ASHRAE AN-04-4-1-2004 Wheel Selection for Heat and Energy Recovery in Simple HVAC Ventilation Design Problems《在简单的暖通空调通风设计问题上的热量和能量回收车轮选择》.pdf

《ASHRAE AN-04-4-1-2004 Wheel Selection for Heat and Energy Recovery in Simple HVAC Ventilation Design Problems《在简单的暖通空调通风设计问题上的热量和能量回收车轮选择》.pdf》由会员分享，可在线阅读，更多相关《ASHRAE AN-04-4-1-2004 Wheel Selection for Heat and Energy Recovery in Simple HVAC Ventilation Design Problems《在简单的暖通空调通风设计问题上的热量和能量回收车轮选择》.pdf（18页珍藏版）》请在麦多课文档分享上搜索。

1、AN-04-4-1 Wheel Selection for Heat and Energy Recovery in Simple HVAC Ventilation Design Problems Yaw Asiedu, Ph.D. Robert W. Besant, P.Eng. Fellow ASHRAE Associate Member ASHRAE Carey J. Simonson, Ph.D., P.Eng. ABSTRACT This paper shows how air-to-air heat and enera system design problems can be fo

2、rmulated for a simple HVAC conjg- uration andsolved for the least life-cycle cost system while still retaining a small payback period. Mathematical expressions anddesign tables arepresented to facilitate the design process. The design process is illustratedfor the city of Chicago where both large he

3、ating and cooling loads occur in HVAC applica- tions. The example design problem presented shows that paybackperiods of less than one year are often achieved for energy wheels and sometimes for sensible heat exchangers. The life-cycle cost savings for auxilialy heating and cooling ventilation air ja

4、r exceeds the capital cost of the energy exchanger (sensible or total), even when only a ten-year life cycle is considered. INTRODUCTION In the past, energy recovery systems in buildings have been designed with the main focus on sensible heat transfer and with little emphasis on latent heat transfer

5、 or moisture transfer (Cames 1984; Sauer and Howell 198 1). Typically, fixed plate, sensible heat transfer wheel, heat pipe, and coil run-around loop heat exchangers have been used (Besant and Simonson 2000). These heat exchangers cannot, however, be used to directly control the indoor humidity. Thi

6、s is usually performed using auxiliary cooling and heating equipment, which is costly to install and operate. The high cost of these systems for buildings, especially in hot, humid climates, has led to interest in energy wheels (enthalpy wheels or desiccant- coated rotary energy exchangers). Althoug

7、h energy wheels can transfer both heat and moisture, their use does not elimi- nate the need for auxiliary heating and cooling equipment during extreme weather conditions. Thus, HVAC design engi- neers are often faced with deciding on the type and size of energy or heat recovery device to use for ea

8、ch given applica- tion to ensure appropriate indoor air quality and humidity and temperature control at minimal cost and maximum benefit to the system user. The sizing of heat exchangers and energy wheels for ventilation heat and moisture exchange in HVAC applications has not been done accurately an

9、d cost-effectively because it is a complex design problem requiring many pieces of informa- tion. Information required includes, not only mass flow rate of ventilation air and capital cost of the energy recovery device, but also the cost of auxiliary heating and cooling equipment and their operating

10、 costs, operating setpoints for temperature and humidity, and the ambient air temperature and humidity properties for an entire typical weather year. The calculations needed to evaluate the operating energy cost involve functions ofthese parameters integrated over time and are quite complex (ASHRAE

11、2000). Besant and Simonson (2000) present a discussion of the various configurations of heat and energy recovery devices and guidelines on how the operating energy cost may be evaluated. It is the purpose of this paper to show how this air-to-air heat and energy system design problem can be formulat

12、ed for a simple HVAC configuration and solved for the least life-cycle cost system while still retaining a small payback period. The design process is illustrated for the city of Chicago, where both large heating and cooling loads occur in HVAC applications. Although detailed mathematical expression

13、s are presented to calculate the operating energy demands, design tables are presented to facilitate the design process. Finally, an example design problem is presented, showing that payback periods of less than one year are often _ Yaw Asiedu is an operational research analyst at the Department of

14、National Defence, Ottawa, Ont. Robert W. Besant is professor emeritus and Carey J. Simonson is an associate professor in the Department of Mechanical Engineering, University of Saskatchewan, Saskatoon, S.K., Canada. 02004 ASHRAE. 381 achieved for energy wheels and sometimes for heat exchang- ers. Th

15、e life-cycle cost savings for auxiliary heating and cool- ing ventilation air far exceeds the capital cost of the heat exchanger or energy wheel, even when only a ten-year life cycle is considered. .- CO Il“ THE WHEEL SELECTION PROBLEM Figure 1 shows a schematic of the airflow paths in an HVAC syste

16、m, which includes a heat exchanger or energy wheel transfeming heat andor water vapor between the supply and exhaust airstreams. It is assumed that the bypass will be used when the complete heat exchanger or energy wheel trans- fer capacity is not needed to meet the design conditions (Tdes, Hdes) fo

17、r the conditioned space (Simonson et ai. 2000a, 2000b). Furthermore, when the heat exchanger or energy wheel cannot meet these required design conditions, recircu- lated air and auxiliary heating and cooling may be used sepa- rately or in combination. More complex HVAC design configurations than tha

18、t shown in Figure 1 are left to other studies. In selecting an energy wheel for an application with air flows, as shown in Figure 1, it is not economical to size the wheel based on only the most severe weather conditions. A better approach is to base the choice on the HVAC application needs and the

19、yearly weather data for the location. The hourly weather data, e.g., temperatures and enthalpies, in a location may be plotted as monotonically increasing functions with the time (number of hours in the year) as the abscissa and enthalpy or temperature as the ordinate. Figure 2 shows an example plot

20、. Time zero corresponds to the lowest ambient hourly temperature or enthalpy and hour 8,760 corresponds to the highest ones. The designer can use the typical hourly weather data for the city of interest; however, to make this chart easier to use in the design of systems, an analytical expression may

21、 be fitted to the temperature and enthalpy distributions. There are two functions that fit the weather data quite well for many North American cities. These are the hyperbolic sine function, T or h = y + a sinh(bt+z) , (1) Recimulaled Air I Conditioncd Auxiliary Space r Figure I Schematic of a HVAC

22、system with air-to-air heat/ energy recovery Heating nnd Cuoling and the inverse Gompertz given by where T and h are the ambient air temperature and enthalpy, respectively, y, z, a, b, k, I, m, and n are constants to be deter- mined through regression analysis, and t is the time in hours of the year

23、. While Equation 2 often fits the yearly weather data better, it is very difficult to integrate without resorting to numerical methods. This makes Equation 1 more suitable for evaluating the various parameters introduced later, which are needed to arrive at a good design. Therefore, for the remainde

24、r of this paper, Equation 1 is used. Arranging the ambient temperature and enthalpy as shown in Figures 2 and 3 is very useful because the area between these lines and supply conditions to the space (Tdes and hdes) represent the annual heating and cooling energy required to condition the outdoor ven

25、tilation air to the supply conditions when no air-to-air energy exchanger is used in the HVAC system. The area in Figure 3a has units of “C.h and is equal to the energy required to heat or cool 1 kg/s ofventilation air from the ambient temperature to the design temperature for supply air to the spac

26、e (kWhkg/s). The area in Figure 3b has units of kJ.h/kg (equates to the kWh/kg/s) and equals the energy required to heat and humid or cool and dehumidi 1 kgs of ventilation air from the ambient enthalpy to the design supply enthalpy. If the HVAC system employs an air-to-air energy exchanger, the aux

27、iliary energy requirements for heating and cooling can be reduced because the auxiliary equipment will only need to condition the ventilation air from Ts,o (/z,) to Tdes (hdes) (see Figure 1). The temperature and enthalpy at the outlet of the air-to-air energy exchanger depends on the effec- Hours o

28、f year “ /Ijr.mperature Figure 2 Distribution of ambient air temperature or enthalpy over the year for a typical city arranged for T(t) or h(t) increasing with time for all of the hours of operation of the HVAC system. 382 ASH RAE Transactions: Symposia Area representing the annual auxiliary sensibl

29、e energy needed to cool the ventilation air brea representing the annual auxiliary sensible energy needed to heat the ventilation air Area representing the annual auxiliary energy needed to cool and dehumidify the ventilation air O IO1 rea representing the annual auxiliary energy needed to heat and

30、humidify the ventilation air 101 Figure 3 Distribution of (a) ambient air temperature and (b) enthalpy showing the areas representing the annual auxiliary energy needed to heat, cool, humidifi, and dehumidifi the ventilation air without an air-to-air heat or energy exchanger: Area representing the a

31、nnual sensible cooling energy recovered by the air-to-air heat exchanger Area representing the total annual cooling energy recovered by the air-to-air energy exchanger Area representing the annual auxiliary sensible energy needed Area represcnting the annual auxiliary energy needed to cool represent

32、ing the annual auxiliary sensible y needed to heat the ventilation air rea representing the annual sensible heating rea representing the total annual heating energy covered by the air-to-air energy exchanger energy recovered by the air-to-atr heat exchanger Figure 4 Areas representing the (a) sensib

33、le and (b) total energy recovered by an air-to-air energy exchanger and the annual auxiliary energy needed to condition the ventilation air to the supply conditions to the space. tiveness of the exchanger, the indoor conditions in the build- In the following sections of this paper, analytical expres

34、- ing, and the outdoor temperature, as will be described sions will be developed to mathematically quantify the areas mathematically later. Hypothetical lines for Ts,o and hs,o are + Ts,;( 1 - = Tc = I-ot,c = J (t )dt - hdes(ttot - des) . (37) *I = Ides In this case, the areas (A:ot,h, , and A;ot,c,

35、 ) are in units of energy per unit mass flow rate (kWh/kg/s or kJ/kgh) and do not need to be converted using Equation 1 1. The associated cost over the life of the system is Auxiliary Winter Energy Requirements When the indoor humidity is controlled in a building all year, the amount of auxiliary en

36、ergy required in the winter is given by * * qaux,h = msAh with (39) and if h, 2 hmi, t, = O otherwise The cost of providing this energy is * * (41) - aux,h - wefhqaux,h . For buildings that do not humidifi the supply air during cold weather, this total energy would be less. Auxiliary Summer Energy R

37、equirements Including the energy required to subcool the supply air and condense water vapor and reheat it to Tde, the amount of auxiliary total energy required in the summer is given by qaux, - qaux, + qaux,c, + 2qaux,cc = m s A cA +m s A cs +2mSAcc * f t * * t * (42) with and (45) The total life-c

38、ycle energy cost for providing auxiliary energy is given by Iux,c = pwefEec(q:ux,cA + qzux,c, + qaux,cc) + hq:ux,cc (46) Recovered Total Energy summer periods are given by The amounts of energy recovered during the winter and * * * * (47) - qrec,h - sArec,h = ms(Atot,h-Ah) 9 The total cost savings i

39、s c:, = Pwef(ehqrec,h + e,q,.ec,c) * * Life-Cycle Cost and Payback Period Considerations The total life-cycle cost for a retrofitted system is : * Ctot,I = ex + Caux 9 (49) * * * where aux = aux,h+ caux,c with a payback period of and, for a new system, the life-cycle cost is with the payback period

40、given by ENTHALPY DISTRIBUTION MODEL AND DESIGN AIDS FOR CHICAGO Figure 10 shows the fitted enthalpy distribution for Chicago and the supply outlet air enthalpy distributions for ASH RAE Transactions: Symposia 391 Figure 1 O Supply outlet enthalpy distribution for energy wheels of diflerent eflectiv

41、eness. hin or hei , hdes Or hso 30 34 36 39 42 45 48 22 33480 43625 52607 6053 1 67493 73580 78868 23 28751 38896 47878 55802 62764 68850 74139 24 24181 34326 43308 51232 58194 6428 1 69569 different effectiveness values for a design enthalpy of 28 J/kg and indoor air enthalpy of 48 J/kg. The fitted

42、 equation is 25 26 27 28 29 h=F(t*)=21.68+ 15.59sinh(0.000402t*-1.569). (53) _ 19769 29914 38896 46820 53782 59869 65 157 15514 25658 34641 42564 49526 55613 6090 1 11412 21557 30539 38462 45424 51511 56799 746 1 17606 26588 345 12 41473 47560 52848 3658 13803 22785 30709 37671 43758 49046 The desig

43、n tables are shown in Tables 5 through 9. Equa- tion 54 is the mathematical expression fitted to the humidity ratio data to enable the development of Table 9. W= U(t*) = 0.003488 + 0.002197 sinh (0.00042 t*- 1.123) (54) EXAMPLE STUDY OF HEAT RECOVERY SYSTEM DESIGN FOR CHICAGO In this section, the se

44、lection of heat recovery systems using Tables 5 through 9 is illustrated. The selections are for all year-round applications in Chicago. The indoor tempera- ture is assumed to be 22C (72F) and 30% RH during the winter and 24C (75F) and 50% RH during the summer. The temperature of the supply air to t

45、he building is assumed to be constant 15C (59F) and 50% RH year-round, and 100% outdoor supply air is assumed. Other parameters are Mass flow rate (riz, ) Specific heat capacity of air (Cp) 1.2 kg/s 1 kJkgK o. 15 $kwh 0.05 $kwh Cooling energy cost (e,) Heating energy cost (eh) Present worth escalati

46、on factor (Pwej) 8.61 (for system life of 10 years) Design for Sensible Energy Recovery For the design temperature (Tdes) of 15“C, Table 4 gives values of 60,050 “C.h (or kWhkg/s) and 28,764 “C.h (or kWh/kg/s) for and Ator, respectively. This is an indi- cation of the level of heating and cooling re

47、quired and may be reduced by the use of a heat exchanger. The design problem is to select the best wheel with respect to payback period andor LCC from the set of wheels shown in Table 1 O. Table 1 1 gives the areas of the various energy recovery, heating, and cooling regions. These values were selec

48、ted from Tables 1 through 3 (Ah from Table 3, ACA from Table 1, and ACB from Table 2). Arec,h and A, represent the heating and cooling energy recovered by the heat exchanger, respectively, and are calcu- lated by Arec,h Atot,h -Ah (55) and I I l I I I I I 392 ASHRAE Transactions: Symposia Table 6. T

49、he Annual Energy Required to Cool the Ventilation Air Leaving the Energy Wheel to the Indoor Conditions Per Unit Mass Flow Rate (A: in kWh/kgls) for Different Indoor Enthalpies and Energy Wheel Effectiveness in Chicago. Et 0.55 0.60 hi, or hei 30 34 36 39 42 45 48 29547 23579 20940 17374 14242 11503 9123 26264 20959 18613 15444 12659 10224 8109 0.65 0.70 0.75 2298 1 18339 16287 13513 11077 8946 7096 19698 15719 13960 11583 9494 7668 6082 16415 13099 11633 9652 7912 6390 5068 1 0.90 II 6566 I 5240