programmable direct digital controller. The control strategy could be programmed and downloaded into the computer, that is being used to control conditions in both chambers i

31、n accordance with research requirements. There are individual watt-hour meters for recording the energy consumption of the two AHU motors, the two pump motors, and the two chillers. ASHRAE Transactions: Research 321 (a) (b) Figure 2 Terminal units inside the environmental chamber: (a) VAV terminal u

32、nit, (b) CAV terminal unit. 1 _._ _ LEGENDS RS - ROOM TEMPERATURE SENSOR HC - ROOM HUMIDITY SENSOR MV - MODULATING VALVE COMPLETE WITH ACTUATOR 1 TS - DUCTTEMPERATURE SENSOR VSD - VARIABLE SPEED DRIVE SPT- STATIC PRESSURE SENSOR Figure 3 Control schematics of the prototype VAV and CAV air-conditioni

33、ng systems. EXPERIMENTAL INVESTIGATION 1. The control systems were designed and operated in a manner similar to actual installed systems in such premises. The thermostat setpoint temperature was 25C (77F) within the specified range of 23-27C (73.4-80.6“F) recom- mended by local guidelines (Public Wo

34、rks Department 1980, 1982). The upper limit of relative humidity was 75% and indoor air velocity was maintained below 75 mimin (246 fdmin). Experimental Design With a view to ensuring that the performance tests simu- late actual operating conditions in a typical air-conditioned ofice building in a h

35、umid tropical country, i.e., Singapore, located 1.2“ N latitude, where average daytime temperature is around 30C (86F) and relative humidity is over 85%, the following measures were taken: 2. ASHRAE Transactions: Research 322 In order that the simulated load conditions be representative of actual cl

36、imatic and operating conditions, the space-cool- ing load (kW/mz, Btu/h e) was varied in a range compara- ble to similar loading in Singapore office buildings (Loh 1988; Gan and Hui 1987; Wong 1988; Ameen and Soon 1990). Effects of ambient climatic conditions on the space cooling loads were negligib

37、le since the twin environmental cham- bers were made of insulated panels and they were located inside a laboratory. Types of Loading 1 Scope of Experiment Sensible Heat Latent Heat Total Heat Space Load Sensible Heat (kW) (kW) (kW) (kW/m2) Ratio 3.9 2.000 5.900 0.295 0.66 The scope of the experiment

38、 was defined as follows: To carry out the relative energy consumption of the CAV and VAV systems under identical simulated load conditions and at a specified room temperature of 25C (77F). To determine the relative energy consumption of the vari- ous components of the two systems. To record the vari

39、ation of temperature and relative humidity with time in both chambers. 2 3.9 3 3.9 Experimental Procedure The energy consumption tests were performed inside the environmental chambers by using two prototype air-condi- tioning systems. The temperature and relative humidity condi- tions were monitored

40、 throughout each test. The energy consumption of the AHU fan motors, chiller pump motors, and chiller compressors for both the CAV and VAV systems 1.671 5.571 0.278 0.70 i .300 5.200 0.260 0.75 was recorded by their individual watt-hour meters at the end of every test. The sensible heat ratio (SHR)

41、and space cooling load (kW/m2, Btu/h fi2) were varied in a range that is representative of the loads in local office buildings. Sensible heat load was provided by three heaters of 2.4 kW (8189 Btu/h), 1.0 kW (3412 Bhdh), and 0.5 kW (1706 BtUni) each, while the latent heat load was supplied by humidi

42、fiers. The systems were run for three hours at the same time for each test, and the temper- ature and the relative humidity of the systems were recorded every 10 minutes. The space loads for each test are shown in Table 1. Three additional tests were performed with three different intermittent loads

43、 (load number 3 above) to study the relative performance of the two systems under variable load condi- tions. The three tests were designed to run at 50%, 67%, and 75% of full load over a duration of three hours, i.e., the loads were turned on and off at intervals with the use of timers. The loading

44、 cycle schedule is shown in Figure 4. 5 3.9 6 3.9 RESULTS AND DISCUSSION 0.688 4.588 0.229 0.85 0.433 4.333 0.217 0.90 Constant Load The first set of tests was performed with different space cooling loads per m2 (fi2), varying from 0.217 to 0.295 kW/m2 (68.81 to 93.54 Btu/h fi2) with corresponding S

45、HR ranging from 0.66 to 0.90. Figure 5 shows the total energy consump- tion ofthe prototype VAV and CAV systems versus space load- ing while Figure 6 shows the total energy consumption of the prototype VAV and CAV systems versus SHR. Sensible Heat Latent Heat Types of Loading (Btdh) (Btu/h) 1 13307

46、6824 Table la. Space Load Details Total Heat Space Load Sensible Heat (Btulh) (Btu/h. ft2) Ratio 20131 93.54 0.66 3 4 13307 4436 17742 82.44 0.75 13307 3327 16633 77.37 0.80 I 4 I 3.9 I 0.975 I 4.875 I 0.244 1 0.80 I 6 13307 1477 14784 68.81 0.90 I 2 I 13307 I 5701 I 19008 1 88.15 1 0.70 I I 5 I 133

47、07 I 2347 I 15654 1 72.61 I 0.85 I I I I I I I I 50% FullLoad 67 % Full Load 75 Yo Full Load Load(kW)h I oad I , I I I bad5yi I , I I I 52 0 0 30 60 90 120 150 180 0 30 45 75 90 120 130 165 180 45 60 105 120 165 180 Time(inin) Figure 4 Loading cycle schedule. Time(min) 18 , 60000 +VAV -t CAV +%savin

48、gs 0.2 0.22 0.24 0.26 0.28 0.3 .5 1 i io 10000 O n Time(min) 65 70 75 80 85 90 95 100 Space Cooling Load (kWlm) Space Cooling Load (Btuh.ff) Figure 5 Comparative performance of VAV and CAV systems with space cooling load. 18 16 s 14 r, 12 5 10 e 08 g6 w - I- s4 2 O I 25 Ia. 0.6 0.65 0.7 0.75 0.8 0.8

49、5 0.9 0.95 Sensible Heat Ratio 60000 5 50000 ai - 40000 5 o e 30000 P 2. 20000 w s 8 10000 O aVAV - CAV +% savings 0.6 0.65 0.7 0.75 0.8 0.85 0.9 0.95 Sensible Heat Ratio 25 20 %, 15 5 d 10 5 i? a. 5 O 25 20 bl m 15 u) % P 10 2 a. 5 O Figure 6 Comparative performance of VAV and CAV systems with sensible heat ratio. 324 ASHRAE Transactions: Research . 50000 - - - - 45000 * -+-CAV(COMP1 +CAV(AHU) 40000 - I $ 35000 30000 E a 25000 20000 15000 r 10000 5000 - C P w +VAV(COMP) +VAV(AHU) 0.21 0.22 0.23 0.24 0.25 0.26 0.27 0.28 0.29 0.3 65 70 75 80