ASHRAE OR-05-11-3-2005 Evaluation of a Ventilation Heat Pump for Small Commercial Buildings《小型商业大厦通风热泵的评价》.pdf

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1、OR-05- 1 1-3 Evaluation of a Ventilation Heat Pump for Small Commercial Buildings Kevin B. Mercer Associate Member ASHRAE ABSTRACT Air-to-air energy recovery can be used to reduce ventila- tion loads for small commercial buildings. Enthalpy exchang- ers are one approach to energy recovery that is co

2、mmonly employed. An alternative technology involves the use of a heat pump that operates between the ventilation and exhaust streams. In this papec the performance of a ventilation heat pump heat recovery unit was studied through simulation. A model of the heat pump was developed and validated using

3、 manufacturers and laboratory test data. This component model was then incorporated within a system simulation to investigate operating cost savings for the heat pump relative to a buse case that does not employ energy recovery but utilizes jxed minimum ventilation rates with an enthalpy economizer.

4、 The operating costs were found to begreater than the base case in moderate climates with relatively small savings in more extreme climates (hot or cold). The performance is not compet- itive with the performance of enthalpy exchangers or demand- controlled ventilation. The breakeven ambient conditi

5、onsfor positive savings utilizing a ventilation heat pump were also investigated. The primary performance penalties associated with the use of a heat pump are added fan and compressor power consumption, along with lost economizer savings. INTRODUCTION For many small commercial buildings, the ventila

6、tion airstream can contribute up to 40% of the building thermal load (ASHRAE 1993). The ventilation airstream is necessary to satisfy the indoor air quality (IAQ) requirements associated with ASHRAE Standard 62-2001. Several new technologies have been introduced to the market for preconditioning the

7、 fresh airstream and reducing primary equipment power James E. Braun, PhD, PE Member ASHRAE consumption. Heat pipes, run-around loops, enthalpy exchangers, and, most recently, heat pump heat recovery have all been introduced as alternative technologies. The investiga- tion described in this paper ev

8、aluates the performance of a heat pump heat recovery (HPHR) system for ventilation air pretreatment. During the cooling season, the ventilation heat pump operates to cool and possibly dehumidify outside air entering the primary rooftop unit (RTU). Heat is rejected from the heat pump into the exhaust

9、 airstream from the building. This exhaust air in turn cools the condenser coil at a lower temper- ature than the ambient. During the heating season, the heat pump operates in reverse to extract heat from the exhaust air and preheat the outside air. Figure 1 illustrates the location of the ventilati

10、on, exhaust, mixed, and return air conditions for a ventilation heat pump. The mixed air is further conditioned with a primary rooftop unit. Generally, the exhaust air is cooler than the ventilation air during the cooling season and warmer than the ventilation air during the heating season. Therefor

11、e, the heat pump operates very efficiently in both cooling and heating modes, and its COP should actually increase as the conditions become more extreme (i.e., more hot and humid for cooling and colder for heating). The heat pump requires an added exhaust an for circu- lating exhaust air. A ventilat

12、ion fan may also be required if the added static pressure of an extra heat exchanger is too high for the RTU supply fan to overcome. The heat pump is designed to be a retrofit to an existing RTU and, as such, all results presented assume a retrofit application. For new designs, the primary RTU could

13、 be downsized because of the addition of the heat pump. Kevin B. Mercer is a product development engineer at Modine Manufacturing Company, Racine, Wisconsin. James E. Braun is a professor of mechanical engineering at Ray W. Hemck Laboratories, Purdue University, West Lafayette, Indiana. 890 02005 AS

14、HRAE. Cooline Mode Ventilation - Exhaust Heatine Mode - Mixed i Ventilation Qm$ Figure I Heut pump conjgurution und operution with u rooftop unit. Heat pump heat recovery is a relatively new technology for application to ventilation air pretreatment and only a few studies have been published that re

15、late to this technology. Riffat and Gillott (2002) tested a revolving heat pipe ventila- tion heat pump system. The authors found an annual energy savings to exist for certain applications. However, the authors noted lower motor and compressor power would still be neces- sary to justi commercial imp

16、lementation of such a system. Another study conducted by Fehrm et al. (2002) addressed exhaust air heat pumps for markets in Sweden and northern Germany. The authors found an exhaust air heat pump reduced energy consumption and peak demand by as much as 20% when compared to a conventional gas-fired

17、boiler. The goal of the study described in this paper was to provide an assessment of heat pump heat recovery technology. A heat pump model was developed based on manufacturers and test data from a sample unit. The model was validated and then implemented into a ventilation strategy assessment tool

18、(VSAT) for analyzing performance (Braun and Mercer 2003). Annual cost savings were evaluated for a range of buildings and climates and compared with savings associated with alter- native technologies for reducing ventilation loads. The breakeven ambient conditions for positive savings with appli- ca

19、tion of a ventilation heat pump were also investigated. LAB TEST SETUP The ventilation heat pump was installed in the outdoor room of an arrangement of psychrometric chambers. The unit was set up for operation between controlled ambient and exhaust airstreams, as shown in Figure 2. In order to provi

20、de the correct air flow velocity for reading pressure drop and relative humidity, a 12x10 inch duct was selected to connect the outdoor and indoor rooms for both the supply and exhaust airstreams. The heat pump was fitted with two switches for managing operation. One switch controlled fan power and

21、the other switch was a toggle between cooling and heating mode. In order to change the fan speeds, a wire setting located on the supply and exhaust air fan was adjusted. The outside air damper was maintained at the factory default Figure 2 Intilution heat pump. setting of 45“ throughout testing. A d

22、rain pan was also fabri- cated to collect any condensate. In order to measure tempera- ture, four thermocouple grids were placed at the ends of each duct section and at the air intake entrance to the heat pump. The temperature data points were then averaged to get an over- all reading. The rated acc

23、uracy of the thermocouple measure- ments was f O. I “C. Two pitot tube arrays, consisting of five pitot tubes each, were constructed to evenly traverse the open cross-sectional area of the supply and exhaust air ducts. These arrays provided a reasonably accurate pressure difference for calculating a

24、ir flow velocity. A voltage output of O to 5 volts DC equated to a velocity pressure range of O to O. 1 in. H,O. Humidity in each airstream was measured using a chilled- mirror dew-point meter and then checked against a reading from a portable thermo-hygrometer. The hygrometer had an accuracy of *3%

25、 relative humidity according to the manufac- turer. Power consumption of the heat pump was monitored using a power meter. The power for each fan was recorded for two different fan settings without the compressor engaged. Throughout the laboratory testing, static pressures in the ducts and the damper

26、 position remained unchanged and, therefore, the fan power was relatively constant. When the compressor was operating, total power of the unit (compressor and fans) was recorded. Compressor power was determined as the difference between the total and fan powers. A wide range of ambient conditions we

27、re simulated; however, return air conditions were kept fairly constant at 75F (23.89“C) for cooling and 70F (2 1.1 1 OC) for heating throughout testing. Characterization tests of the heat pump were performed for a high and low ventilation air flow rate (high and low fan speed settings). The high and

28、 low air Bow rates were approximately 1050 cfin and 775 cfm, respectively. A wide range of ambient conditions for heating and cooling modes were considered with steady-state data recorded at each point. Afer a steady-state condition was achieved at each test ASH RAE Transactions: Symposia 891 Table

29、1. Ventilation Heat Pump Rated Conditions Ambient Air Wet- Bulb Temp. OF (OC) 75 (23.89) - I I I Capacity COP SHR KBtdh (kW) 22.5 (6.59) 4.515 0.9 22.2 (6.50) 7.425 - Airflow Rate Return Air Temp. Ambient Air Dry- Bulb Temp. cfmlton OF (OC) - Cooling mode 533 75 (23.89) Heating mode 540 70 (21.11) O

30、F (“C) 95 (35) 33 (0.56) point, data were collected every 5 seconds for a total of 2 % minutes. This resulted in approximately 30 data readings at a single test point. These data readings were averaged to obtain a single steady-state data point at each operating condition. MODEL DEVELOPMENT AND VALI

31、DATION Capacity and COP The heat pump was modeled in both cooling and heating modes using a semi-empirical approach for direct-expansion air conditioners found in the ASHRAE HVAC Toolkit (Bran- demuehl et al. 2000). The model corrects the capacity and COP given at a rating point using empirical corr

32、elations. A simple analytical model is employed for correcting the rated sensible heat ratio (SHR) for cooling. Outlet air conditions are calculated from mass and energy balances for both cooling and heating modes. For the heat pump considered in this study, Table 1 gives capacities, COPS, and SHR a

33、t rating conditions for heating and cooling modes. The COP is the ratio of the cooling or heating capacity to the compressor power input. The capacity does not include the effect of the primary supply air fan. The sensible heat ratio is the ratio of sensible to total cooling capacity and is only def

34、ined for cooling mode. The total capacity and COP for cooling or heating are determined from the rated values and correction factors for air flow rate and inlet conditions using Equations 1 and 2. Op coprat fCOP,/( Tamb,Tre/) .fcoP,m(m/mrg/) (2) The four correction factors are computed using the fol

35、lowing empirical equations: fcop,m(m/firat) = a4 + b, (filmrat) (6) Separate correlations (i.e., a, b, c, d, and f values) are utilized for heating and cooling mode. The air mass flow rate used within the correlations is the ventilation flow rate, which is assumed to be equal to the exhaust air flow

36、 rate. In cooling mode, the ambient temperature ( Tamb) used in the above rela- tions is the inlet ventilation stream wet-bulb temperature, whereas in heating mode it is the inlet ventilation stream dry- bulb temperature. In both modes, the return temperature (T,J is a dry-bulb temperature. The empi

37、rical coefficients (a, b, c, etc.) of the correction factor equations are calculated using a linear regression applied to available data for the heat pump. Values of the coefficients are given for cooling and heating mode in Tables 1 and 2 of Appendix A. In cooling mode, the correction factor equati

38、ons (Equa- tions 3-6) for determining total capacity and COP utilize inlet air wet-bulb temperature as a variable. This is appropriate for a wet coil, but for a dry coil performance depends on dry-bulb temperature only. For a given air dry-bulb temperature, coil performance should be approximately c

39、onstant for all humid- ity ratios below the value where condensation would begin. This fact is used in order to apply a single set of correction factor equations for both wet and dry conditions. In cooling mode, the coefficients in the correlations are determined using data where condensation occurr

40、ed. During application of correlations in a model, the coil is assumed to be wet if the calculated sensible heat ratio (see next subsection) is less than 1. If the sensible heat ratio is found to be greater than one, then the inlet wet-bulb temperature is adjusted iteratively for the given dry-bulb

41、until sensible heat ratio is exactly equal to 1. At the point where condensation begins, a wet and dry analysis both give the same result. Therefore, the wet analysis deter- mined with this artificial wet-bulb temperature and SHR equal to 1 give a correct result for a dry coil operating with the act

42、ual dry -bulb temperature. Sensible Heat Ratio The model for estimating sensible heat ratio utilizes the concept of a bypass factor. The rated bypass factor is deter- mined from cooling inlet and outlet rated air conditions, the rated air flow rate per unit cooling capacity, and rated sensible heat

43、ratio. An evaporator conductance parameter based on the rated bypass factor and air mass flow rate is computed for use 892 ASHRAE Transactions: Symposia in the steady-state sensible heat ratio calculations at any oper- ating condition. The bypass factor model for an evaporator coil considers two dif

44、ferent airstreams: one airstream is in close contact with the coil and leaves as saturated air at the effective temperature of the coil surface and the other airstream bypasses the coil completely and remains at the entering air condition. The effective coil surface temperature is often termed the a

45、pparatus dew point. Mass and energy balances on the airstreams give the following relationships: . %vap.out am + mbw (7) - aadp +lout = m (8) - moutu)evap,out mappWadp + mb,vpwevap,in mouthevap,out =m app h adp mbyphevap,in (9) Tdp Tevap,oui Tevap.,” The bypass factor is defined as the ratio ofbypas

46、s fow to total tions according to air flow. The bypass factor is related to coil operating condi- 3 Psychrometric depiction Of bYPassfactor The outlet air state lies on a straight line that connects the inlet air state with the apparatus dew point. This is depicted in Figure 3. As the bypass factor

47、increases, the outlet air state moves closer to the inlet state. It can be shown that the bypass factor is related to the coil sensible heat transfer and described in terms of the number of coil transfer units, or NTU. For a given bypass factor and inlet and outlet air enthalpy, the saturated air en

48、thalpy at the apparatus dew point is calculated from hevapjn- hevap,out hevap,in - 1 -BF The apparatus dew-point humidity ratio is then determined from psychrometric relations using the saturated air enthalpy with a relative humidity of 100%. Outlet air humidity ratio is then determined by subsectio

49、n is used to determine a fictitious wet-bulb temper- ature where SHR is unity for the actual dry-bulb temperature. The fictitious wet-bulb temperature is then used as the inlet condition for capacity and COP calculations. Val id at ion Both manufacturers and laboratory data for cooling and heating mode were used to determine coefficients of the correction factor equations and to evaluate their accuracy in estimating capacity and COP. The conductance of the evapo- rator coil was also optimized for use in calculating sensible heat ratio. Lab test data extended beyond the limit