ASHRAE 4837-2006 Modeling and Testing of a Utility Peak Reducing Residential Hot Dry Air Conditioner (HDAC) Using Microchannel Heat Exchangers《用微通道换热器的高峰减少住宅热 干空调器(hdac)效用建模和测试》.pdf

《ASHRAE 4837-2006 Modeling and Testing of a Utility Peak Reducing Residential Hot Dry Air Conditioner (HDAC) Using Microchannel Heat Exchangers《用微通道换热器的高峰减少住宅热 干空调器(hdac)效用建模和测试》.pdf》由会员分享，可在线阅读，更多相关《ASHRAE 4837-2006 Modeling and Testing of a Utility Peak Reducing Residential Hot Dry Air Conditioner (HDAC) Using Microchannel Heat Exchangers《用微通道换热器的高峰减少住宅热 干空调器(hdac)效用建模和测试》.pdf（9页珍藏版）》请在麦多课文档分享上搜索。

1、4837 Modeling and Testing of a Utility Peak Reducing Residential Hot/Dry Air Conditioner (HDAC) Using Microchannel Heat Exchangers Clark W. Bullard, PhD John Proctor Joseph Brezner Fellow ASHRAE Member ASHRAE Member ASHRAE Kevin B. Mercer Associate Member ASHRAE ABSTRACT This paper describes the des

2、ign, simulation, building, and testing of a 10.5 kW (%ton) residential air conditioner for hot/ dry climates. The unit was optimized to field-measured indoor and outdoor conditions und duct system losses, subject to the same packaging constraints as conventional systems. A new type of microchannelfl

3、at tube plate fin evaporator and condenser were designed and fabricated for this project. The system was tested in a psychrometric room at extreme hot /dry conditions and met the project goal by reducing peak power demand by more than 25% compared to u typical SEER 12, R-41 OA baseline unit. Compone

4、nt andsystem performance data were compared to the results of simulution modeling conducted during the design process. Results suggest that additional performance improvements might be obtainable with minor changes in system configuration. INTRODUCTION California is a summer peaking utility region w

5、ith air conditioning as the largest electric load at peak demand peri- ods. Californias peak electric demand raises many environ- mental issues related to the need for additional power plants and transmission infrastructure. Even high-performance air- conditioning systems are not optimized to reduce

6、 peak electric demand and energy under hot and dry ambient conditions. This projects primary design criterion is to design an affordable air conditioner that reduces peak electrical demand in hotldry regions. A hot/dry region is defined as a climate in which the outdoor humidity is lower than the in

7、door humidity at summer average and summer design conditions. From field data, it was determined that this unit should be optimized for an outside temperature of 46.1 “C (1 15F) and inside condi- Robert A. Davis Member ASHRAE tions of 26.7”C (80F) dry bulb and 17.2”C (63F) wet bulb (Downey and Proct

8、or 2002). Field data also dictated that the design should be optimized to a duct system K factor of 0. 136 m3/s-Pa” (1442 cfdiwc”), the resistance to airflow that produces 125 Pa (0.50 iwc) static pressure (excluding the evaporator coil) at 0.48 1 m3/s (1020 cfm) (Proctor and Parker In California, t

9、ypical split-system air conditioners consist of a condensing unit and an evaporator in the supply plenum of an attic-mounted gas furnace. The packaging constraints for the hotldry air conditioner (HDAC) indoor unit were kept the same as the baseline system. This system was designed to reduce peak lo

10、ad, improve energy savings, and meet the typi- cal requirements of an air conditioner installed in California. In hot/dry climates such as California, Nevada, and Arizona, dehumidification is rarely required since infiltration of dry air is sufficient to dehumidiSl the space. Therefore, this project

11、 focused on sensible capacity and sensible energy effi- ciency ratio (EER) as the metrics of interest. However, the unit is also controlled to dehumidify when conditions require it. 2000). SYSTEM SI MU LATI ON A SEER 12, R-4 1 OA air conditioner was laboratory tested to provide a baseline for a vari

12、ety of conditions, including the HDAC design conditions. The baseline unit was then modeled using the public domain system simulation developed by Rice (2002). The simulation model was then adjusted to match the laboratory results. Starting from this simulation, heat exchanger parameters, compressor

13、s, fans, motors, and expan- sion devices were evaluated to assess their performance under the design conditions. Clark W. Bullard is professor of mechanical engineering at the University of Illinois at Urbana-Champaign. John Proctor and Joseph Brezner are with Proctor Engineering Group, San Rafael,

14、Calif. Kevin B. Mercer is with Modine Manufacturing Co., Racine, Wis. Robert A. Davis is with Pacific Gas and Electric Co., San Ramon, Calif. 1 62 02006 ASHRAE. Table 1. Specifications for Baseline and HDAC Systems Baseline Round Tube HDAC Flat Tube HDAC Compressor Reciprocating Reciprocating Recipr

15、ocating Type Rated Capacity (Btuh at 45C; 130F) 30,600 27,235 27,235 Evaporator Round CdAI Round CulAl Flat AI/A1 0.41; 4.42 0.39; 4.20 Face Area (m2; ft2) 0.32; 3.46 Air-side Area (m2; fi2) 20.8; 224 27.1; 292 18.4; 198 Fin Type Slit Louvered Louvered Fin Pitch; Density (mm; fpi) 1.8; 14.5 1.8; 14.

16、5 2.1; 12 Rows Deep 3 3 2- Core Depth (mm; in.) 57; 2.25 57; 2.25 44; 1.73 Mass (kg; Ibrn) 6.3; 13.9 8.1; 17.8 8.1; 17.8 Airflow (m3/s; cfm) Configuration Slab coil A-coil A-coil 0.566; 1200 0.519; 1100 0.515; 1092 Condenser Round Cu/AI Round CdA1 Flat Al/AI Face Area (m2; fi2) 1.38; 14.80 1.55; 16.

17、66 1.39; 14.96 Air-side Area (m2; ft2) 49.8; 537 55.6; 598 33.3; 358 Fin Type Wavy Louvered Louvered Fin Pitch; Density (mm; fpi) 1.0; 25 1.3; 20 2.1; 12 Rows Deep 1 1 1 Core Depth (mm; in.) 19; 0.75 19; 0.75 22; 0.87 Mass (kg; Ibrn) 12.4; 27.3 14.0; 30.8 13.1; 28.9 Airflow (m3/s; cfm) 1.31; 2813 1.

18、42; 3000 1.5; 3178 Refrigerant R4 1 OA R4 1 OA R4 1 OA Initially, the potential for achieving the project goals using conventional round tube plate fin heat exchanger tech- nology was evaluated. It was found that improved perfor- mance required more heat transfer surface in the form of additional fa

19、ce area, tube rows, or fin density, so the packaging constraints would have needed to be relaxed, as shown in Table 1, if conventional round tubes were used. The project also explored the potential advantages of microchannel heat exchanger technology, developed for compact applications such as autom

20、otive air-conditioning condensers but not yet optimized for residential applications. The model used for this purpose was a detailed system model composed of extensively validated component models and based on a Newton-Raphson solver (Klein and Alvarado 2004; Jain and Bullard 2004). The model employ

21、ed air-side heat transfer and pressure drop correlations developed for micro- channel heat exchangers by Chang and Wang (1 996, 1997); refrigerant-side areas in the flat tubes are typically -5 x larger than round copper tubes, so results are relatively insensitive to the refrigerant-side correlation

22、s employed. Results indicated that the high air-side performance would allow relatively thin heat exchangers to achieve extremely small approach temper- ature differences and that low air-side pressure drops would make it possible to increase airflow rates to provide a larger heat source and sink. F

23、or a real system to approach the ideal vapor compression cycle efficiency-evaporating and condensing at the indoor and outdoor air temperatures, respectively-infinite heat exchanger UAs are not enough. The heat exchangers must also be provided with access to an infinite heat source and sink-that is,

24、 infinite airflow rates. Finite airflow rates undergo finite temperature changes, thus limiting overall cycle efficiency. Recall that extensive field measurements revealed that duct pressure losses faced by split systems averaged 125 Pa at 0.481 m3/s (0.50 in H20 at 1020 cfm), far greater than ASHRA

25、E Transactions: Research 163 Table 2. Modeled Performance of HDAC Units at 46.1126.7117.2“C (I 15/80/63“F)* Round Tube HDAC Baseline (Measured Data) Flat Tube HDAC Performance EER (BtuWh) Sensible EER (BtuWh) 6.97 6.23 8.63 8.25 8.90 8.90 Sensible Eficiency Increase 30% 48% Dry Climate Peak Reductio

26、n 24% 30% Capacity (kW, kBtuh) 8.0; 27,s 7.9; 27.0 8.1; 27.5 Sensible Capacity (kW; kBtuk) 7.2; 24,7 7.5; 25.6 8.1; 27.5 Power Consumption Compressor (W) Evaporator Blower (W) Condenser Fan (W) 3250 536 150 2572 361 196 2682 281. 134 Fans YO of Total 17% 18% 13% Total 3936 3129 3097 Saturation Tempe

27、ratures Suction (OC; OF) 6.7; 44 8.33; 47 11.4; 52 Discharge. (“C; OF) 57.4; 135 55.6; 132 53.7; 129 * Condenser air entenngreturn dry-bulbheturn wet-hulh temperatures. the 37.5 Pa at 0.566 m3/s (1200 cfm) used in the AR1 and DOE standard tests. Therefore, the indoor air flow rate cannot be increase

28、d in retrofit situations to provide a larger heat source for the air-conditioning cycle without increasing blower power penalties. It was, however, possible to increase the outdoor airflow by taking advantage of the low air-side pres- sure drop afforded by the flat tubes and larger fin pitch. The mo

29、deling results led to the component specifications and performance predictions shown in the third columns of Tables 1 and 2, respectively. System capacities shown are net, after accounting for heat added by the blower. Since cost compar- isons were outside the scope of this project, heat exchanger s

30、urface areas and masses are included in Table 1 to facilitate such analyses. The SEER 12 baseline unit was tested at hoddry condi- tions and the resulting data are shown in Table 2 along with the simulation results for the HDAC prototype systems in columns 2 and 3. The baseline units 9.5 mm (0.375 i

31、n.) round evaporator tubes were rifled and the condenser tubes were smooth. The HDAC prototypes flat tube outer dimensions were 1.9 x 19 mm (minor x major). MICROCHANNEL EVAPORATOR AND CONDENSER The first step in designing both heat exchangers was to sure drops and empirical correlations based on da

32、ta from the specific fin and tube geometries. Figure 1 shows details of the fin and tube assembly selected for these prototypes. The plate fins are fundamentally different from conventional serpentine (folded) fins used in automotive applications. These fins were designed with residential and commer

33、cial HVAC applications in mind to reduce manufacturing costs and address condensate shedding issues that sometimes prove problematic with conventional serpentine fins (Kim et al. 2001,2002). Only one fidlouver design configuration and two tubes were available from the manufacturer at the time of thi

34、s experiment. The degrees of freedom for design were reduced because of these limitations, thus allowing only fin pitch and face area to be changed. After modeling each coil, a triangular port tube, as shown in Figure 1, was chosen for both the evaporator and condenser. Both the evaporator and conde

35、nser are all-alumi- num heat exchangers. A microchannel condenser enabled design targets to be met with less than 1% increase in face area, 16% deeper core, and about half the fin density, as shown in Table 1, when compared to the baseline. A three-pass design was employed to provide connections on

36、opposite ends and ensure nominal pressure drop through the coil. A total of 71 tubes were used in the condenser, with 38 tubes in the first pass, 24 tubes in the examine their contributions to overall system performance using the system simulation model. Detailed design was performed later by the ma

37、nufacturer using proprietary models that contained more detailed representations of header pres- second pass, and 9 tubes in the third pass. The low air pressure drop of the outdoor condenser enabled a larger (heat sink) airflow rate that was crucial to increasing system efficiency. The outdoor coil

38、 is shown in Figure 2. 164 ASHRAE Transactions: Research 4Q Ports Tube Assembly - Front side Figure 1 Tube and)n assembly detail. Since the indoor airflow rate could not be increased due to the high duct pressure losses, the microchannel evaporator did not offer the same advantage as did the condens

39、er. However, its high air-side heat transfer coefficient (despite the lower face velocity) did allow the core depth to be reduced substantially. The microchannel and plate fins were selected in order to test their condensate shedding performance for wet- coil conditions where SEER is determined and

40、to satisfi pack- aging constraints. The face area of the A-coil prototype was approximately 22% larger and its core 23% thinner than that of the baseline slab-coil evaporator, but it was still within the established cabinet dimensions. Note in Figure 1 that the curvature of the plate fins adds about

41、 25% to the heat transfer surface area compared to a conventional microchannel heat exchanger with serpentine fins at the same nominal pitch. The microchannel A-coil was a single-pass design with six baffled inlet sections and one outlet connection centered in the outlet manifold. The coil was shape

42、d by folding it into a cross-counter-flow orientation. Folding resulted in the A-coil being two rows deep with all connections located just outside the apex of the coil for easy attachment to the suction line and expansion device. The prototype evaporator was initially tested by the manufacturer in

43、a transparent wind tunnel to better understand issues related to condensate drainage, refrigerant distribution, and overall coil performance. The evaporator as it was posi- tioned in the wind tunnel is shown in Figure 3. The coil was Plate Fin Figure 2 Prototype microchannel condenser: tested assumi

44、ng an upflow design, where a blower would be located below the coil. The test conditions and data recorded are listed in Table 3. A dry condition at -27C db/l3“C wb (80/56“F) and a wet condition at -27/19“C (80/67“F) were tested; measured capacity and air-side pressure drop from these tests are show

45、n ASHRAE Transactions: Research 165 Table 3. Test Condition and Data for A-coil Evaporator Inlet Air Inlet Air Airflow Wet-Bulb Dry-Bulb Coil Ref. Mass Inlet Pressure Outlet Outlet (lbmlh; g/s) (psia; kPa) (psia; kPa) (OF; OC) (cfm; m3/s) Rate Temperature Temperature State (OF; OC) (OF; OC) Flow at

46、Manifold Pressure Temperature 416; 52 134; 926 133; 918 51.8; 11.0 1200; 0.56 55.5; 13.1 80.0; 26.7 DRY 358; 45 134; 925 133; 919 52.8; 11.6 1000; 0.47 56.3; 13.5 80.0; 26.7 DRY 282; 35 135; 93 1 134; 928 55.3; 13.0 800; 0.38 57.4; 14.1 80.0; 26.7 DRY 568; 71 132; 914 130; 899 53.0; 11.7 1200; 0.56

47、67.1; 19.5 80.0; 26.6 WET 512; 64 133; 917 131; 905 51.5; 10.8 1000; 0.47 66.9; 19.4 80.0; 26.6 WET 397; 50 133; 919 132; 912 52.7; 11.5 800; 0.38 66.4; 19.1 80.0; 26.7 WET Figure 3 Prototype microchannel evaporator: in Figure 4. Air-side pressure drop was very low for the A-coil at all airflow rate

48、s of 0.378,0.472, and 0.566 m3/s (800, 1000, and 1200 cfm). In addition, for the wet case, condensate drain- age at all three airflows proved to be very good. Figure 4 also shows a comparison of the air-side pressure drop for a dry and wet coil at each airflow rate. After the 0.566 m3/s (1200 cfm) a

49、irflow rate test point was recorded, the airflow was increased to a maximum value exceeding 1.42 m3/s (3000 cfm). No condensate blow-off was observed even at this high airflow rate. Condensate was observed draining rapidly down the inside of the heat exchanger along the exposed edges of the tubes. Hantz et al. (2005) were able to demonstrate satisfac- tory water drainage for this same fin design when used in a heat pump application. The predicted simulation results are taken from the manu- facturer?s detailed model, which uses correlations derived from testing of this uniqu