A simplified method for defining air humidification.pdf

ORIGINAL A simplifi ed method for defi ning air humidifi cation and dehumidifi cation requirements Azra KorjenicThomas Bednar Received: 5 August 2011/Accepted: 23 May 2012/Published online: 14 June 2012 ? Springer-Verlag 2012 AbstractNowadays, the use of ventilation systems is very common. In museums and exhibition rooms, they are used to guarantee the optimal temperature and humidity conditions for valuable and irreplaceable objects. Ventila- tion systems are increasingly used even in residential and offi ce buildings. In Austria and other similar climates, winter outdoor air is very dry. Because of the constant air exchange using a ventilation system, the relative humidity of indoor air is often lower than the comfort limit, so that air humidifi cation is necessary. In the same rooms, air must be dehumidifi ed during the summer months. Humidifi ca- tion and dehumidifi cation should be minimized as they are processes that consume a large amount of energy. The problem is that until now, it has been unclear how much humidifi cation or dehumidifi cation is necessary and could not be calculated easily. In this investigation, a simplifi ed method for determining the humidifi cation and dehumidi- fi cation demands taking into account the effective moisture capacity of the room was developed. The developed model is used for the calculation of the energy balance and will be integrated into the program for energy certifi cation in Austria. 1 Introduction In the last decades, energy effi ciencies in buildings have developedgreatly.Thebuildingenvelopehasbeen developed to the extent that the effect of transmission heat losses, leakages and thermal bridges have been minimized. Because of these developments, ventilation is currently playing a central role. Ventilation systems are being used more frequently to reduce heat losses by natural ventilation in winter, or to supply precooled fresh air to the building during the warm months. Exterior air may be either humidifi ed or dehumidifi ed as not only the temperature, but the humidity is important in order to achieve the prede- termined comfort conditions. The use of renewable energy for humidifi ers has also been developed. In 1, a humidi- fi er unit required for the air conditioning system is shown which uses solar energy instead of electricity with a pan humidifi er. A mathematical simulation of the solar humidifi er was also presented based upon weather data of a hot day in the region. The heat or moisture transfer amount from the exhaust air to the supply air depends on the recovery coeffi cient of the heat and moisture exchanger. Common dimensioning of heat and moisture exchangers is based upon the ability of the exchanger to transfer suffi cient heat and moisture from the exhaust to the supply air for defi ned indoor and outdoor temperatures and humidities. It is already known that humidifi cation and dehumidi- fi cation levels are much more diffi cult to adjust than heat recovery and requires more energy. Humidifi cation is a signifi cant energy consumer. Humidifying to 50 % instead of no humidifi cation may increase the heat demand by more than 30 % 2. Signifi cant amounts of energy can be saved with a suitable reduction of the ventilation rate and appropriate set points of indoor conditions according to current standards and guidelines 3. The formulas are given in 3 for the energy required for heating, cooling, humidifi cation and dehumidifi cation of outdoor air to the desired indoor conditions. A. Korjenic (su, and the humidity recovery values, gx;su, are always equal, and the selected transmission rate for both heating and cooling for the whole year is constant. It should be noted that for the previously stated constant annual heat and humidity recovery values, the heat recovery values and the humidity Fig. 1 Scheme of a full air-conditioning system 1792Heat Mass Transfer (2012) 48:17911802 123 values are always assumed to be equal; however, these values are equal only for certain conditions. For the cases with heat recovery (HR), air fl ows through the exhaust air or outdoor air and thus, transfers heat for heating scenarios or transfers cold for cooling scenarios. If the outside air is supplied directly to the building bypassing the heat recovery system, the bypass is dependent upon whether the supply air temperature after HR has a similar temperature to the established indoor air temperature. If, using HR, no reduction of the temperature difference between outdoor and indoor air set point is reached, no heat is transferred and the thermodynamic changes in outside air will only be made through the heating and/or cooling register and the humidifi er. The use of a heat recovery system with a hygroscopic surface is not a deciding factor for the operation of the HR if, in addition to temperature, moisture transfer is energetically advantageous. Tempera- ture is the only reference for heat recovery operation. The effi ciency of the various components of the HVAC system and the resulting change in end energy demand were not considered here. No differentiation was made whether the outside air cooling was through a direct system (direct expansion) or through an indirect system (cold water set). The energy demand of various support aggre- gates such as pumps, actuators, and the rotary heat exchanger drive were also not considered. To which temperature or humidity the supply air is conditioned depends on the difference between the exhaust and the desired room air. The lower the exhaust air tem- perature, the higher the air temperature. The same applies to the supply air humidity level. Because of the comfort criteria, the supply air temperature is limited to 50 ?C. The maximum acceptable supply air humidity is determined by the dew point temperature of the supply air. The maximum supply air humidity is 16 g/kg for a supply air temperature of 50 ?C. The supply air state results from the sum of the target state, which is approximately the difference between the exhaust air conditions and the set point conditions. When the conditions of the exhaust air are equivalent to the set point conditions, the air supply and the set point parameters are the same. The calculation of supply air temperature, in contrast to humidity, is calculated as an additional increased air tem- perature correction generated by the supply fan. Basically: Winter mode: xExhaustxSet air humidifi cation is needed TExhaustTSetheating is needed Summer mode: xExhaustxSet air dehumidifi cation is needed TExhaustTSetcooling is needed xair humidity (g/kg) Tair temperature (K) Grounded on the calculated values of the climate data- set, relative humidity, exterior air temperature, and the total air pressure of the related location, the following constants and equations were used for the calculations: Constants according 8 or 9: Molar mass of airML= 28.96 g/mol Molar mass of water vaporMD= 18.02 g/mol Specifi c gas constant of airRL= 287.1 J/kg K Specifi c gas constant of water vapor RD= 461.40 J/kg K Specifi c heat capacity of aircpL= 1.01 kJ/kg K Specifi c heat capacity of water vapor cpD= 1.86 kJ/kg K Evaporation enthalpy of water at t = 0 ?C ro= 2,501 kJ/kg Air density at standard conditionsqL;0= 1.292 kg/m3 Air pressure (standard conditions)po= 1.013 bar Temperature (standard conditions)To= 273.15 K Gravitational accelerationg = 9.81 m/s2 Air pressure at the calculated location p pL pD1 pLPartial pressure of dry air (Pa) pDWater vapor pressure (Pa) Density: qL pL RL? T ; qD pD RD? T 2 Moisture content mixing ratio: x 0:62198 pD pL 3 Relation x andpD;sat;w:pD;sat;w x 0:62198 x ? p4 Enthalpy: h cpL? # x ? r0 cpD? # ? 5 The Eqs. 6 and 7 for saturated vapor pressure are included in the recommendations by the WMO (World Meteorological Organization) 9. Saturated vapor pressure of humid air above water (-45 to 60 ?C), pD;sat;w fp ? 6:112 ? exp 17:62 ? # 243:12 # ? in hPa6 Saturated vapor pressure of humid air above ice (-65 to 0 ?C), pD;sat;i fp ? 6:112 ? exp 22:46 ? # 272:62 # ? in hPa7 fp 1:0016 3:15 ? 10?6? p ? 0:074 ? p?18 Heat Mass Transfer (2012) 48:179118021793 123 # in ?C p in hPa. Dew point and freezing point: #dew 243:12 ? ln pD;sat;w?6:112 ? fp ? 17:62 ? ln pD;sat;w?6:112 ? fp ? ? Water ?45 ?C bis 60?C 9 #frost 272:62 ? ln pD;sat;i?6:112 ? fp ? 22:46 ? ln pD;sat;i?6:112 ? fp ? ? Ice ?65 ?C bis 0?C 10 CIMO Guide, 7th Guide to Meteorological Instruments and Methods of Observation Relative humidity u pD pD;sat;w 11 Theindividualconnectionsofaheatexchanger, analogousthenomenclatureEN13141-710and O NORM EN 308 11 is presented in Fig. 2. Heat recovery index (ex = relation on exhaust air side, su = relation on supply air side) gt;ex T12? T11 T11? T21 ? ja11 ja22 gt;su T22? T21 T11? T21 ? ja22 ja11 12 T11 Exhaust air temperature infl ow (before recovery) in K T21 Supply air temperature infl ow (before recovery) in K T22 Supply air temperature outfl ow (after recovery) in K ja11 Air mass fl ow exhaust air in kg/s ja22 Air mass fl ow supply air in kg/s Moisture recovery index (ex = relation on exhaust air side, su = relation on supply air side) gx;ex x12? x11 x11? x21 ? ja11 ja22 gx;su x22? x21 x11? x21 ? ja22 ja11 13 x11 Exhaust air moisture content infl ow (before recovery) g/kg x21 Supply air moisture content infl ow (before recovery) g/kg x22 Supply air moisture content outfl ow (after recovery) g/kg Air temperature increase by a fan: DT DpVent gges? cpL? ja 14 DTTemperature increase in Kelvin DpVentFan pressure increase for the considered exhaust air duct length in Pa gges Total fan effi ciency factor ja Air mass fl ow in kg/s. 2.2 Cooling case calculation If the supply air temperature is reduced by the temperature increase caused by the supply air fan, and is found to be lower than the outside air temperature, cooling is required. Figure 3 shows the process fl ow of a system with heat and moisture recovery with gt;su gx;su 45%. If the fresh air humidity after passing through moisture recovery is higher than the desired supply air humidity, the air must be dehumidifi ed by cooling the fresh air to the dew point of the requested supply air humidity. As for applications in air conditioning, the dew point temperature is almost always lower than the allowable supply air temperature; the supply air must be heated after dehumidifi cation to the dew point temperature to the supply air temperature. Temperature losses between the primary and secondary sides in the heat exchanger, and cooling the supply air under the dew point temperature as exists in cooler operation were not taken into account by calculation. System variant: gt;su gx;su 45%; xin 9g=kg Outsideaircondition:0outside= 28.1 ?C,xout= 8.95 g/kg Air condition H ? MR: 0HR= 27.6 ?C, xMR= 10.32 g/kg Supply air condition: 0in= 20 ?C, xin= 9 g/kg The calculation of the instantaneous cooling energy demand is the product of the outside air density, volume fl ow, and the enthalpy difference between the air condition at the heat recovery output and the dew point, or in the case of exclusively sensible cooling, of the state of the supply air. 2.3 Heating case calculation If the supply air temperature is reduced by the temperature increase by the supply fan, and is higher than the exteriorFig. 2 Designation the individual connections of a heat exchanger 1794Heat Mass Transfer (2012) 48:17911802 123 air temperature, heating is required. If the exterior air humidity, or in systems with moisture recovery, the humidity of the supply air after humidity recovery, is less than the desired indoor humidity, outside air humidifi cation is neces- sary. A humidifi er humidifi es the air in an isothermal process. If the outside air humidity, or the air humidity after moisture recovery, is equal to the required supply air humidity, the exterior air will be heated to the supply air temperature. Air conditioning for the required supply air humidity of 6 g water vapor per kg dry air with heat and moisture recovery, gt;su gx;su 45%, is presented in Fig. 4. System variant: gt;su gx;su 45%, xin= 9 g/kg Outsideair condition:0outside= 5.1 ?C, xout= 5.19 g/kg AirconditionexitH ? MR:0HR= 12.71 ?C, xMR= 8.03 g/kg Supply air condition: 0in= 20 ?C, xin= 9 g/kg The calculation of the heating energy demand will be determined similarly to the cooling energy demand. The cooling energy demand is the product of the exterior air density, volume fl ow and the enthalpy difference between the air leaving the heat recovery unit, and in the case of dehumidifi cation, the difference between the dew point and the supply air states. 3 Simplifi ed approach for defi ning the air humidifi cation and dehumidifi cation demands The humidifi cation and dehumidifi cation demands are based upon the amount of moisture that must be supplied or removed so that the desired air humidity is reached for a defi ned area. The amount of air humidity depends on the room use (residence, offi ce, museum, etc.). For the calculation, it is necessary to indicate the desired humidity level, the observable tolerance, the use of room (heat and moisture release), and the climate as a part of the room defi nition. The required range description can be specifi ed using a lower and an upper limit for the absolute indoor air humidity. For the calculation of the humidifi cation and dehumid- ifi cation demands, it is necessary to determine the moisture gain or loss in the ventilation system, taking moisture recovery into account. The hourly exterior temperature Fig. 3 Exemplary course, the change in state of the cooling case Fig. 4 Exemplary course, the change in state of the heating case Heat Mass Transfer (2012) 48:179118021795 123 data is used for the calculations. Hourly HVAC energy needs are calculated for the whole year. The time-depen- dent given fl ow rate is taken into account by the use pattern (the operating hours of the building). In the Fig. 5, the essential components are presented that for the determination of the humidifi cation and dehumidi- fi cation demand are necessary. In Fig. 6, the registers are added to that are relevant for the calculation of the energy demand for the shown confi guration. 3.1 Determination of the air humidifi cation and dehumidifi cation needs The starting point for the calculations is the balance equa- tions for the humidifi cation demand (HD), and dehumidifi - cation demand (DHD). For the humidifi cation demand, the ventilation losses (convection) are the reciprocal to ventila- tion gains and interior moisture production. The dehumidi- fi cation demand isthe inverse to the humidifi cation demand. HD MLoss;Conv? gm? MGain;Conv MGain;Use ? 15 DHD MGain;Conv MGain;I ? ? gm? MLoss;Conv16 HD monthly humidifi cation demand in kg DHD monthly dehumidifi cation demand in kg MLoss;Convmonthly moisture losses through ventilation (convection) in kg MGain;Convmonthlymoisturegainsbyventilation (convection) in kg MGain;Usemonthly moisture gains by use in kg The degree of HD utilization can be determined using the win-loss ratio (c) and the time constant (s). The c and s are obtained in accordance with O NORM EN ISO 13790, in the following manner: cm MGain MLoss a a0 sm s0 gm 1 ? ca m 1 ? ca1 m 17 adimensionless numerical parameter dependant on the time constant a0dimensionless reference numerical parameter smmoisture time constant s0reference time constant To determine the time constant, the specifi c losses (Lm) and the moisture capacity (Cm) of the observed zone must be calculated. sm Cm Lm ? Lm MLoss ci;soll? ce ? Cm X n A ? vm18 vmEffective moisture capacity in kg/m2/(kg/m3) Aarea in m2. 3.2 Effective thermal and hygric capacity of constructions If the temperature rises in a room, heat is absorbed by all building constructions. Analogously, the surfaces absorb moisture when the water vapor pressure in the room is higher than the surfaces. When the air temperature in the room and the surface temperatures are similar, this also applies to the absolute humidity. The effective heat capacity related to area is the coef- fi cient which links the amplitude of the temperature range with the amplitude of heat fl ow. The longer the period length is (instead of daily, weekly or monthly oscillation), the effect of larger surface areas increases the effective heat storage capacity and distribution. The effective area- related moisture capacity is the coeffi cient linking the moisture oscillation amplitude with the moisture fl ow amplitude. The heat capacity defi nition results from the basic Eq. 19, according to 13. Q v ? 2 ?T Ztp 0 maxqt;0 ? dt19 vEffective heat capacity J/m2K Fig. 5 Demonstration of calculation model exchanger ? room 1796Heat Mass Transfer (2012) 48:17911802 123 The moisture capacity can be described analogous to the heat capacity: M vm? 2 ? c Ztp 0 maxmt;0 ? dt20 vmEffective moisture capacity kg/m2/(kg/m3) c maximum moisture concentration kg/m3 The starting point for the calculation of a multilayer capacity is the relationship between the sinusoidal tem- perature amplitude and the heat fl ow amplitude on both sides of a layer. The Fourier equation for heat transfer was solved there- forewiththesocalledmatrixmethod14,whichisalsoused fortheheatstorageofconstructions.Thehygriccapacitycan be analog formulated. The Layer model with amplitude of thermal/hygric parameters are shown in Fig. 7. Tb q b ? Z11Z12 Z21Z22 ? ? Ta q a ? c b jb ? W11W12 W21W22 ? ? c a ja ? 21 To determine the elements of the matrix according to EN ISO 13786, the periodic penetration depth is needed 15: d ffi ffi ffi ffi ffi ffi ffi ffi ffi ffi ffi ffi ffi