Evaporative cooling is used in almost all climates. The wet-bulb temperature of the entering airstream limits direct evaporative cooling. The wet-bulb temperature of the secondary airstream limits indirect evaporative cooling.

Design wet-bulb temperatures are rarely higher than 78°F, making direct evaporative cooling economical for spot cooling in kitchens, laundries, agricultural, and industrial applications. In regions with lower wet-bulb temperatures, evaporative cooling can be effectively used for comfort cooling, although some climates may require mechanical refrigeration for part of the year.

Indirect applications lower the air wet-bulb temperature and can produce leaving dry-bulb temperatures that approach the wet-bulb temperature of the secondary airstream. Using building return air as the secondary airstream can further enhance performance of the indirect cooler, especially if the building has a high capacitance for moisture absorption. Incorporating sensible precooled air in the secondary airstream further enhances the indirect evaporative cooler’s cooling capability.

Direct evaporative cooling is an adiabatic exchange of energy. Heat must be added to evaporate water in the supply airstream. The air into which water is evaporated supplies the heat; thus, the dry-bulb temperature is lowered and the moisture content increases. The amount of heat removed from the air equals the amount of heat absorbed by the water evaporated as heat of vaporization. If the direct evaporative cooler sump water is recirculated in the direct evaporative cooling apparatus, the water temperature in the reservoir approaches the wet-bulb (wb) temperature of the air entering the process. By definition, no heat is added to, or extracted from, an adiabatic process; the initial and final conditions of the process air fall on a line of constant wet-bulb temperature, which nearly coincides with a line of constant enthalpy on the psychrometric chart (Figure 1).

The maximum reduction in dry-bulb temperature is the difference between the entering air dry- and wet-bulb temperatures. If air is cooled to the entering air wet-bulb temperature in a direct evaporative cooling process, it becomes saturated and the process would have 100% wet bulb depression effectiveness (WBDE). WBDE is the depression of the dry-bulb temperature in the process divided by the difference between the entering air dry- and wet-bulb conditions.

When a direct evaporative cooling unit alone cannot provide desired conditions, several alternatives can satisfy application requirements and still be energy-effective and economical to operate. The recirculating water supplying the direct evaporative cooling unit can be increased in volume and chilled by mechanical refrigeration to provide lower leaving wet- and dry-bulb temperatures and lower humidity. Compared to the cost of using mechanical refrigeration only, this arrangement reduces operating costs by as much as 25 to 40%. Indirect evaporative cooling applied as a first stage, upstream from a second, direct evaporative stage, reduces both the entering dry- and wet-bulb temperatures before the air enters the direct evaporative cooler. Indirect evaporative cooling may save as much as 60 to 75% or more of the total cost of operating mechanical refrigeration to produce the same cooling effect for 100% outdoor air (OA) systems. Systems may combine indirect evaporative cooling, direct evaporative cooling, heaters, and mechanical refrigeration, in any combination.

The psychrometric chart in Figure 1 shows what happens when air is passed through a direct evaporative cooler. In the example shown, assume an entering condition of 95°F db and 75°F wb. The initial difference is 95 − 75 = 20°F. If the effectiveness is 80%, the depression is 0.80 × 20 = 16°F db. The dry-bulb temperature leaving the direct evaporative cooler is 95 − 16 = 79°F. In the adiabatic evaporative cooler, only part of the water recirculated is assumed to evaporate and the water supply is recirculated. The recirculated water reaches an equilibrium temperature approximately the same as the wet-bulb temperature of the entering air.

Figure 1. Psychrometrics of Evaporative Cooling

The performance of an indirect evaporative cooler can also be shown on a psychrometric chart (Figure 1). Many manufacturers define effectiveness similarly for both direct and indirect evaporative cooling equipment. With indirect evaporative cooling, the cooling process in the primary airstream follows a line of constant moisture content (constant dew point). Indirect evaporative cooling effectiveness is the dry-bulb depression in the primary airstream divided by the difference between the entering dry-bulb temperature of the primary airstream and the entering wet-bulb temperature of the secondary air. Depending on heat exchanger design and relative quantities of primary and secondary air, effectiveness ratings may be as high as 85%.

Assuming 60% effectiveness, and assuming both primary and secondary air enter the apparatus at the outdoor condition of 95°F db and 75°F wb, the dry-bulb depression is 0.60(95 − 75) = 12°F. The dry-bulb temperature leaving the indirect evaporative cooling process is 95 − 12 = 83°F. Because the process cools without adding moisture, the wet-bulb temperature is also reduced. Plotting on the psychrometric chart shows that the final wet-bulb temperature is 71.5°F. Because both wet- and dry-bulb temperatures in the indirect evaporative cooling process are reduced, indirect evaporative cooling can substitute for part of the refrigeration load in 100% OA systems. This sensible cooling process can make a second-stage direct evaporative cooler more effective in arid climates, because the sensible cooling it contributes allows reduction of mechanical cooling.

The benefit of humidification by the adiabatic direct evaporative cooling component is often overlooked in evaporative cooling design. During cold winter conditions, the outdoor air required to meet building code ventilation requirements quickly drives indoor relative humidity below acceptable levels.

Figure 2 shows a schematic of an air-handling unit design that provides hydration to the dry outdoor air in winter without greatly affecting building heating energy costs. The variable-air-volume (VAV) design concept (Scofield et al. 2016) uses a heat pipe air-to-air heat exchanger, selected for indirect evaporative cooling (IEC) in summer, to recover heat from the building return air to (1) increase the fresh air fraction introduced to the building and (2) provide the heat necessary for evaporation of water for humidification furnished by the direct evaporative cooler/humidifier (DEC/H). In mild winter climates, such as Sacramento, California, this design can provide 100% outdoor air to a VAV cooling-only air-handling unit with free hydration of the building supply air sufficient to hold indoor relative humidity above 32% at 72°F indoor temperature when ambient conditions are 22°F in winter.

Combining an all-air VAV cooling design with an air-to-air heat exchanger provides significant heating and humidification energy savings, compared to a more conventional air-side economizer design without heat recovery. When ambient temperatures drop in winter and air mass flow decreases through the heat exchanger, the heat pipe becomes more effective. The percentage of heat recovery increases as air-side parasitic losses (caused by static pressure) decline. When dwell time in the heat exchanger increases, effectiveness increases. The same is true for the DEC/H heat exchanger. In the Sacramento application, the heat pipe dry-to-dry effectiveness increases from 65.7% at cooler ambient bin weather conditions, when the VAV system is at 70% of full flow, up to 68.3% effective, at the winter design ambient of 22°F and a flow of 50%. The IEC cooling sprays in Figure 2 would be off at all ambient bin conditions below 62°F db. Heat recovery is achieved using the 72°F room return air condition in winter.

Recent studies (Karim et al. 1985; Tang 2014; Taylor 2014) identify low room relative humidity as a factor contributing to the buoyancy, viability and spread of some airborne pathogens, such as the flu virus, within the human breathing zone. Maintaining indoor humidity between 40 and 60% at a comfortable room temperature can reduce the risk of human exposure to a number of respiratory infections (Sterling et al. 1985).

Figure 2. Schematic of VAV Heat Recovery Economizer Unit to Provide Free Adiabatic Humidification during Cold Ambient Conditions (Scofield et al. 2016)

Another factor in the spread of pathogens within the human breathing zone indoors is air turbulence. Artificially high room air change rates will push airborne contaminants farther away from a human host thereby exposing healthy humans farther away from a cough or sneeze (Pantelic and Tham 2013). Reducing a VAV supply air temperature set point in cold weather reduces not only air change rates, but also fan energy for both the supply and return fans in Figure 2 (English et al. 2015).

Dilution of indoor-generated contaminants and airborne pathogens with a filtered supply of outdoor air is just one advantage of an all-outdoor-air design. Human productivity has been shown to increase 3 to 4% when outdoor air is increased from the code minimum of 15 cfm per person up to 106 cfm per person (Seppanen et al. 2005). Absentee rates for business employees were shown to decrease by 33% when fresh air rates went from 24 to 48 cfm per person (Milton et al. 2000).

The installed cost of a DEC/H component in an air-handling unit, such as the one shown in Figure 2, adds about $1.50 to $1.75 per supply air cubic foot per minute to the first cost. This added cost has a very short payback because of the payroll savings from increased productivity and reduced absenteeism. If summer sensible cooling and winter hydration of outdoor air are not a design option, the outdoor air increase shown in Table 1 would still make the heat recovery economizer (HRE) a viable option, at a first cost of approximately $2.00 to $2.50 per supply air cubic foot per minute installed. In cold climates, the air-to-air heat exchanger offers additional protection against winter freeze-up of water coils downstream in the air-handling unit. Blending preheated OA with building return air becomes a much simpler process. Cold-climate blending devices may often be eliminated without concern for hot- and cold-air stratification causing water freezing problems. An HRE alone offers outdoor ventilation rates well above code minimums in even the coldest climates without preheating costs, unlike a conventional air-side economizer without heat recovery. Outdoor air dampers with OA preheating through heat recovery admit more outdoor air than they would without preheating. Thereby, the fresh air fraction to the building is increased above code minimums prescribed by ASHRAE Standard 62.1 (Scofield and Bergman 1997).

Recirculated Water. Except for the small amount of energy added by shaft work from the recirculating pump and the small amount of heat leakage through the unit enclosure, evaporative humidification is strictly adiabatic. As the recirculated liquid evaporates, its temperature approaches the thermodynamic wet-bulb temperature of the entering air.

The airstream cannot be brought to complete saturation, but its state point changes adiabatically along a line of constant wet-bulb temperature. Typical saturation or humidifying effectiveness of various air washer spray arrangements is between 50 an 98%. The degree of saturation depends on the extent of contact between air and water. Other conditions being equal, low-velocity airflow is conducive to higher humidifying effectiveness.

Preheated Air. Preheating air increases both the dry- and wet-bulb temperatures and lowers the relative humidity; it does not, however, alter the humidity ratio (i.e., mass ratio of water vapor to dry air) or dew-point temperature of the air. At a higher wet-bulb temperature but with the same humidity ratio, more water can be absorbed per unit mass of dry air in passing through the direct evaporative humidifier. Analysis of the process that occurs in the direct evaporative humidifier is the same as that for recirculated water. The desired conditions are achieved by heating to the desired wet-bulb temperature and evaporatively cooling at constant wet-bulb temperature to the desired dry-bulb temperature and relative humidity. Relative humidity of the leaving air may be controlled by (1) bypassing air around the direct evaporative humidifier or (2) reducing the number of operating spray nozzles or the area of media wetted.

Heated Recirculated Water. Heating humidifier water increases direct evaporative humidifier effectiveness. When heat is added to the recirculated water, mixing in the direct evaporative humidifier may still be modeled adiabatically. The state point of the mixture should move toward the specific enthalpy of the heated water. By raising the water temperature, the air temperature (both dry- and wet-bulb) may be raised above the dry-bulb temperature of entering air. The relative humidity of leaving air may be controlled by methods similar to those used with preheated air.

Direct evaporative coolers may also be used to cool and dehumidify air. If the entering water temperature is cooled below the entering wet-bulb temperature, both the dry- and wet-bulb temperatures of the leaving air are lowered. Dehumidification results if the leaving water temperature is maintained below the entering air dew point. Moreover, the final water temperature is determined by the sensible and latent heat absorbed from the air and the amount of circulated water, and it is 1 to 2°F below the final required dew-point temperature.

The air leaving a direct evaporative cooler being used as a dehumidifier is substantially saturated. Usually, the spread between dry-and wet-bulb temperatures is less than 1°F. The temperature difference between leaving air and leaving water depends on the difference between entering dry- and wet-bulb temperatures and on certain design features, such as the cross-sectional area and depth of the media or spray chamber, quantity and velocity of air, quantity of water, and the water distribution.

Direct evaporative coolers of all types perform some air cleaning. See the section on Air Cleaning and Sound Attenuation for detailed information.

Where the heat exchanger is directly wetted, a water hardness monitor for the recirculation water sump is recommended. Water hardness should be kept within 200 to 500 parts per million (ppm) to minimize plating out of dissolved solids from the sump water. To maintain its set point, the hardness monitor may initiate a sump dump cycle when it detects increased water hardness. In addition, the sump should have provisions for a fixed bleed so that extra makeup water is continuously introduced to dilute dissolved solids left behind when water evaporates from the wetted heat exchanger surface. Sumps should always be drained at the end of a duty cycle and refilled the next day when the system is turned on. For rooftop applications, sumps should be drained for freeze protection during low ambient temperatures.

Air-side control for a cooling system with a 55°F supply air set point may be set up as follows. The heat exchanger’s wet-side sprays or indirect evaporative cooling component should be activated whenever ambient dry-bulb temperatures exceed 65°F, if room return air is used on the wet side of the air-to-air heat exchanger. Air-conditioned buildings have a stable return air wet-bulb condition in the range of 60 to 65°F. Outdoor air may be usefully precooled when ambient dry-bulb temperatures exceed the return air wet-bulb condition.

Where outdoor air is used on the cold-air side of the heat exchanger, cooling may begin at ambient temperatures above 55°F, because the wet-bulb condition of outdoor air is always lower than its dry-bulb condition.

Parasitic losses generated by the heat exchanger static pressure penalty to supply and return air fans and by the water pump need to be evaluated. These losses may be mitigated by opening bypass dampers around the heat exchanger for pressure relief and shutting off the pump in the ambient temperature range of 55 to 65°F db. Where outdoor air is used on the wet side (scavenger air) of an air-to-air heat exchanger, this temperature range may be reduced somewhat. Comparing the energy penalty to the precooling energy avoided determines the optimum range of ambient conditions for this control strategy.

For variable-air-volume (VAV) supply and return fan systems, the static penalty reduces by the square of the airflow reduction from full design flow at summer peak design condition. As airflow rates decrease across an air-to-air heat exchanger, the WBDE increases, thereby providing better precooling. Where scavenger outdoor air is used for indirect evaporative cooling, the wet-side airflow rate is usually constant volume.

Winter heat recovery may be initiated at ambient temperatures below the 55°F supply air set point. Where building return air is used with an air-to-air heat exchanger, the 70 to 75°F return air condition is used to preheat makeup air for the building. For a VAV supply air system, Figure 8 shows the increased ventilation potential of a heat pipe air-to-air heat exchanger that uses face and bypass dampers on the supply air side to mix unheated outdoor air with preheated outdoor air to maintain the 55°F building supply air set point (Scofield and Bergman 1997). The heat pipe leaving air temperature may also be controlled with a tilt control (see Chapter 26 of the 2020 ASHRAE Handbook—HVAC Systems and Equipment). With a heat pipe economizer, a minimum outdoor air ventilation rate of 20% would not be breached until ambient temperatures dropped below −15°F.

Figure 8. Increased Winter Ventilation

Runaround coils control leaving supply air temperature with a three-way valve (see Figure 6). Because of their higher parasitic losses, these systems may require a wider range of ambient conditions where pressure-relief bypass dampers are open and the pump system shut down. Some projects limit activation of these recovery systems to ambient temperatures above 85°F or below 40°F.

Coupling indirect and direct evaporative cooling to a variable-air-volume (VAV) delivery system in arid climates can effectively eliminate requirements for mechanical refrigeration in many applications. Many cities in the western United States have summer design conditions suitable to deliver 55°F or lower supply air to a building using a 70% WBDE indirect and a 90% effective direct evaporative cooling system.

Figure 9 shows plan and elevation views of an air-handling unit using a sprayed heat pipe air-to-air heat exchanger and a wetted-media direct evaporative cooling section augmented by a final-stage chilled-water cooling coil (Scofield and Bergman 1997). The 70% indirect WBDE is achieved with a direct-sprayed heat pipe using a sump and a recirculation water system on the building return air side of the heat exchanger. The 90% effective direct evaporative cooling medium is split into two sections for two-stage cooling capacity control of the 55°F leaving air temperature. The direct evaporative cooling system also uses a sump and water recirculation. Supply-side heat pipe face and bypass dampers control the final supply air temperature (55°F) in both summer (when indirect sprays are on) and winter, to control the heat pipe’s heat recovery capacity. Heat pipe dampers on both sides of the heat exchanger are powered to full open to mitigate system parasitic losses during ambient temperature conditions when the value of energy recovered is exceeded by the fan energy penalty. The recirculation damper is used for morning warm-up of the building and for blending building return air with preheated outdoor air during extreme cold ambient conditions (see Figure 8).

Figure 9. Heat Pipe Air-Handling Unit

Table 3 uses ASHRAE bin weather data for the semi-arid climate of Sacramento, California, to illustrate potential cooling energy savings for a 10,000 cfm VAV design that turns down to 5000 cfm at winter design (Scofield and Bergman 1997). Compared to a conventional-refrigeration cooling VAV design with a 25% minimum outdoor air economizer, the two-stage evaporative cooling system reduces peak cooling load by 49% while introducing 100% outdoor air. For a building duty cycle of 8760 h per year, the ton-hour savings is 34,037, or a 60% reduction compared to the conventional air-side economizer system with mechanical cooling only. For ambient bin conditions of 62°F db/54°F wb to 57°F db/52°F wb, there are 2376 cooling hours per year (27% of the annual cooling hours) where a 90% wet-bulb depression efficiency (WBDE) direct evaporative cooling system may be used for the 55°F supply air requirement without refrigeration.

Figure 10 uses typical meteorological year (TMY) data for 14 cities in the western United States to illustrate the evaporative cooling annual refrigeration avoidance per 10,000 cfm of VAV supply air, compared to a 25% minimum outdoor air economizer (Scofield and Bergman 1997). For thermal energy storage (TES) applications, the two-stage evaporative cooling design may significantly reduce chiller plant storage capacity and refrigeration equipment first cost.

Figure 10. Refrigeration Reduction with Two-Stage Evaporative Cooling Design

Benefits of this design in dry climates include the following:

Areas with mild winter climates (e.g., the western U.S. coast) may use the heat available in building return air, through the air-to-air heat exchanger, to overheat supply air and add building humidification during the driest season of the year. The 4 in. section of direct evaporative cooling media (see Figure 8) is used in cool ambient conditions as a humidifier. Table 4 (Scofield and Bergman 1997) extends the Table 3 bin weather data for the Sacramento, California, site into winter ambient conditions. The table shows that 100% outdoor air may be introduced and humidity controlled between 54 and 32% for ambient conditions down to 37°F with a 60% heat pipe recovery effectiveness. There are only 146 bin hours below the 37°F ambient threshold during which the building recirculation air damper (see Figure 8) would have to open or additional heat be added with the hot-water coil to maintain the 55°F air delivery set point. The average winter temperature in Sacramento is 52.7°F.

In indirect evaporative cooling, outdoor supply air passes through an air-to-air heat exchanger and is cooled by evaporatively cooled air exhausted from the building or application. The two airstreams never mix or come into contact, so no moisture is added to the supply airstream. Cooling the building’s exhaust air results in a larger overall temperature difference across the heat exchanger and a greater cooling of the supply air. Indirect evaporative cooling requires only fan and water pumping power, so the coefficient of performance tends to be high. The principle of indirect evaporative cooling is effective in most air-conditioned buildings, because evaporative cooling is applied to exhaust air rather than to outdoor air.

Indirect evaporative cooling has been applied in a number of heat recovery applications (Mathur et al. 1993), such as plate heat exchangers (Scofield and DesChamps 1984; Wu and Yellot 1987), heat pipe exchangers (Mathur 1998; Scofield 1986), rotary regenerative heat exchangers, and two-phase thermosiphon loop heat exchangers (Mathur 1990). In residential air conditioning, the outdoor condensing unit can be evaporatively cooled to enhance performance (Mathur 1997; Mathur and Goswami 1995; Mathur et al. 1993). Indirect evaporative cooling with heat recovery is covered in detail in Chapter 26 of the 2020 ASHRAE Handbook—HVAC Systems and Equipment.

Both direct and indirect evaporative cooling may be used for area or spot cooling of industrial buildings. Both can be controlled either automatically or manually. In addition, evaporative coolers can supply tempered air during fall, winter, and spring. Gravity or power ventilators exhaust the air. Area cooling works well in buildings where personnel move about and workers are not subjected to concentrated, radiant heat sources. Area cooling may be used in either high- or low-bay industrial buildings, but may provide significant advantages in high-bay construction where cooling loads associated with roofs, lighting, and heat from equipment may be effectively eliminated by taking advantage of stratification. When cooling an area, ductwork should be designed to distribute air to the lower 10 ft of the space to ensure that cooler air is supplied to the workers.

Cooling requirements change from day to day and season to season, so if discharge grilles are used, they should be adjustable to prevent drafts. The horizontal blades of an adjustable grille can be adjusted so that air is discharged above workers’ heads rather than directly on them. In some cases, the air volume can be adjusted, either at each outlet or for the entire system, in which case the exhaust volume may need to be varied accordingly.

Spot cooling is a more efficient use of equipment when personnel work in one area. Cool air is brought to the spot at levels below 10 ft, and may even be delivered from floor outlets. Duct height may depend on the location of other equipment in the area. For best results, air velocity should be kept low. Controls may be automatic or manual, with the fan often operating throughout the year. Workers are especially appreciative of spot cooling in hot environments, such as in chemical plants and die casting shops, and near glass-forming machines, billet furnaces, and pig and ingot casting.

When spot-cooling a worker, the air volume depends on the throw of the air jet, worker activity, and amount of heat that must be overcome. Air volumes can vary from 200 to 5000 cfm per worker, with target velocities ranging between 200 to 4000 fpm. Outlets should be between 4 to 10 ft from workstations to avoid entrainment of warm air and to effectively blanket workers with cooler air. Workers should be able to control the direction of air discharge, because air motion that is appropriate for hot weather may be too great for cool weather or even cool mornings. Volume controls may be required to prevent overcooling the building and to minimize excessive grille blade adjustment.

Spot cooling is useful in rooms with elevated temperatures, regardless of climatic or geographical location. When the dry-bulb temperature of the air is below skin temperature, convection rather than evaporation cools workers. In these conditions, an 80°F airstream can provide comfort regardless of its relative humidity.

Electrical generators and motors are generally rated for a maximum ambient temperature of 104°F. When this temperature is exceeded, excessive temperatures develop in the electrical windings unless the load on the motor or generator is reduced. By providing evaporatively cooled air to the windings, this equipment may be safely operated without reducing the load. Likewise, transformer capacity can be increased using evaporative cooling.

Heat emitted by high-capacity electrical equipment may also be sufficient to raise the ambient condition to an uncomfortable level. With mill drive motors, an additional problem is often encountered with the commutator. If the air used to ventilate the motor is dry, the temperature rise through the motor results in a still lower relative humidity, at which the brush film can be destroyed, with unusual brush and commutator wear as well as the occurrence of dusting.

As a rule, a motor with a temperature rise of 25°F requires approximately 120 cfm of ventilating air per kilowatt-hour of loss. If inlet air to the motor is 95°F, air leaving the motor would be 120°F. This average motor temperature of over 107°F is 3°F higher than it should be for the normal 104°F ambient. The same quantity of 95°F db inlet air at 75°F wb can be cooled by a direct evaporative cooler with a 97% saturation effectiveness. The resulting 88°F average motor temperature eliminates the need for special high-temperature insulation and improves the motor’s ability to absorb temporary overloads. By comparison, an air quantity of 185 cfm is required if supplied by a cooler with 80% saturation effectiveness.

Figure 18 shows three basic arrangements for motor cooling. Air from the evaporative cooler may be directed on the motor windings, or into the room; the latter requires greater air volume to compensate for the building heat load. Direct evaporative cooler operation should be keyed to motor operation to ensure that (1) saturated or nearly saturated air is never introduced into a motor until it has had time to warm up, and (2) if more than one motor is served by a single system, air circulation through idle motors should be prevented.

Combustion turbines used for electric power production are normally rated at 59°F. Their performance is greatly influenced by the compressor inlet air temperature because temperature affects air density and therefore mass flow. As ambient temperature increases, demand on electric utilities increases and combustion turbine capacity decreases. Capacity recovery due to inlet air cooling is approximately 0.4%/°F (cooling). Direct and indirect evaporative cooling is beneficial to gas turbine performance in almost all climates because when the air is the hottest, it generally has the lowest relative humidity. Expected increases in output using direct evaporative cooling range from 5.8% in Albany, New York, to 14% in Yuma, Arizona. In addition to increasing gas turbine output, direct evaporative cooling also improves heat rate and reduces NO_x emissions.

Figure 18. Arrangements for Cooling Large Motors

For an installation of this type, the following precautions must be taken: (1) mist eliminators must be provided to stop entrainment of free moisture droplets, (2) coolers must be turned off at a temperature below 45°F to prevent icing, and (3) water quality must be monitored closely (Stewart 1999).

In the manufacture of textiles and tobacco and in processes such as spray coating, the required accurate relative humidity control can be provided by direct evaporative coolers. For example, textile manufacturing requires relatively high humidity and the machinery load is heavy, so a split system is customarily used to introduce free moisture directly into the room. The air handled is reduced to approximately 60% of that normally required by an all-outdoor-air, direct evaporative cooler.

Laundries have one of the most severe environments in which direct evaporative air cooling is applied, because heat is produced not only by the processing equipment, but by steam and water vapor as well. A properly designed direct evaporative cooler reduces the temperature in a laundry 5 to 10°F below the outdoor temperature. With only fan ventilation, laundries usually exceed the outdoor temperature by at least 10°F. Air distribution should be designed for a maximum throw of not more than 30 ft. A minimum circulated velocity of 100 to 200 fpm should prevail in the occupied space. Ducts can be located to discharge the air directly onto workers in exceptionally hot areas, such as pressing and ironing departments. For these outlets, manual control should be provided to direct the air where it is desired, with at least 500 to 1000 cfm at a target velocity of 600 to 900 fpm for each workstation.

Wood-processing plants and paper mills are good applications for evaporative cooling because of the high temperatures and gases associated with wood-processing equipment. Wood dust should be kept out of the recirculation sumps of evaporative coolers, because the dust contains microorganisms and worm larvae that will grow in sumps.

Because of the types of gases and particulates present in most paper plants, water-cooled systems are preferred over air-cooled systems. The most prevalent contaminant is wood dust. Chlorine gas, caustic soda, sulfur, hydrogen sulfide, and other compounds are also serious problems, because they accelerate the corrosion of steel and yellow metals. With more efficient air scrubbing, ambient air quality in and about paper mills has become less corrosive, allowing use of equipment with well-analyzed and properly applied coatings on coils and housings. Phosphor-free brazed coil joints should be used in areas where sulfur compounds are present.

Heat is readily available from processing operations and should be used whenever possible. Most plants have good-quality hot water and steam, which can be readily geared to unit heater, central station, or reheat use. Newer plant air-conditioning methods, including evaporative cooling, that use energy-conservation techniques (such as temperature stratification) lend themselves to this type of large structure. Chapter 26 has further information on air-conditioning of paper facilities.

An appropriate air-cooling system can be selected once preliminary heating and cooling loads are determined and criteria are established for temperature, humidity, pressure, and airflow control. The same considerations for selection apply to power-generating facilities and industrial facilities.

Chapter 29 describes evaporative cooling methods for mines.

The design criteria for farm animal environments and the need for cooling animal shelters are discussed in Chapter 24. Direct evaporative cooling is ideally suited to farm animal shelters because 100% outdoor air is used. Fresh air removes odors and reduces the harmful effects of ammonia fumes. At night and in the spring and fall, direct evaporative cooling can also be used for ventilation.

Equipment should be sized to change the air in the shelter in 1 to 2 min, assuming the ceiling height does not exceed 10 ft. This flow rate usually keeps the shelter at or below 80°F. In addition, conditions can be improved with portable or packaged spot coolers.

For poultry housing, most applications require an air change every 0.75 to 1.5 min, with the majority at 1 min. Placing the fans at the ends or the center of the house, with the direct evaporative cooler located at the opposite end, creates a tunnel ventilation system with an air velocity of 300 to 500 fpm. Fans are generally selected for a total pressure drop of 0.125 in. of water, which means that the direct evaporative cooling media cannot have a pressure drop in excess of 0.075 in. of water. Thus, to prevent an inadequate volume of air being pulled through the poultry house, the designer must carefully size the media selected.

Using direct evaporative cooling for poultry broiler houses decreases bird mortality, improves feed conversion ratio, and increases the growth rate. Poultry breeder houses are evaporatively cooled to improve egg production and fertility during warm weather. Evaporative cooling of egg layers improves feed conversion, shell quality, and egg size. When the ambient outdoor temperature exceeds 100°F, evaporative cooling is often the only way to keep a flock alive. Direct evaporative cooling is also used to cool swine farrowing and gestation houses to improve production.

Potatoes. Direct evaporative cooling for bulk potato storage should pass air directly through the pile. The ventilation and cooling system should provide 1.0 to 1.5 cfm/100 lb of potatoes. Average potato density is 45 lb/ft³ in the pile. Pile depths range from 12 to 20 ft, which creates a static pressure of 0.15 to 0.25 in. of water. Ventilation consists of fresh air inlets, return air openings, exhaust air openings, main air ducts, and lateral ducts with holes or slots to distribute air uniformly through the pile. Distribution ducts should be placed no farther apart than 80% of the potato pile depth, and should extend to within 18 in. of the storage walls. Ducts, the direct evaporative cooling media, and any refrigeration coils cause a static pressure ranging from 0.5 to 1.0 in. of water. Typically the total static pressure ranges from 0.75 to 1.25 in. of water, depending on the equipment. Air speed through each of the openings in the ventilation/cooling system should be as listed in Table 5.

Direct evaporative cooling media should be 90 to 95% effective, depending on the climate. In arid regions, 95% effective media are recommended. In more humid climates, such as in the midwestern and eastern United States, 90% effective media are commonly used. Air speed through the media should be 500 to 550 fpm to ensure high pad efficiency with low static-pressure penalty.

For more information, see Chapter 37 of the 2018 ASHRAE Handbook—Refrigeration.

Apples. Direct evaporative cooling for apple storage without refrigeration should distribute cool air to all parts of the storage. The evaporative cooler may be floor-mounted or located near the ceiling in a fan room. Air should be discharged horizontally at ceiling level. Because the prevailing wet-bulb temperature limits the degree of cooling, a cooler with maximum reasonable size should be installed to reduce the storage temperature rapidly and as close to the wet-bulb temperature as possible. Generally, a cooler designed to exchange air every 3 min (20 air changes per hour) is the largest that can be installed. This capacity provides a complete air change every 1 to 1.5 min (40 to 60 air changes per hour) when the storage is loaded.

For further information on apple storage, see Chapter 35 of the 2018 ASHRAE Handbook—Refrigeration.

Citrus. The chief purpose of evaporative-cooling fruits and vegetables is to provide an effective, inexpensive means of improving storage. However, it also serves a special function in the case of oranges, grapefruit, and lemons. Although mature and ready for harvest, citrus fruits are often still green. Color change (degreening) is achieved through a sweating process in rooms equipped with direct evaporative cooling. Air with a high relative humidity and a moderate temperature is circulated continuously during the operation. Ethylene gas, the concentration depending on the variety and intensity of green pigment in the rind, is discharged into the rooms. Ethylene destroys chlorophyll in the rind, allowing the yellow or orange color to become evident. During degreening, a temperature of 70°F and a relative humidity of 88 to 90% are maintained in the sweat room. (In the Gulf States, 82 to 85°F with 90 to 92% rh is used.) The evaporative cooler is designed to deliver 11 cfm per pound of fruit.

Direct and indirect evaporative cooling is also used as a supplement to refrigeration in the storage of citrus fruit. Citrus storage requires refrigeration in the summer, but the required conditions can often be obtained using evaporative cooling during the fall, winter, and spring when the outdoor wet-bulb temperature is low. For further information, see Chapter 36 of the 2018 ASHRAE Handbook—Refrigeration.

Proper regulation of greenhouse temperatures during the summer is essential for developing high-quality crops. The principal load on a greenhouse is solar radiation, which at sea level at about noon in the temperate zone is approximately 200 Btu/h · ft². Smoke, dust, or heavy clouds reduce the radiation load. Table 6 gives solar radiation loads for representative cities in the United States. Note that the values cited are average solar heat gains, not peak loads. Temporary rises in temperature inside a greenhouse can be tolerated; an occasional rise above design conditions is not likely to cause damage.

Not all solar radiation that reaches the inside of the greenhouse becomes a cooling load. About 2% of the total solar radiation is used in photosynthesis. Transpiration of moisture varies by crop, but typically uses about 48% of the solar radiation. This leaves 50% to be removed by the cooler. Example 2 shows a method for calculating the size of a greenhouse evaporative cooling system.

Example 2..

A direct evaporative cooler is to be installed in a 50 by 100 ft greenhouse. Design conditions are 92°F db and 73°F wb, and average solar radiation is 138 Btu/h · ft². An indoor temperature of 90°F db must not be exceeded at design conditions.

Solution: The direct evaporative air cooler is assumed to have a saturation effectiveness of 80%. Equation (2) may be used to determine the dry-bulb temperature of the air leaving the direct evaporative cooler:

The following equation, a modification of Equation (1), may be used to calculate the airflow rate that must be supplied by the direct evaporative cooler:

(4)

where

A	=	greenhouse floor area, ft2
It	=	total incident solar radiation, Btu/h · ft2 of receiving surface
60ρcp	=	density times specific heat of air times 60 min/h ≈ 1.0 Btu/ft3· °F at design conditions

For this problem

Figure 19. Schematics for 100% Outdoor Air Used in Hospital

Horizontal illumination from the direct rays of noonday summer sun with clear sky can be as much as 10,000 footcandles (fc); under clear glass, this is approximately 8500 fc. Crops such as chrysanthemums and carnations grow best in full sun, but many foliage plants, such as gloxinias and orchids, do not need more than 1500 to 2000 fc. Solar radiation is nearly proportional to light intensity. Thus, the greater the amount of shade, the smaller the cooling capacity required. A value of 100 fc is approximately equivalent to 3 Btu/h · ft². Although atmospheric conditions such as clouds and haze affect the relationship, this is a safe conversion factor. This relationship should be used instead of Table 5 when illumination can be determined by design or measurement.

Direct evaporative cooling for greenhouses may be under either positive or negative pressure. Regardless of the type of system used, the length of air travel should not exceed 160 ft. The temperature rise of the cool air limits the throw to this value. Air movement must be kept low because of possible mechanical damage to the plants, but it should generally not be less than 100 fpm in areas occupied by workers.

When used in a makeup air system comprised of a mixture of outdoor air and recirculated air, direct evaporative coolers function as scrubbers and reduce some gaseous contaminants found in outdoor air. These contaminants may concentrate in the recirculating water, so some water must be bled off. For more information regarding control of gaseous contaminants, see Chapter 46.

Evaporative coolers near sources of airborne nitric acid, chlorine, or ammonia absorb these chemicals, which can damage the cooler. The amount of soluble gases cleaned from the air depends on the air/water mixing, retention time, the water’s pH, and the bleed rate. When exposed to soluble gases, evaporative coolers should be operated with a high bleed rate.

Ozone levels of the airstream can be reduced using evaporative coolers and air washers. Ozone is fairly unstable in a watery solution, decaying to ordinary diatomic oxygen. The stability of ozone absorbed in water depends on water temperature, ozone concentration, and length of holding time. Higher room humidity can vastly improve the rate at which it decays back to oxygen (Sterling et al. 1985).

Legionnaires’ Disease. There have been no known cases of Legionnaires’ disease with air washers or wetted-media evaporative air coolers. This is can be attributed to the low temperature of the recirculated water, which is not conducive to Legionella bacteria growth, as well as the absence of aerosolized water carryover that could transmit the bacteria to a host (ASHRAE Guideline 12-2000).

Evaporative cooler media can attenuate sound attenuation somewhat. This insertion loss varies, depending on whether the media is wet or dry and whether the sound is traveling counter to (reverse flow) or with (forward flow) the airstream. Sound attenuation for 12 in. depth of rigid media can be found in Table 8. Components in the path of airflow in an air-handler plenum or duct provide some sound attenuation; one of the more effective at reducing sound pressure levels is the rigid-media direct evaporative cooler (DEC). Periannan (2013) performed tests in accordance with ASTM Standard E477-90 to measure insertion losses (1) with airflow in the direction of sound and (2) with reverse airflow. Measurements were also made at different velocities and for different media depths. As Table 8 shows, a rigid-media DEC is quite effective in reducing low-frequency noise levels, which are usually the most difficult to reduce. Rigid media should likely not be used purely as a sound attenuator, but noise reduction can be an additional benefit when these types are selected.

Direct evaporative cooling may be used in all climates to save cooling and humidification energy. In humid climates, the benefits of direct evaporation are realized during periods when outdoor air is warm and dry, but cooling savings are unlikely to be realized during peak design conditions. In more arid areas, direct evaporative cooling may partially or fully offset mechanical cooling at peak load conditions. Humidification energy savings may be realized during the heating season when outdoor air is used to provide cooling and humidification. If properly controlled, direct evaporative cooling can use waste heat otherwise rejected from buildings when outdoor air is used for cooling.

Indirect evaporative cooling may be used in all climates to save cooling and, in some applications, heating energy. In humid climates, indirect evaporative cooling may be used throughout the cooling cycle to precool outdoor air. Indirect evaporative cooling can be used to extend the range of 100% outdoor air ventilation to both higher and lower temperatures, and to increase the percentage of outdoor air a system can support at any given temperature through heat recovery. In high-humidity areas, indirect evaporative cooling may be used to (1) partially offset mechanical cooling requirements at peak load conditions and (2) provide better control over low-load humidity conditions by allowing use of smaller refrigeration equipment to provide ventilation over a wider range of outdoor air conditions. The cost of heating may be reduced when operating below temperatures at which minimum outdoor air quantities exceed the rates of ventilation required for free cooling by using heat recovered from building exhausts.

Typically, domestic service water is used for evaporative cooling to avoid excessive scaling and associated problems with poor water quality. In designing evaporative coolers, the cost of water treatment is included in the overall project cost. However, water cost is typically ignored for evaporative coolers because it is usually an insignificant part of the operational cost. Depending on the ambient dry-bulb temperature and wet-bulb depression for a specific location, the cost of water could become a significant part of the operational cost, because the greater the differential between dry- and wet-bulb temperatures, the greater the amount of water evaporated (Mathur 1997, 1998).