1、I n tern at i on a 1 J o u r n a 1 of H e at in g, Ve n ti 1 at i II g, Air-conditioning and Refrigerating Research HVACRRri.rurh (ISS 1078-9669) is published quarterly as a peer- reviewed archival research journal for the R largely because of their need for ultra-clean environments during manufactu
2、re. This is particularly true when the manufacturing process involves repeated steps that require ultra-clean integrated circuit compatible technologies. Also, the more complex the device, the more times these process steps are likely to be required. Exceptions do exist. There are many cost effectiv
3、e MEMS devices on the market today. A list of the most prolific includes the ink jet cartridge, and the air bag accelerometer, which are low cost because they are manufactured in such large numbers ( lo5) . Other examples can be found in the medical industry, where MEMS devices are used as medical i
4、nsertion sensors in situations for which there is no other way to make the measurements. These devices offer something new, something that is done better than any analogue device could ever do. Do MEMS fit into our industry? If so, where? One thing that HVAC STD-ASHRAE SRCH IJHVAC q-1-ENGL 1778 D 07
5、57b50 05322b7 737 Supply Inlet I VOLUME 4, NUMBER 1, JANUARY 1998 5 Supply Outlet Develop analytical design methods to extend or extrapolate the range of laboratory test data for HPHXs to similar operating conditions; Evaluate the influence of the angle of inclination as a means of controlling the h
6、eat rate of an HPHX at several different operating conditions in both supply and exhaust air streams; and Develop a methodology to design and predict the performance of the heat-pipe heat exchanger for HVAC applications where air properties and flow rates are variables for supply and exhaust air str
7、eams. EFFECTIVENESS The performance of a heat pipe air-to-air heat exchanger is usually expressed as its effective- ness for (1) sensible, (2) latent, and (3) total energy transfer. Referring to Figure 1, the equation that can be used to calculate the effectiveness for the heat pipe air-to-air heat
8、exchanger is pro- vided in ASHRAE Standard 84-1991 as follows: where E is sensible heat, latent heat, or total heat effectiveness; X is either the dry-bulb tempera- ture T for sensible energy effectiveness, or the humidity ratio W for latent energy effectiveness, or the enthalpy h for total energy e
9、ffectiveness; and where rh, is the dry air mass flow rate of the supply; he is the dry air mass flow rate of the exhaust; and mmin is the smaller of h, and me. In order to calculate the sensible, latent, and total energy effectiveness of the HPHX, the mass flow rate, temperature, humidity ratio, and
10、 enthalpy for each air stream inlet and outlet must be known. The parameters required for Equation (1) can be calculated by measuring the atmo- spheric pressure and the temperature, humidity, and pressure drop across the airflow measuring device in each air stream. The following assumptions are made
11、 for the measurement and calculation of effectiveness: Uniform flow and properties entering the HPHX, and well mixed properties at each measur- ing station The HPHX operates under steady state conditions for supply and exhaust air streams during each test (i.e. the apparatus surface temperatures rem
12、ain steady, the mass flow rate for each air stream is steady, and the humidity and temperature distributions at all measuring stations are steady) No auxiliary energy is supplied into the HPHX between the measuring stations by electric motors and/or heaters Figure 1. Effectiveness ratings (counterfl
13、ow) STD-ASHRAE SRCH IJHVAC Li-1-ENGL 1998 0757b50 05322b8 875 6 HVAC.y / V / , / Figure 4. Inclination angle of heat-pipe heat exchanger STD-ASHRAE SRCH IJHVAC Y-1-ENGL 1998 W 0759b50 0532270 423 8 HVAC 045 I) 40 VOLUME 4, NUMBER 1, JANUARY 1998 - - +Sensible Energy e, +Total Energy e, -u- Average S
14、ensible Energy E, +Average Total Energy E, - 8 0.55 t y. 5 .- 0.50 Y - 0 5 0.45 (A 1.40 I .ho 1.W 2.w 2.20 2.40 2.m 2.80 3.00 MassFlux (kg/n/Crnax)e (9) For balanced supply and exhaust flow, m, = me and CrniJCrn, = 1, Equation (9) can be rewritten in the form 1 1 + 1/mu E= where NTU = U,AICrnin is t
15、he number of transfer units, Ur is the overall heat exchanger heat transfer coefficient, A is the heat exchanger surface area, and C is the heat capacity rate for sen- sible heat transfer rates. Thus Equation (10) becomes 1 1 + MC,/( UIA) E= STD-ASHRAE SRCH IJHVAC Li-1-ENGL 1998 = 0757b50 0532277 88
16、8 B VOLUME 4, NUMBER I, JANUARY 1998 15 In this equation, U, need not be independent of mass flow rate h. For plate fin heat exchang- ers, such as the HPHX tested, Ut can be written as (Shah 1986) A where RHp is the heat pipe thermal resistance, Ali,q,A, is the heat pipe supply side air thermal resi
17、stance, and Alh,qtKA, is the heat pipe exhaust side air thermal resistance. A is the reference area for the heat exchanger, and A = A, = A, for equal supply and exhaust side areas for heat transfer. h, and h, are the average heat transfer coefficients on the supply and exhaust sides. qts and q, are
18、the supply and exhaust side overall, or total, fin efficiency, which are expected to be nearly equal for small temperature differences (q, = ql, = qJ. For the HPHX tested, it was expected that the exhaust and supply air thermal resistance would be nearly equal and, for the test conditions of Figures
19、 5 to 10 (Le., a = O), much larger than R, so that Equation (12) became: where hq, is the average heat transfer coefficient and fin efficiency product. Because the airflow in the HPHX is considered turbulent flow between the fins and over the tubes as determined by the maximum Reynolds number throug
20、h the fins (2000 O, r = l/r = m,/me ; that is O aCrirl) the effectiveness decreases rapidly with increasing a. At a slightly larger inclination angle (a = aCrir2), the effectiveness reaches a minimum. For Figure 14 with TI = - 10C acrirl = O and aCNr2 = 3“. For Figure 15 with T, = 40“C, acrit, = O a
21、nd acrit2 = 8.9“. The ratio of minimum to maximum effectiveness, or throttling ratio, is about 0.28 for both high and low inlet supply temperatures, but the angle required to reduce the heat flow to its minimum aCrir2 depends on the supply inlet operating temperature. The drop of the effectiveness w
22、ith increasing inclination angle is very sensitive to a. This implies that accurate inclination tilt angle controls are required to control the heat rate for an HPHX. DESIGN METHODOLOGY The test results for an HPHX can also be used to develop a methodology to design and predict the performance of th
23、e HPHX in HVAC applications. An evaporative water spray cooler should always be considered, since the HPHX cannot transfer moisture, and the tilting method of heat flow control cannot achieve zero heat flow. Otherwise, heating of the supply air occurs for certain STD-ASHRAE SRCH IJHVAC Li-1-ENGL 399
24、8 I759b50 0532283 2117 m VOLUME 4, NUMBER I. JANUARY 1998 19 0.60 0.50 g o.4o o ._ - 8 0.30 o) - e B c% 0.20 0.IC 0.M -10 + Mass Fiux = 1.574 (kgm“2s) +Mass Fiux = 1.846 (kgm“2s) -A- Mass Flux = 2.106 (kglm“2) 4 Mass Flux = 2.378 (kg/m“2) -x- Mass Rux = 2.643 (kgh2s) -5 O 5 Inclination Angle (degree
25、) 10 15 Figure 14. Sensible effectiveness as a function of heat-pipe inclination angle for ms = me, T, = -1O“C, T3 = 23“C, and $3 = 15% rh 0.60 0.50 yt 0.40 o) Y .- CI 8 0.30 a - e Y v1 0.20 0.10 0.00 +MassFlux = 1.574 (kgimA2s) +-Mass Flux = 1.846 (kg/m“Zs) +-Mass Flux= 2.106(kg/mA2s) -Mass Flux =
26、2.378(kg/mA2s) *Mass nux = 2.643(kg/mA2s) IO -8 -6 -4 -2 O 2 4 6 8 IO Inclination Angle (degree) Figure 15. Sensible effectiveness as a function of heat-pipe inclination angle for m, = me, TI = 4WC, T3 = 23“C, and $3 = 15% rh STD-ASHRAE SRCH IJHVAC 4-L-ENGL L998 U759b50 0532282 145 a 20 HVAC,(e) is
27、obtained from the building supply and exhaust airflow rates, and T,(8) and T3(8) are the ambient and exhaust air tempera- tures, respectively. The exhaust air temperatures may be nearly constant for heating (e.g., 24C) and cooling (e.g., 26C) for typical commercial buildings, but the ambient air tem
28、peratures change dramatically over the year. If a yearly energy simulation is performed and the HPHX system is used for both heating and cooling, C,(X) is calculated as the summation of heat rates for each hour for the associated temperature and humidity and appropriate energy utility cost. During t
29、he cooling season, which requires the chilling of supply air, Equation (22) uses the elec- trical energy cost EC and chiller coefficient of performance r while the heating season uses sea- sonal heating costs for natural gas and electricity and boiler efficiency r. The equation for the operating cos
30、t, C,(X), can be described by where qloud(8) is the ventilation heating or cooling load or rate without the energy recovery sys- tem, EC is the electrical utility energy cost, rF is the fan efficiency, Pfun-s is the increase in sup- ply fan power caused by the HPHX, and Pfan-e is the increase in exh
31、aust fan power caused by the HPHX. In order to avoid excessively high fan power costs, the HPHX should be sized for the largest ventilation flow rates (Johnson et al. 1995). For commercial buildings, the ventilation heating load tends to increase in a linear manner as the ambient temperature drops b
32、elow about 15C. The HPHX can be used to meet all of this load STD-ASHRAE SRCH IJHVAC 4-1-ENGL 1998 E 0759b50 0532284 T1A E HVAC&R RESEARCH 22 if it has a high efficiency, but below - 15“C, frost growth on the exhaust side of the HPHX may occur. If the HPHX is throttled back using the tilt angle cont
33、rol, frost growth on the heat exchanger can be minimized. During cold weather, air filters in the supply, upstream of the HPHX, should be removed to avoid the blockage of the filter by frosting. The cooling load for ventilation air depends, in part, on the ambient air humidity. For very dry climates
34、, this cooling load curve tends to be linear but for warm moist climates, the loads will change on a daily basis with ambient air humidity. A high effectiveness HPHX may meet a large fraction of the ventilation cooling load requirements but most often some auxiliary supply air chilling will be requi
35、red, especially for large fractions of return airflow (RA). Reheating of the supply air after the cooling coil (CC) may be required if the air in the cold deck is to be at 15C. The capital costs, C,(X), will include the HPHX costs, the extra fan costs, the reduced costs of boiler and chiller equipme
36、nt, the costs of installation, and the costs of instrumentation for the system. These cost estimates should be based on the manufacturers cost data and installation cost estimates. The best design for an HVAC system with an HPHX should be the lowest level life-cycle cost in the system considering th
37、e full range of operating conditions. In order to get the minimum life-cycle cost for a HVAC system with an HPHX, it is evident from Equation (22) that E, should be large so that the ventilation heating or cooling loads are kept small. Johnson et al. (1995) and Dhital et al. (1995) showed that the l
38、east LCC of run-around air-to-air heat exchangers vary nearly inversely with the heat exchanger effectiveness when the effectiveness is close to 50%, implying that the effectiveness should be high. However, effectiveness greater than 70% is not likely to be cost effective for office buildings in mod
39、erate and cold climates, even when all the heating coil design variables are optimized. A similar result can be expected for an HPHX in an office build- ing, because it is a sensible energy exchanger device similar to the run-around system. The influence of inclination angles a on E, is very large.
40、A positive inclination angle (the heating zone of the heat pipe is higher than the cooling zone) may cause E, to decrease sharply. This means that an improper installation of the HPHX in an HVAC application may result in both too low and too high rates of energy recovery unless the controls are corr
41、ectly set. Thus, the HPHX must be accurately installed in HVAC applications if the expected potential operating cost savings are to be realized. CONCLUSIONS In this paper, ASHRAE Srundurd 84-1991 was used as a guide for testing an HPHX. Several changes were introduced to maintain high accuracy in th
42、e computed effectiveness. The most important addition was the calculation of uncertainty according to ASME PTC 19.1 for all effec- tiveness results. It was found that the HPHX effectiveness, and corresponding uncertainties var- ied with the mass flow rate of air, average air temperature, mass flow r
43、ate ratio of supply and exhaust air, and tilt angle of the heat-pipe. Operating conditions that would result in condensa- tion effects were avoided during testing. The main conclusions from this work are: i. For the HPHX tested, the average sensible and total effectiveness decreased with increasing
44、flow rate but increased with average air temperature and mass flow ratios that were both less than and greater than 1 .O. The corresponding uncertainties in effectiveness decreased with flow rate from 7.4% to 2.2% for cold supply air temperatures and from 8.2% to 3.2% for hot humid supply air. Effec
45、tiveness results with uncertainties larger than 5% should not, accord- ing to ASME PTC 30- 199 1, be reported. 2. Measured data was used to demonstrate the accuracy for extrapolating a particular effective- ness value to other operating conditions. The differences between the measured results and th
46、eoretical sensible effectiveness were within t2%, with a variable, overall heat transfer coefficient (,), over the range of airflow rates investigated. These results imply that the STD-ASHRAE SRCH IJHVAC 4-2-ENGL 1998 0759b50 0532285 954 M VOLUME 4, NUMBER I. JANUARY 1998 23 sensible effectiveness o
47、f the HPHX can be extrapolated within accepted uncertainty limits as a function of m, m,/m, T, and CY. The equations for these extrapolations can be used in HPHX system design as a empirically validated design model. Fewer tests will be needed to establish the effectiveness of an HPHX system when em
48、pirical extrapolation equations are available and, for large HPHX, empirical extrapolation equations may be necessary because there is no test data except those for small HPHXs. Negative heat pipe inclination angles, with the hot end down (u acNrl), the effectiveness decreases rapidly with increasin
49、g a. For u acrir2, the effectiveness again appears to be constant but at a low value. For cold supply air, acrifl = O and acrir2 = 3“. For hot supply air, acritl = O and ucrit2 = 9“. These results indicate that the control of the inclination angle for HVAC applications must be done accurately if the full operating cost benefit of an HPHX is to be realized. The effectiveness, from measurements and/or theoretical extrapolations for the HPHX, can be used to develop a methodology to design and predict the performance of the HPHX in HVAC applications with air pro
