1、4725 Minimizing TEWI in a Compact Chiller for Unitary Applications Philip R. Barnes ABSTRACT A simulation model was developed to investigate strate- gies for reducing total equivalent warming impact (TEWI) in compact water chillers. The focus was on minimizing refrig- erant charge while increasing e
2、jciency, using R-41 OA as an example. Compactflat plate heat exchangers with refrigerant channels similar in scale to microchannels appear capable of reducing total system charge by about 80% compared to conventional air-air split systems. Results are also compared to those obtainedfor highly ejcien
3、t air-to-air unitary systems, in which minimum-TEWIdesign strategies require larger heat exchangers having greater chaTe. Overall, the two approaches achieve comparable reductions in global warming impacts; the chiller depends more on reducing direct emis- sions, compared to unitary systems dependen
4、ce on reducing indirect emissions through use of jlat multiport tubes with foldedjns. INTRODUCTION The phaseout of hydrochlorofluorocarbons (HCFCs) has led to their replacement by hydrofluorocarbons (HFCs), which are greenhouse gases. To characterize the global warm- ing effects of new systems, incl
5、uding those using natural refrigerants, Sand et al. i proposed the total equivalent warming impact (TEWI) to account for the release of refrig- erant into the atmosphere (direct effects) and the release of carbon dioxide from electricity generation (indirect effects). Many alternatives to HCFCs are
6、not greenhouse gases but may be either toxic or hazardous (e.g., hydrocarbons, ammo- nia), making it necessary to minimize charge when those refngerants are used. Charge-minimizing heat exchangers and systems differ substantially from conventional designs, and Clark W. Bullard, Ph.D. Fellow ASHRAE p
7、rototypes are costly to fabricate. Therefore, it is economical to employ simulation modeling to conduct initial explorations of the parameter space, quantifying trade-offs between two ways for minimizing TEWI-increasing efficiency and reduc- ing charge. Previous analyses have provided rough comparis
8、ons of TEWI for residential space conditioning (heating and air conditioning combined) systems, including the option of using a secondary (e.g., chilled water or glycol) loop to keep flammable or toxic refrigerants outdoors i. While establish- ing that the direct contribution to TEWI is a small frac
9、tion of the indirect, they were not focused on the geometric trade-offs involved in designing low-charge heat exchangers, for either air-to-air systems or chillers. This paper addresses those geometric trade-offs explicitly and examines their effect on the directhndirect fraction of TEWI. The next s
10、ection describes two residential-scale “base- line” systems from which improvements can be measured. One is a conventional North-American-style split system; the other is a hermetic chiller utilizing compact brazed plate heat exchangers. The residential scale was chosen only to provide a familiar st
11、arting point for the analyses, which are normalized per unit cooling capacity, in the interest of generalizing the resulting insights across a broader range of unitary air-condi- tioning system and chiller sizes. The next sections briefly describe the optimization process and explain the results of
12、the TEWI minimization and then describe the sensitivity analysis performed near the opti- mum with respect to several model assumptions. Descriptions of the baseline systems and the methods for modeling the brazed plate heat exchangers were detailed by Barnes and Bullard 2. Philip R. Barnes is a mec
13、hanical design engineer at Florida Power 4%/yr refrigerant leak rate 5; and 900 h/yr runtime (typical of Washington, D.C., Indianapolis, Los Angeles) 6. For the conventional system with -258 g/kW refrigerant charge i, this corresponds to a 90/10 indirecvdirect TEWI fraction, which, of course, can va
14、ry greatly with assumptions about system runtime (climate, insulation, and shading) and refrigerant leakage. While the following analyses illustrate the methodology for the most common case of HFC refrigerant R-410A where flammability and toxicity are generally not at issue, they also show how the p
15、resence of charge constraints could influence strategies for minimizing TEWI by limiting heat exchanger size. Using the same model, Kirkwood and Bullard 7 explored the extent to which TEWI could be reduced in air-to- air split systems by using microchannel heat exchangers. They examined microchannel
16、s because of their compactness for a given heat transfer capacity and low pressure drop, compared to traditional round-tube/plate-fin heat exchangers. Their simulations suggested that TEWI could be reduced by approx- imately 13% compared to a conventional R-410A system, at the AR1 210/240-B standard
17、 rating condition (26.7“C dry bulb/19.4“C wet bulb indoor; 27.8“C dry bulb outdoor found in DOE 6). This improvement was achieved by increasing system COP (3.8 to 43, which required large heat exchang- ers, thus limiting charge reduction (258 to 235 gkW). There remains some potential for improvement
18、 in the microchannel design either by extruding smaller microchannel ports or decreasing liquid line length (10.8 m in Kirkwoods simula- tions) to reduce total charge. However the overall result (diffi- culty of reducing charge while increasing efficiency) is not surprising; due to the dominance of
19、air-side heat transfer resis- tance, large heat exchangers are needed and refrigerant charge increases proportionally. Compact Hermetic Chiller Another way to minimize the direct TEWI effect would be to minimize charge by building a small chiller to take advantage of the compactness obtainable with
20、refrigerant-to- water heat exchange instead of refrigerant-to-air. Commer- cially available compact brazed plate heat exchangers (CBEs) are used in a wide variety of applications (food processing, chemical processes, and pharmaceutical industries). Due to their very compact nature, high surface-volu
21、me ratios, rel- tively low pressure drops, and their ability to utilize chevrons and bumps imprinted on the plates, they rely more on heat transfer coefficient and less on area to transfer heat. Many studies have examined liquid-liquid heat transfer and pressure drop in CBEs 8-18. However, only a fe
22、w have examined evaporation and condensation in CBEs 19-23, with only Yan 22,23 providing correlations. Those correla- tions for two-phase heat transfer and pressure drop were used in a simulation model with correlations from Shah and Focke I51 for single-phase heat transfer and Focke et al. i i for
23、 single-phase pressure drop. Criteria for choice were described by Barnes and Bullard 2. The compact chiller using CBEs was simulated for the ARI 550/590 standard rating condition for chillers (0.054 L/s per kW at 29.4“C inlet condenser water, 0.043 LIS per kW at 6.7“C outlet evaporator water) 24. I
24、t was assumed that refrig- erant lines could be quite short, as indicated in Table 1. To Table 1. Hermetic Chiller Model Inputs 336 ASHRAE Transactions: Research Conventional CBE Chiller provide chilled water to every room, total pipe lengths were conservatively assumed to bel 50 m and 50 m for the
25、cold and hot loops, respectively. Outdoor fan power requirements were assumed identical to the split system (1 9 WkW) and halved indoors (20 WkW) due to the absence of ductwork and the ability to use thinner coils with larger face area, such as those typically found in ductless split systems. Superh
26、eat was set to 7“C, as recommended by a manufacturer of compact brazed plate heat exchangers 25. Subcooling was set equal to 2C in order to minimize the amount of liquid in the condenser while still ensuring full condensation. The compressors isentropic efficiency was assumed to be 0.7, and its UA =
27、 15.8 WK 2. Water pumping power was based on a combined pumpimotor efficiency of 0.6 26. The chiller requires more compressor power due to the difference in standard rating conditions. As shown in Table 2, the split system condenses at 39C and evaporates at 9C. The CBE system with 60 plates in both
28、the condenser and evapo- rator (denoted 60 x 60) has average condensing and evaporat- ing temperatures of 37C and 5“C, respectively. This increases compressor power to 208 W/kW versus 198 W/kW for the conventional split system. The two water pumps add 9 WkW that did not exist in the conventional spl
29、it system. However, due to the absence of indoor ductwork, total power require- ments are about equal. Flat Plate Chiller Conventional Split System MINIMUM-TEWI CHILLER Initial analyses suggested that the optimal heat exchanger geometries would lie well outside the range of channel aspect ratios use
30、d to develop the correlations for the chevron plate CBEs. The CBE model was therefore modified by replacing the heat transfer and pressure drop correlations, as described in Table 3. These correlations were developed for flow in smooth tubes and applied here to rectangular channels using the hydraul
31、ic diameter 27-321. The heat flux and mass flux values were within the range of the correlations. The optimization results tended to favor diameters smaller than those for which the correlations were developed, but previous experiments with R-4 10A in microchannels suggest that extrapolation errors
32、are small 33. TEWI was minimized by iteratively maximizing COP, then minimizing charge using variable metric searches. The search algorithm thus identified geometric parameters and approach tempera- ture differences characteristic of near-optimal heat exchang- ers. Surprisingly, a substantial charge
33、 reduction was achieved at plate spacings small enough to inflict a condenser pressure drop penalty that matched the temperature glide, offsetting it by adding area to reduce approach temperature difference to almost O. 1 OC, as shown in Figure 1. The minimum-TEWI condenser has many long narrow plat
34、es (Np = 200, L = 3.8 m, w = 1.9 mm) spaced closely together (b,= 0.4 mm), as shown in Table 3. The search algo- rithm pointed toward a global optimum having an infinite number of short plates spaced infinitesimally close together, CBE 60 x 60 so these parameters were constrained arbitrarily to 200
35、and 0.4 mm, respectively, based on manufacturing considerations. nel only 25% higher than the conventional split system. To accommodate the increased heat flux at reasonable Dressure The compact design of the CBEs, and the shorter liquid line in the packaged ystern, requires leSS charge (loo gkw The
36、 refrigerant channels create amass flux in each chan- s. 258 gkw)* system than the conventional Split system but still 7% higher TEW1 is 7% leSS for the than the more energy-efficient microchael split system. Since CBEs were developed for liquid-liquid applications, performance could be improved by
37、optimizing for refrigerant evaporation and condensation heat transfer, as described below. drop, water-side plate spacing must be 4.6 mm, resuiting in a water mass flux 3.9 times higher than the conventional CBE. CBEs have many chevron bumps that restart the laminar boundary layer to create high wat
38、er heat transfer coefficients (-9500 W/m2.K) at low Reynolds numbers (-660). Since no ASHRAE Transactions: Research 337 65 60 55 50 a 5“ 45 40 35 30 12 10 .-. o L I- 8 6 1 I I I o 1 2 3 4 Length m O 0.1 0.2 0.3 0.4 0.5 0.6 0.7 Length m Figure 1 Condenser temperature profile. published correlations w
39、ere available for smaller plate spac- ing, the current investigation used correlations for smooth flat plates very closely spaced, relying on turbulent flow (Re N 2 100) to achieve higher heat transfer coefficients (-4750 W/m2.K) than could be obtained with laminar flow between smooth plates. Althou
40、gh the minimum-TEWI chiller has a condenser 3.8 m long (compared to 0.4 m for the CBE) and higher Reynolds number, water-side pressure drop is 8.6 kPa, while the baseline CBE water pressure drop is 12 kPa due to the chevron bumps. The minimum-TEWI evaporator is much more sensitive to refrigerant pre
41、ssure drop than the condenser, despite the fact that it is much shorter than the condenser (0.69 m vs. 3.8 m). The evaporator also tended toward the minimum plate spacing on the refrigerant side, 0.4 mm, but the plates at 19 mm are much wider than the condenser, (yet narrower than conven- tional CBE
42、 width of 71 mm). The evaporating temperature could not increase above 5C due to the low water temperature required for the secondary loop and the 7C superheat required by the TXV that was assumed to be present. See the temper- ature profile in Figure 2. Despite the lower evaporating temperature (5O
43、C vs. SOC), the chillers compressor power was nearly identical to the conventional air-air split system, due to its lower average condensing temperature (34C vs. 39C). As shown in Table 4, total water and air pumping power for the optimal chiller was 2 1% lower than for the conventional split system
44、. Due to the 82% charge reduction, the TEWI is 13% less than the conventional split system. At 47 g/kW, charge is about 50% lower than CBE baseline chiller. Less than half of this charge is in the heat exchangers; the rest is in the refrigerant lines or dissolved in the compressor sump. Figure 3 com
45、pares the charge distribution to that ofthe baseline aidair system, which, in turn, is essentially the same for the microchannel aidair system. Figure 2 Evaporator temperature profile. The optimally sized condenser is also very compact. It would be 3.8 m long, but only about 2 mm thick and about 0.6
46、 m wide (allowing 0.2 mm channel dividers), as shown in Figure 4. Such a plate might be made of copper, steel, or aluminum and bent to fit into a more compact package, Ends could be cut to accommodate brazed or welded fittings that would separate the refrigerant and water channels at the inlets and
47、outlets. Of course, there are many other ways of arranging proximate refrigerant and water heat transfer surface in a more manageable package, but this simple geom- etry is sufficient to illustrate the overall size and trade-offs involved. It could be easily rolled into a very compact cylin- drical
48、spiral standing 1 m tall. TRADE-OFFS NEAR THE OPTIMUM Standard Operating Conditions The results presented above described a chiller geometry in which TEWI was minimized at the maximum COP achiev- able at the given operating conditions. The simulations showed how another 28% reduction in charge could
49、 be obtained-if necessary to meet safety constraints-at a cost of 7% reduction in COP. Figure 5a shows minimum charge trade- offs at various COP values, while Figure 5b shows the asso- ciated plate width and length trade-offs for the evaporator and condenser. As COP decreased from the maximum of 3.96, charge was minimized by decreasing condenser length with little change to condenser plate width and by simultaneously decreasing evaporator width with little change to evaporator plate length, as shown in Figure 5b. For both heat excha
