ASHRAE IJHVAC 3-1-1997 International Journal of Heating Ventilating Air-Conditioning and Refrigerating Research《供暖 通风 空调和制冷研究的国际期刊 第3卷第1号 1997年1月》.pdf

《ASHRAE IJHVAC 3-1-1997 International Journal of Heating Ventilating Air-Conditioning and Refrigerating Research《供暖 通风 空调和制冷研究的国际期刊 第3卷第1号 1997年1月》.pdf》由会员分享，可在线阅读，更多相关《ASHRAE IJHVAC 3-1-1997 International Journal of Heating Ventilating Air-Conditioning and Refrigerating Research《供暖 通风 空调和制冷研究的国际期刊 第3卷第1号 1997年1月》.pdf（104页珍藏版）》请在麦多课文档分享上搜索。

1、I n t e r n a tio n al Jo u r n al of He at in g ,Ven t il a tin g , Air-conditioning and Refrigerating Research Volume 3, Number 1, January 1997 STD*ASHRAE SRCH IJHVAC 3-L-ENGL 1777 0759650 052b757 2b5 W International Journal of Heating, Ventilating, Air-conditioning and Refrigerating Research Edit

2、or Raymond Cohen, Ph.D., P.E., Professor of Mechanical Engineering and Hemck 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 Engineer

3、ing and Mechanics, Rensselaer Polytechnic institute, U.S.A. Science, University of Oxford, United Kingdom Fire Research Laboratory, National Institute of Standards and echnology, U.S.A. Arthur L. Dexter, D.Phil., C.Eng Reader in Engineering Science, Department of Engineering David A. Didion, D.Eng.,

4、 P.E., Leader, Thermal Machinery Group, Building and Ralph Goldman, Ph.D Senior Consultant, Arthur D. Little, Inc., U.S.A. Hugo Hens, Dr.ir., Professor, Deparunent of Civil Engineering, Laboratory of Building Physics, Katholieke Universiteit, Belgium Ken-Ich Khura, Dr. Eng., Professor, Department of

5、 Architecture, Waseda University 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

6、, Institut fr Kltetechnik und Angewandte Wiinnetechnik, Jean J. Lebrun, Ph.D., Professor, Laboratoire de Thermodynamique, John W. Mitchell, R.D., P.E., Professor, Mechanical Engineering. Dale E. Seborg, Ph.D., Professor, Chemical Engineering, Policy Committee Laurance S. Staples, Jr., chair Frank M.

7、 Coda Hans O. Spauschus Fritz W. Steimle W. Stephen Comstock Raymond Cohen Editorial Assistant Jenny met-Jakovljevic Publisher Frank M. Coda Publishing Director ASHRAE Editorial and Publishing W. Stephen Comstock Services Staff Robert A. Parsons, Handbook Editor Scott A. Zeh, Publishing Services Man

8、ager Nancy F. Thysell 01997 by the American Sociey of Heating. Refrigerating and Air-Con- ditioning Engineers, Inc 1791 Tullie Circk, Atlanta, Georgia 30329. AU rights reserved. Periodicals postage paid at Atlanta, Georgia, and additional mailing oficts. HVAC Ms1 Mo;+Mp;X f. = - - + -x - MPi Ms; + M

9、pi 1 +- Msi If m zones have the same f in a family of n zones, Le., fi = f2 = . = fm, then Mom + MpmX Msl + Mpl Ms2 + Mp2 Msm + MPm - . - - Mol + MplX Mo2 + Mp2X - From the laws of ratio and proportion, m i=l a combined f can be expressed in the following summations. rn rn - i=l i=l c MOi + x c Mpi

10、- rn m c MSi+ c Mpi i=l i=l - Mocomb + MPcombX MScomb + Mpcomb - comb + KcornbX + Kcomb The subscript comb denotes a single imaginary zone that is the combination of the zones whose f values are the same. Equation (6) proves that different thermal zones with the same f can be treated as a single ven

11、- tilation zone. Furthermore, the proof here is also valid for the MSE. In a shutoff VAV system, therefore, different thermal zones with the same z value can be treated as a single zone from the STD-ASHRAE SRCH IJHVAC 3-L-ENGL 1747 0757b50 052b7b4 4T5 VOLUME3, NUMBER I, JANUARY 1997 7 ventilation vi

12、ewpoint. When zoning an air-conditioned building, adjacent rooms with a similar thermal load pattern are usually treated as a single thermal zone for the air distribution system. However, these rooms may have different patterns of ventilation requirements. From an improved indoor air quality control

13、 perspective, it is better not to group rooms with varied venti- lation requirements into a single zone. Thus, not only the patterns of thermal loads, but also the ventilation requirements, have to be considered in zoning. This implies that a designer should zone rooms with similar patterns of both

14、ventilation requirements and thermal loads. For this purpose, z andf, respectively, are useful ventilation indicators for zoning fan-powered and shut- off VAV systems. IMPACTS OF INCREASING PRIMARY AIRFLOW TO CRITICAL ZONE(S) Raising the primary airflow rate to the critical space(s) can reduce the r

15、equired OA entering the AHU when the system is not in 100% OA economizer mode. Although this concept is accepted by the heating, ventilating, and air-conditioning community, no analytical proof has been found. The following derivations remedy the lack of proof. Consider the MSE, Equation (i). It is

16、true that increasing the critical spaces primary airflow lowers the 2 and, consequently, decreases the Y. It is not obvious, however, that increasing the critical spaces primary air Ms, can reduce the required OA Moy, because all X, Y, and Z terms contain Ms, explicitly or implicitly. The same state

17、ment is also valid for the GMSE. To show the trend more clearly, replace the Y and X in the numerator of Equation (3) with their definitions: - - Ms, I+X-F Multiply by Msn on both sides then substitute Equation (4) into F, Mon MOy = 1 +X-F - Mon - Z+K.X 1 +x- l+K - Mon - x-z 1 +- l+K 1+K - - Monl+K+

18、X-Z Take the total derivative of Moy with respect to Ms, noting that X, Z and K are dependent vari- ables of Ms,. Expand the partial derivatives: STD.ASHRAE SRCH IJHVAC 3-1-ENGL 1777 m 0757b50 052b7b5 331 m 8 HVAC - M$)q =1 fi= Mdi where fi = newfi value with new primary airflow rate Ms; = Msi + AMs

19、i, new primary airflow rate of zone i AMsi = increment of primary airflow rate of zone i Msn = C Msj =new total primary airflow i=l Rearranging Equation (22) yields, STD-ASHRAE SRCH IJHVAC 3-L-ENGL 1997 D 0757b50 052b7bd 040 w VOLUME 3, NUMBER I. JANUARY 1997 11 Sum up Ms, of all zones whose primary

20、 airflow rates must be increased to make 6 I 1, The subscriptf 1 denotes the summation of zones needing adjustment. Note that C Msi = C Ms;+ C AMsi fl f1 f 1 and Ms, = Ms,+ C AMsi f 1 The increments are only for zones withfi 1 initially. Therefore, CAMS; = results in f 1 AMs; and Ms, = Ms, + CAMsi (

21、26) Substituting Equations (25) and (26) into Equation (24) yields, This should be the minimum total increment of primary airflow. Substituting into Equation (26) yields, Substituting the above equation into Equation (23) yields the new primary airflow rate for zone i, STD-ASHRAE SRCH IJHVAC 3-1-ENG

22、L 1997 = 0759b50 052b7b9 T? 12 HVAC , must satisfy the GMSE, .e., Mon Mo; = Mo, Ms, + AMs, Mo, l+- - Md, Ms, + AMs, Md, Rearrange, VOLUME 3, NUMBER I, JANUARY 1997 Mo, Msc+MsC Mo, Mo, Md, Ms, +AMs, Md, - Mo; - - 1 +- Ms,+AMs, Mdc Mo, Mo, Ms,+AMs, Mo, Mo Md, - +-1) Ms,+AMs, - Md, Mo,-Md, -+ Ms,+AMs,

23、Mo; Mo, let Md, Mo,-Md, -+ =A Mo; Mo, Substitute into and solve Equation (42) for AMs, AMs, - Ms, 1-1 AMs, = 17 (43) Note that the subscript c may stand for a single critical zone or a group of critical zones. For a group of critical zones, the total increment from Equation (43) can be distributed t

24、o each of the critical zones by using Equation (36). Nevertheless, the use of Equation (36) cannot change the ventilation criticalness hierarchy. If it changes, apply Equation (40) to the group of originally critical zones first. This step determines the size of the increment of total primary airflo

25、w that increases the number of critical zones. Distribute this increment according to Equation (36). Then, apply Equation (43) to the new group of critical zones and determine the extra increment of total primary airflow needed to reach the specified OA quantity. CONCLUSIONS By increasing the primar

26、y airflow rate to the critical zone(s) according to the Multiple Spaces Equation and tempering as necessary to prevent overcooling, the total outdoor airflow rate to the building can be reduced when the system is not in 100% outdoor air economizer mode. This reduction, not necessarily to the lowest

27、value, in the outdoor air requirements is a critical part of optimization strategies aimed at reducing the system total energy consumption, since condition- ing outdoor air expends significant energy. Furthermore, to guarantee that only the minimum increment of primary airflow in VAV systems is appl

28、ied, the ventilation criticalness hierarchy cannot change. For this purpose, Equation (19) provides mathematical proof that the total required outdoor airflow rate will decrease with incremental increases in the primary airflow to the critical zone(s). Equation (6) provides mathematical proof that s

29、paces with identical ventila- tion criticalness can be combined into a single ventilation zone. Equation (27) provides the analytical basis for determining the required primary airflow to underventilated critical space(s) served by series fan-powered VAV systems. Equation (36) pro- vides the analyti

30、cal basis to ensure that the distribution of an incremental increase in the total primary airflow for the group of critical zones will not alter their ventilation criticalness hierar- chy. Equation (40) provides the analytical basis to increment the critical zones primary airflow so that its ventila

31、tion criticalness equals that of the zone that is the next lower in the ventilation criticalness hierarchy. Equation (43) provides the analytical basis for incrementing the critical zones primary airflow so that the outdoor airflow rate is equal to a specified value. STD.ASHRAE SRCH IJHVAC 3-1-ENGL

32、1997 m 0759b50 052b775 280 m 18 HVAC (2) liquid line restriction; (3) leaky compressor valves; (4) fouled condenser coil; and (5) dirty evaporator filter. Only the fault detection and diagnostic steps were considered. Further work would be necessary to establish methods for fault evaluation. One pri

33、mary goal was to identify the sensitivity of the algorithm in detecting and diagnosing each of these faults. Refrigerant loss detection could be an immediate application for the proposed FDD method, since fault evaluation is not necessary for this fault. A refrigerant leak should be repaired as soon

34、 as it is detected and diagnosed. As a step towards implementation, the impact of the number of sensors on the sensitivity for detecting and diagnosing refrigerant leakage was also studied. The performance of FDD methods for this study was evaluated using a combination of simu- lations and laborator

35、y experiments. Diagnostic rules were developed through simulation and checked within the laboratory. The sensitivities of the FDD method for detecting and diagnosing each of the five faults were estimated through simulation. Both simulations and experiments were used to investigate the impact of the

36、 number of sensors on FDD performance for detecting and diagnosing refrigerant leaks. EVALUATION TOOLS For the simulations, a vapor compression system model developed and validated by Rossi (1995) was used. It is a modular, steady-state model that solves mass, energy, and momentum bal- ances for any

37、 set of entering air conditions. A steady-state model is appropriate because the fluid flow and heat transfer dynamics are generally much faster than the dynamics of the load and ambi- ent conditions. The model allows the introduction of all of the faults considered in this study. STDmASHRAE SRCH IJ

38、HVAC 3-L-ENGL 1997 0759b50 052b78L 584 24 HVACBrR RESEARCH 1. Evaporator fouling * 3. Loss of charge tank 8 scale 5. Inadequate compressor performance hot gas bypass 4. Liquid line restriction a. ,!. !. . *valve .o 2. Condenser fouling Figure 3. Rooftop Air Conditioner with Simulated Faults A fully-

39、instrumented, three ton rooftop air conditioner was used for testing FDD algorithms. The system has fixed-speed condenser and evaporation fans, a fixed orifice expansion device, and a single-stage, ordoff controlled reciprocating compressor. Figure 3 illustrates the manner with which faults were sim

40、ulated in the test unit: (1) condenser fouling: paper was placed on the air-side of the coils, (2) evaporator filter fouling: paper was placed on the air-side filter, (3) leaky compressor valves: this effect was represented using a hot gas bypass line with a manual valve, (4) liquid line restriction

41、: a manual valve was located in the liquid line before the expan- sion device, and (5) refrigerant leakage: refrigerant charge removal was controlled by a valve located on the high pressure side of the unit and monitored with a scale. All refrigerant cycle temperature measurements were made using K-

42、type thermocouples mounted to the exterior surfaces of refrigerant piping and insulated. On the air side, tempera- tures of entering and leaving air were measured using platinum resistance temperature devices (RTDs) for both the evaporator and condenser streams. The thermocouples and RTDs were cali-

43、 brated by immersion in a mixed ice bath. The measured error for each sensor was used for an offset correction and to estimate the accuracy of the measurement. The error of all the thermo- couple measurements was less than 0.4 K, while the RTDs were within 0.3 K. The relative humidity of air enterin

44、g the evaporator was determined from measurements of dew point using a chilled mirror dewpoint hydrometer. This device is accurate to within 0.5 K and results in an error in relative humidity that is less than 0.05 for the range of conditions considered in this study. FDD TECHNIQUE Figure 4 shows a

45、block diagram of the fault detection and diagnostic system in relation to input and output “measurements” from a vapor compression cycle. For fixed-speed fans, an air- to-air vapor compression cycle is driven only by the air inlet states into the evaporator and con- denser (U). These states are char

46、acterized by the ambient temperature (Tumb) and the return air temperature (T,) and relative humidity of air stream entering the evaporator. In steady- state, every internal state of a normally operating cycle depends solely on the driving conditions. In this investigation, the performance of the cy

47、cle was characterized using a vector of tempera- ture measurements (Y) only. At most, seven temperatures were considered: (I) evaporating tem- perature (vuy), (2) suction line superheat (Tsh), (3) condensing temperature (Tcond), (4) liquid STD.-ASHRAE SRCH IJHVAC 3-1-ENGL 1797 9 0757b50 052b782 Li10

48、 VOLUME 3, NUMBER I, JANUARY 1997 25 3. +a 3. Tcond 4. Tsc 5. Thg 6. ATa 7. ATea Figure 4. Fault Detection and Diagnostic System line subcooling (TJ, (5) hot gas line or compressor outlet temperature (Thy), (6) air temperature rise across condenser (AT,), and (7) air temperature drop across evaporat

49、or (ATea). The preprocessor portion of the FDD system contains two major components: a steady-state model and the preprocessor portion of a steady-state detector. The measured inputs are used by the model to predict each measured output for the vapor compression cycle under normal oper- ation (.e., no faults). The difference between measured and model predictions of the operating states (residuals) are used by the fault detection and diagnostic classifiers for decision making. A steady-state model is appropriate for the faults in this s