1、Volume 3, Number 2, April 1997 American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. STD.ASHRAE SRCH IJHVAC 3-2-ENGL 1777 757b50 0527b23 977 I International Journal of Heating, Ventilating, Air-conditioning and Refrigerating Research Editor Raymond Cohen, Ph.D., P.E., Profe
2、ssor 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 Mechanics, Rensselaer Poly
3、technic Institute, U.S.A. Science, University of Oxford, United Kingdom Fire Research Laboratory, National Institute of Standards and Technology, U.S.A. Arthur L. Dexter, D.PhiI., C.Eng., Reader in Engineering Science, Department of Engineering David A. Didion, D.Eng., P.E., Leader, Thermal Machiner
4、y Group, Building and Ralph Goldman, Ph.D., Senior Consultant, Arthur D. Little, Inc., U.S.A. Hugo Hens, Dr.Ir., Professor, Department of Civil Engineering, Laboratory of Building Physics, Katholieke Universiteit, Belgium Ken-Ichi Kirnura, 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
6、und Angewandte Wrmetechnik, Jean J. Lebrun, Ph.D., Professor, Laboratoire de Thermodynamique, John W. Mitchell, Ph.D., P.E., Professor, Mechanical Engineering, Dale E. Seborg, Ph.D., Professor, Chemical Engineering, Policy Committee Laurance S. Staples, Jr., chair Frank M. Coda Hans O. Spauschus Fri
7、tz W. Steimle W. Stephen Comstock Raymond Cohen Editorial Assistant Jenny tlet-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 Manager Nancy F. Thysell O199
8、7 by the American Society of Heating, Refrigerating and Air-Con- ditioning Engineers, Inc., 1791 Tullie Cucle, Atlanta, Georgia 30329. All rigbu reserved. Periodicals postage paid at Atlanta. Georgia, and additional mailing offices. HVAC nor may any pari of this book be reproduced, stored in a retri
9、eval system. or uansmitted in any form or by any means-elecuonic, photocopying, recording, or other-without permission in writing from ASHRAE. Abshcts-Abstracted and indexed by Engineering Information, Inc. Available electronically on Compendex Plus and in print in Engineer- ing Index. Disclaimer-AS
10、HRAE“ has compiled this publication with care, but ASHRAE has not investigated, and ASHRAE expressly disclaims any duty to investigate. any product, service. process. procedure, design, or the like which may be described herein. The appearance of any techni- cal data or editorial material in this pu
11、blication does not constitute endorsement, warranty, or guaranty by ASHRAE of any product, ser- vice, process, procedure. design, or the like. ASHRAE does not warrant that the information in this publication is free of errors, and ASHRAE does not necessarily agree with any statement or opinion in th
12、is publica- tion. The entire risk of the use of any information in this publication is assumed by the user. Postmaster-Send form 3579 to: HVAC and (Ta - Tb) is the average temperature difference of the two fluids. The sec- ond equation applies to the thermal capacitance of either fluid where c(T, -
13、To,) is replaced by latent heat, 1, if the fluid is in the two phase state. The two equations are of course interdepen- dent in that the quantity of heat transferred is equal to that gained or lost by the respective fluid streams. Also the Ta and Tb are determined by the respective fluids? Tin and T
14、o, values. For a given q, one could decrease the heat exchanger material by increasing the average tem- perature difference, (Ta - Tb), but that would inherently increase the entropy production of the process and thus decrease the system COP. The preferred solution is to increase the overall con- du
15、ctance per unit length. This solution is being used at an increasing rate, primarily through the use of enhanced heat transfer surfaces. These enhanced surfaces might use microchannels, vari- ous other structured surfaces, or reentrant cavity surfaces, etc. on either or both the interior and exterio
16、r of the heat exchanger tubes. These changes spark the need for new heat transfer coeffi- cient and pressure data, representing not only the new surfaces in the multitude of heat exchanger arrangements but also with the new refrigerants and lubricants. The combination of a refrigerant in a low quali
17、ty state with the liquid lubricant entrained in the grooved surface requires more research before optimization is possible. The air side of heat exchangers is also 99 STD.ASHRAE SRCH IJHVAC 3-2-ENGL 1777 0757b50 0527b25 7bL 100 HVAC the thermal energy required was subtracted from the energy release
18、of combus- tion. The species concentrations and enthalpy of the gas mixture were updated at each time step and the effective disappearance of the intermediate species CO determined the end of flame zone. The combustion chamber was modeled as an incompressible, plug flow zone with twenty dis- crete e
19、lements of equal volume. A typical residence time in the combustion chamber was 0.35 seconds, roughly the same as in the main heat exchanger. No chemical reactions were consid- ered, excepting the formation of thermal NO,. As the energy change in this reaction is small, NO, formation was calculated
20、from Equation (7) by post-processing, using the temperatures obtained from heat transfer calculations. As the flue gas normally enters the heat exchanger at a temperature less than lOOO“C, further chemical reactions were not accounted for in the heat exchanger. This model was found to give good agre
21、ement with the measured values in a commercial boiler testing program (Teekaram 1994) for the rated operating point and predicted correctly the trends of NO, concentrations under varying operating conditions. The alternate modeling method is applicable to both oil- and gas-firing situations, and was
22、 based on an empirical equation in the form of where NO, is the NO, concentration (ppm wet), O2 is the volume percentage of free oxygen in the flue gas (% wet) and Tfthe adiabatic flame temperature (K) calculated for the operating con- ditions of the boiler. The regression constants m and c were eva
23、luated from experimental data for oil- and gas-firing, respectively (Teekaram 1994). This approach has the advantage of computational simplicity; also, it provides reliable predic- tion for the NO, formation over the normal range of boiler operation conditions. However the drawback is that, for stoi
24、chiometric or rich combustion, Equation (1 1) predicts no NO, as the residual oxygen tends to zero. In addition, it does not predict the change of NO, formation with firing rate, since the flame temperature and the concentration of residual oxygen are independent of the firing rate. CO Formation. Th
25、e most significant influences on the formation of carbon monoxide in the flue gases are imperfect mixing of the fuel and air (properties of the burner head) and rapid quenching of pockets of combustion chamber gas near cool surfaces (determined by the dimen- sions of the combustion chamber). It is n
26、ot feasible to describe these effects deterministically using a one-dimensional model, hence an empirical approach was used. The basis of the approach was the equilibrium value CO, at the mean combustion chamber temperature, corrected for the relative sizes of the flame and the combustion chamber. C
27、onsider- ation of boiler test data produced a relationship of the form STD.ASHRAE SRCH IJHVAC 3-2-ENGL 1777 0757b50 0527b3L Tb5 106 HVAC typical operating conditions are within the range 15 to 25%. The variation of NO, emissions with firing rate depends on burner design-for a typical blast tube burn
28、er, the curve is rather flat. Figure 4 shows that, at typical operational levels with excess air, the model predicts a slight increase in NO, levels as the firing rate is reduced. As conditions in the flame zone are substantially unchanged, the prompt NO, remains constant. Despite the smaller size o
29、f the post-flame high temperature zone, the longer residence time in this region leads to a higher predicted level of thermal NO,. Incorporation into System Simulation To illustrate the feasibility of the use of this model in the context of a system simulation, the model was rewritten in the form of
30、 a TRNSYS program. The program contained 3 input variables, 7 output variables, and 39 parameters. To enable the steady-state boiler model to be used in a sys- 108 HVAC-Tg dt boiler -_- dT, TG-T, - -_- dt boiler In Equations (14) and (15), the time constant rboiler = RCboiler, where Cbojler is the b
31、oiler ther- mal capacitance (metal + water, typically 1025 WIK) and R the overall thermal resistance between the gas and water. The superscript o denotes the steady-state value. For simplicity, the transport delay was not taken into account. STD-ASHRAE SRCH IJHVAC 3-2-ENGL 1777 W 0757b50 0527b3Li 77
32、4 U VOLUME 3, NUMBER 2, APRIL 1997 109 An additional thermal capacity was incorporated in the system model to take into account the dynamics of the connected load. The boiler-load system model (Figure 5) enabled the boiler to be subjected to an arbitrary load pattern as a function of time. I Boiler
33、- I A I .1 Load L I System thermai capacity I I I I Figure 5. Boiler-load system The prototype boiler modeled in this case was a gas-fired unit with a rated output of 190 kW (Strebe1 Ltd. 1995). The boiler had a two stage burner, operating on high- and low-fire, and con- trolled according to the req
34、uired throttling range of the water flow temperature. Two scenarios were modeled: Start-up from an initial system temperature of 24“C, with the thermal load assumed to be in proportion with Tw Operation with a fixed load of 170 kW over a period of 4 hours. The magnitude of this load fell between the
35、 high- and low-firing rates of the boiler. c Time (hour) O 0.5 1 t.5 Time (hour) Figure 6. System start up The results of the simulation of the start-up phase are shown in Figure 6. This scenario was chosen as the conditions allowed a direct comparison with a single dynamic test carried out as part
36、of a commercial boiler testing program (Teekaram 1994). The test showed the following characteristics: Very rapid attainment of steady-state values of species concentrations Rapid initial rise in flue gas temperature followed by a first-order rate of temperature increase First-order temperature rise
37、s of both flow and return water temperatures STD*ASHRAE SRCH IJHVAC 3-2-ENGL 1797 0757b50 0527b35 bOO 110 HVAC however, it can detach from the ceiling if its negative A. Vahab Hassani is senior engineer at the National Renewable Energy Laboratory, Thermal Systems Branch,Golden, CO. Alian Kirkpatrick
38、 is a professor of Mechanical Engineering in the Solar Energy Applications Laboratory, Colo- rado State University, Fort Collins, CO , and Pasquale M. Sforza is a professor in the Department of Mechanical, Aero- space, and Manufacturing Engineering in the Polytechnic University, Brooklyn, NY. 112 ST
39、D-ASHRAE SRCH IJHVAC 3-2-ENGL 1997 m 0757b50 0527b38 3LT VOLUME 3, NUMBER 2, APRIL 1997 113 Wail mixing layer-like (a) k- x, -j/Separation point (b) Figure 1. (a) Turbulent wail jet with positive influence of buoyancy forces, (b) Turbulent wail jet with negative influence of buoyancy forces buoyancy
40、 is greater than the positive pressure difference. Therefore, jet separation can become a problem in buildings applications as relatively high velocities and temperature differences result in unacceptable thermal comfort levels in the occupied zone. Consequently, predicting nega- tively buoyant jet
41、behavior, and particularly its separation, is an important engineering issue. The published literature on turbulent isothermal wall jets is extensive. The research in this area has been directed toward determination of the various jet decay regimes, and associated large scale characteristics, such a
42、s the jet decay coefficients and growth rates. Glauert (1956) developed analytical models for isothermal wall jets for both laminar and turbulent conditions. Viets and Sforza (1966) discussed classification of the jet decay regimes into three distinct regions. A review study comparing various experi
43、mental findings related to three-dimensional wall jets is contained in Launder and Rodi (1983). Sforza and Herbst (1970) measured centerline velocity decay, as well as lateral and vertical growth for turbulent jets issuing from rectangular outlets. Centerline velocity decay measurements from isother
44、mal radial jets were also presented by Engle and Kirkpatrick (1993). The velocity distribution in the transverse directions for an isothermal wall jet were measured by Padmanabham and Gowda (1991). Additional wall jet ref- erences are also introduced in appropriate sections of this paper. Research r
45、esults concerning the separation of non-isothermal, especially negatively bouyant wall jets, are more limited. Previous work for horizontal wall jets has been reported by Grimitlin (1970), Rodahl (1977), Nielsen and Moller (1988), and Kirkpatrick and Hassani (1993). Empiri- cal expressions have been
46、 developed which relate the separation length to the air jet conditions at the nozzle outlet. The relations have usually been presented as functions of the Richardson number or Archimedes number at the nozzle outlet. STD-ASHRAE SRCH IJHVAC 3-2-ENGL 1997 I0759L50 U527b39 25b HVAC ax ay az p ax ;*i) (
47、 ax ay aZ u-+ v- + w- = - - - + v - - -u2 +-u/+ -u w The z-momentum equation is not needed for the integral analysis pursued here. An order of magnitude analysis (following Hinze 1975) was carried out for the y-momentum equation. It indicated that the term a(m)/az, which appears in the y-momentum eq
48、uation may be neglected if it is assumed that the RyzLJLz called the “Core Zone” (CZ) or “Potential Core” near the jet outlet where the maximum velocity is equal to the outlet velocity; (2) a “Characteristic Decay” (CD) region or a transition zone (zone 2) where the centerline velocity decay of the
49、jet is given by: rn 1 - u0 J., and (3) a “Radial Decay” (RD) region or an extensive zone (zone 3), called the fully established turbulent zone, where the centerline velocity decay is: rn 1 u, x+x -oc- P (9) The parameter xp is the distance from the nozzle to the virtual origin of the wall jet. Experi- mental observations suggest that in the CD and RD regions of a wall jet, the temperature differ- ence and velocity profiles can be approximated as self-similar in both the y and z directions when normalized by the local centerline velocity and temperature differen
