1、I n tern at i on a 1 J o u r n a 1 of H eat in en ti 1 a tin g, Air-Conditioning and Refrigerating Research HVA Volume 2, Number 2, April 1996 ASHRAE TITLE*IJHVAC 2-2 b W 0759b50 0523L185 920 m International Journal of Heating, Ventilating, Air-conditioning and Refrigerating Research Editor Raymond
2、Cohen. Ph.D P.E., Professor of Mechanical Engineering and Herrick Professor of Engineering. Purdue University, U.S.A. Associate Editors Arthur E. Bergles, Ph.D., P.E., John A. Clark and Edward T. Crossan Professor of Engineering, Department of Mechanical Engineering, Aeronautical Engineering and Mec
3、hanics. Renssclaer Polytechnic Institute. U.S.A. University of Oxford, United Kingdom Fire Research Laboratory, National institute of Standards and Technology, U.S.A. Arthur L. Dexter. D.Phi1 C.Eng University Lecturer, Department of Engineering Science, David k Diion. D.Eng P.E Leader. Thermal Machi
4、nery Group, Building and Ralph Goldman. Ph.D., Senior Consultant. Arthur D. Little. Inc U.S.A. Hugo Hens. Dr.lr., Professor, Department of Civil Engineering, Laboratory of Building Physics, Katholieke Universiteit. Belgium Ken-Ichi Kimura, Dr. Eng., Professor, Department of Architecture. Waseda Univ
5、ersity and President, Society of Heating, Air-conditioning and Sanitary Engineers of Japan, Japan Universitt Hannover, Germany Universit de Lige, Belgium University of Wisconsin-Madison. U.S.A. University of California, Santa Barbara. U.S.A. Horst Kruse. Dr.-Ing Professor, institut fr Kltetechnik un
6、d Angewandte Wrmetechnik. Jean J. Lebrun. Ph.D Professor, Laboratoire de Thermodynaniique. John W. Mitchell. Ph.D., P.E., Professor. Mechanical Engineering, Dale E. Seborg, Ph.D., Professor, Chemical Engineering, Policy Committee Ronald J. Kessner, chair Frank M. Coda Eugene Stamper Fritz W. Steimle
7、 W. Stephen Comstock Raymond Cohen Editorial Assistant Jenny Otlet-Jakovljevic Publisher Frank M. Coda Publishing Director W. Stephen Comstock ASHRAE Editorial and Pubiishing Services Staff Robert A. Parsons. Handbook Editor Scott A. Zeh. hblishing Services Manager Nancy F. Thysell BI996 by the Amer
8、ican Society of Heating. Refrigerating and AirConditioning Engineers. Inc 1791 Tullie Circle. Atlanta. Gwrgia 30329. NI rights reserved. Second class postage paid at Atlanta. Georgia. and additional mailing olfires. HVAC and they can modi$ their behavior through the actuators of the EMCS, as well as
9、 use monitoring and control algorithms implemented in the software of the EMCS to process information. The modem EMCS is a powerful distributed microcomputer system consisting of per- haps hundreds of networked outstations. Many of the latter have the computing power of a high speed workstation or p
10、ersonal computer. The use of sophisticated graphical user interfaces in EMCS supervisor is already commonplace. Future generations of EMCS will probably make use of special-purpose hardware that can directly implement low cost artificial neural networks and the primitive operations of fuzzy logic. S
11、ensors are being developed with built-in local processing power that can be used to enhance and val- idate the sensor readings, and transmit the measurement data to the outstation by radio link or by direct interface to the EMCS network. Open standards have recently emerged for network communication
12、 that should remove many of the hurdles that had previously restricted the transfer of data between the different management and control systems (environmental, fire, security, facilities) in the building. Significant progress has also been made on the development of more sophisticated monitoring an
13、d control algorithms for WAC applications. For many years, researchers have made attempts to apply modem control techniques to building environmental con- trol problems. Auto-tuning and self-adaptive control schemes have been proposed that can automatically modi the controller parameters online. Suc
14、h controllers are capable of commissioning themselves initially, and retuning themselves in response to long term changes in behavior associated with seasonal variations in operating point, or deteriora- tion in plant performance. Model-based controllers have been developed that use predic- tions of
15、 future behavior to determine the best current control action in WAC systems involving significant time delays. Optimal control schemes have been suggested that use online optimization to minimize cost functions that define the desired performance of either the whole building or individual items of
16、its plant, such as chillers or boilers. Sim- ilar optimization schemes have also been developed to ensure optimum scheduling for the maintenance of WAC plant. 105 ASHRAE TITLE*IJHVAC 2-2 76 0757650 O523487 7T3 106 HVACBrR RESEARCH More recently, techniques which mimic human intelligence have been ap
17、plied to var- ious aspects of HVAC plants and controls. Expert systems that make inferences based on qualitative statements about relationships have been used to analyze test data obtained during the commissioning of building control systems before hand-over. Such systems also can evaluate the ongoi
18、ng performance of an WAC plant from normal operating data obtained when the building is occupied. Artificial neural networks can be trained to describe complex relationships or to classi. Enkemann and Arnemann (1994) site several advantages to using a volatile secondary refrigerant, including: highe
19、r volumetric heat capacity, constant temperature evaporation, and lower viscosity. The advantages to using CO, as compared to HFCs include: an extremely low global warming potential, a high enthalpy of evaporation, its non-flammability and non-toxicity, as well as its low cost and availability. Base
20、d on these advantages, the use of CO, as a volatile secondary refrigerant is also proposed here. In addition, CO, has been proposed together with NH, in a cascade system for supermarket refrigeration Kauffeld 1995bl. However, a cascade system, which is only suitable for low temperature applications,
21、 is not considered in this inves- tigation due to the difficulty in controlling such a system, and the availability of CO2 ASHRAE TITLExIJHVAC 2-2 96 0759b50 0523490 298 = VOLUME 2, NUMBER 2. APRIL 1996 1 o9 compression technology for the low pressure stage. The single pressure recirculation system
22、used in the secondary-loop configuration can be implemented using a liquid pump and commercial piping. The use of CO2 as the secondary refrigerant, and ammo- nia as the primary refrigerant, will result in a refrigeration system which uses only nat- ural refrigerants. Refrigerants which occur natural
23、ly are not expected to have any future, unforeseen harmful effects on the environment. One potential drawback to CO, is its relatively low critical temperature (31C. 87.8“F) and high critical pressure (7.375 MPa, 1070 psi). During normal operation, this would not present a problem in the secondary-l
24、oop configuration at low temperature applications around -30C (-22F). For medium temperature applications, typically -15C to -1OT (5 to 15F). system pressures could approach upwards of 3 MPa. Current refrigeration sys- tems are commonly designed to withstand a maximum operating pressure of 2.5 MPa,
25、thereby imposing new design criteria. Designing for such pressures however is not unrea- sonable. The high saturation pressure of CO, presents a much larger problem during the defrost cycle when temperatures in the evaporator approach room temperature. Three possible solutions to this problem were g
26、iven by Enkemann and Amemann (1994): 1. A safety valve that releases the CO2 to the environment when a maximum pressure 2. An expansion container that the CO2 could expand into during system down time. 3. Mixing the CO2 with a liquid solvent. The disadvantages to these alternatives include the need
27、to continually recharge sys- tem (i), the necessary size of expansion container (2), and the decreased performance of CO2 when mixed with solvent (3). Another possibility, proposed here, is to bypass the dis- play case refrigeration coil during defrost and continually provide a small amount of cool-
28、 ing capacity in the evaporator of the primary loop to keep the CO, in the two-phase region. ANALYSIS Figures 1 and 2 shows a schematic of the conventional and secondary-loop refrigera- tion systems. A multistage compression cycle with flash tank (economizer system) is shown in the primary-loop (con
29、ventional and secondary-loop systems). This cycle was simulated by the computer models described here. A multistage cycle is necessary in order to allow the systems to produce the low temperatures required in supermarket dis- play cases. The economizer cycle was chosen in order to reduce the outlet
30、refrigerant tem- peratures from the second compressor as much as possible. The states throughout each of the systems are referenced according to the numbers shown on these schematics throughout the analysis. The secondary-loop system model predicts the power requirements and capacity of the system b
31、ased on user inputs. Inputs to the model include: the air-coil geometry; air-coil evaporation temperature: inlet air conditions to the air-coil: length, diameter, is reached. ._._._. / Expansion Device (4 Figure 1. Conventional refrigeration system ASHRAE TITLExIJHVAC 2-2 96 = 0759b50 0523493 I124 1
32、10 HVAc condensing temperature: the outlet temperature difference of the refrigerant streams at the inter- mediate heat exchanger; and the number of degrees of superheat of the primary and secondary refrigerants at the exit of the intermediate heat exchanger and air-coil. Fig- ure 3 shows an end vie
33、w of a typical air-coil with a triangular tube arrangement. This figure illustrates the inlet air properties and some of the geometrical parameters speci- fied by the user of the model. In addition to those parameters indicated in the figure. the user specifies the length of the tubes, the tube and
34、fin materials (aluminum or cop- per), and the fin thickness. An inline tube arrangement may also be selected. The refrigeration systems shown in Figure 1 were modeled by the procedure outlined in Figure 4. The primary components shown in this flow chart are the refrigeration cycle analysis and the a
35、ir-coil analysis. A description of each follows. Refrigeration Cycle Analysis The refrigeration cycle analysis calculated the thermodynamic properties at each of the states in Figures 1 and 2. The refrigerant mass flow through the air coil was unknown. Therefore, it was necessary to make an initial
36、guess at this quantity to begin the analysis. Through this analysis the heat transfer rates in the vapor and liquid lines were found, as well as the required pump power, and the heat transfer in the intermediate heat exchanger. The air-coil capacity was also calculated (referred to as Qr in Figure 4
37、). This capacity was checked with that computed in the air-coil analysis described below. The difference in coil capacity predicted by these two analyses was used to update the refrig- erant mass flow rate through the air-coil. Pressure drops and heat gains in the interconnecting piping were conside
38、red. The pressure drops were computed by the procedure outlined in Chapter 2 of ASHRAE ASHRAE TITLE*IJHVAC 2-2 96 0759650 0523492 060 VOLUME 2. NUMBER 2, APRIL 1996 Refrigerant Mass Flow Rate 1 . A .“-es Q =Coil capacity calculated by Refrigeration Cycle Analysis Q a=Coil capacity calculated by Au-c
39、oil Analysis - Fan Power 111 Figure 4. Flow chart for conventional and secondary-loop refrigeration systems (1993). The refrigerant temperature change in the vapor line was computed by the fol- lowing equation, provided by Incropera and DeWitt (1990). for heat transfer from a pipe to a constant temp
40、erature fluid: where T, is the evaporation temperature. The value of UAp,o is given by 1 l/hoAp,o+ln-/2nLpk,+ln/2rrLpk+ l/h,.Ap,i D. DO P Di p Do, p UAp,. = with h, computed by the Dittus-Boelter equation and ho a constant value of 6 W/(m2.K) as an approximation for natural convection over a long ho
41、rizontal cylinder. The heat gain was then found from an energy balance on the vapor line. Because heat transfer to the liquid line occurred partially in the two-phase region, the change in refrigerant tem- perature as it flowed from the pump to the air coil was small. This allowed the heat transfer
42、to the liquid line to be approximated by: with T, as the average temperature at states 5s and 4s. Equations (i), (21, and (3). together with the energy equation, define the inlet and outlet states to the vapor and liq- uid lines. The remaining states throughout the systems were found by energy balan
43、ces on the cycle components. The analyses of the compressors, flash tank, condenser, and expansion devices in both models were based on the following assumptions: The outlet states of the condenser and flash tank (states 3, 5. and 7) are saturated. The expansion devices are adiabatic. Pressure drops
44、 in the intermediate heat exchanger and condenser are neglected. Compressor power is computed using an isentropic efficiency. ASHRAE TITLESIJHVAC 2-2 96 0759650 0523493 TT7 112 HVAC heat transfer k thermal conductiviw, kW/(m.K) i. length. m coefficient, kW/(m2.K) I j- fan mass flow, kg/s Q , heat tr
45、ansfer rate, kW i in. insulation AQ heat transfer rate for discrete coil section, longitudina1 kW 11 liquid line R thermal resistance, K.m2/kW S spacing. m T temperature, K V velocity, m/s) W power, kW rn metal, mean o outside, out P pipe, pump r refrigerant s surface U overall heat transfer coeff.,
46、 kW/m2.K) t tubes, transverse REFERENCES ASHRAE. 1988. 1988 ASHRAE Handbook-Equipment. Atlanta: ASHRAE. ASHRAE. 1993. 1993 ASHRAE Handbook-Fimiamentais. Atlanta: ASHRAE. Enkemann, T., and M. Amemann. 1994. Investigation of CO, as a Secondary Refrigerant. In Proc. of IIR Conference. Hanover, Germany.
47、 Godwin, D.S. 1994. Results of Soft-Optimized System Tests in ARIs R-22 Alternative Refrigerants Evaluation Program. In Proc. of International Refrigeration Conference at Purdue University. West Lafayette. IN: Purdue Univ. Gray, D., and R. Webb. 1986. Heat Transfer and Friction Correlations for Plat
48、e Finned-Tube Heat Exchangers Having Plain Fins. In Proc. of the 8th International Heat Transfer Conference 6: Harrison, M.R., R.C. Keeney, and T.P. Nelson. 1995. Pilot Survey of Refrigerant Use and Emissions Hesse. U. 1995. Secondary Refrigerant System Options for Supermarket Refrigeration. In Proc
49、. of Incropera, F.P., and D.P. DeWitt. 1990. findamentais of Heat and Mass Transfer, 3rd ed. New 2745-50. from Retail Food Stores. ASHRAE Transactions lOl(1): 25-33. International CFC and Halon Alternatives Conference. York: John Wiley and Sons, Inc. ASHRAE TITLE*IJHVAC 2-2 96 0759650 0523502 83T VOLUME 2. NUMBER 2. APRIL 1996 121 Kauffeld. M. 1995a. Neue NH3-Technologie-NH3 mit COz als Kaltetrager. DIE KALTE und KLimat- echnik 11: 93 1-932. Kauffeld, M. 1995b. Ammoniak, Kohlendioxid und Wasser-Die gewerblichcn Kaltemittel der Zukunft DIE KALTE und Klima
