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Coils are used for air cooling with or without accompanying dehumidification. Examples of cooling applications without dehumidification are (1) precooling coils that use well water or other relatively high-temperature water to reduce load on the refrigerating equipment and (2) chilled-water coils that remove sensible heat from chemical moisture-absorption apparatus. The heat pipe coil is also used as a supplementary heat exchanger for preconditioning in air-side sensible cooling (see Chapter 26). Most coil sections provide air sensible cooling and dehumidification simultaneously.

The assembly usually includes a means of cleaning air to protect the coil from dirt accumulation and to keep dust and foreign matter out of the conditioned space. Although cooling and dehumidification are their principal functions, cooling coils can also be wetted with water or a hygroscopic liquid to aid in air cleaning, odor absorption, or frost prevention. Coils are also evaporatively cooled with a water spray to improve efficiency or capacity. Chapter 41 has more information on indirect evaporative cooling. For general comfort conditioning, cooling, and dehumidifying, the extended-surface (finned) cooling coil design is the most popular and practical.

In finned coils, the external surface of the tubes is primary, and the fin surface is secondary. The primary surface generally consists of rows of round tubes or pipes that may be staggered or placed in line with respect to the airflow. Flattened tubes or tubes with other nonround internal passageways are sometimes used. The inside surface of the tubes is usually smooth and plain, but some coil designs have various forms of internal fins or turbulence promoters (either fabricated or extruded) to enhance performance. The individual tube passes in a coil are usually interconnected by return bends (or hairpin bend tubes) to form the serpentine arrangement of multipass tube circuits. Coils are usually available with different circuit arrangements and combinations offering varying numbers of parallel water flow passes within the tube core (Figure 1).

Cooling coils for water, aqueous glycol, brine, or halocarbon refrigerants usually have aluminum fins on copper tubes, although copper fins on copper tubes and aluminum fins on aluminum tubes (excluding water) are also used. Adhesives are sometimes used to bond header connections, return bends, and fin-tube joints, particularly for aluminum-to-aluminum joints. Certain special-application coils feature an all-aluminum extruded tube-and-fin surface.

Common core tube outside diameters are 8, 10, 12.5, 16, 20, and 25 mm, with fins spaced 1.4 to 6.4 mm apart. Tube spacing ranges from 15 to 75 mm on equilateral (staggered) or rectangular (in-line) centers, depending on the width of individual fins and on other performance considerations. Fins should be spaced according to the job to be performed, with special attention given to air friction; possibility of lint accumulation; and frost accumulation, especially at lower temperatures.

Tube wall thickness and the required use of alloys other than copper are determined mainly by the coil’s working pressure and safety factor for hydrostatic burst (pressure). Maximum allowable working pressure (MAWP) for a coil is derived according to ASME’s Boiler and Pressure Vessel Code, Section VIII, Division 1 and Section II (ASTM material properties and stress tables). Pressure vessel safety standards compliance and certifications of coil construction may be required by regional and local codes before field installation. Fin type and header construction also play a large part in determining wall thickness and material. Local job site codes and applicable nationally recognized safety standards should be consulted in coil design and application.

This type of air-cooling coil normally has a shiny aluminum air-side surface. For special applications, the fin surface may be copper or have a brown or blue-green dip-process coating. These coatings protect the fin from oxidation that occurs when common airborne corrosive contaminants are diluted on a wet (dehumidifying) surface. Corrosion protection is increasingly important as indoor air quality (IAQ) guidelines call for higher percentages of outdoor air. Baked-on or anodized coating improves the expected service life compared to plain aluminum fins under similar conditions. Uncoated fins on non-dehumidifying, dry cooling coils are generally not affected by normal ambient airborne chemicals, except, to some extent, in a saline atmosphere. Once the coil is installed, little can be done to improve air-side protection.

Incoming airstream stratification across the coil face reduces coil performance. Proper air distribution is defined as having a measured airflow anywhere on the coil face that does not vary more than 20%. Moisture carryover at the coil’s air leaving side or uneven air filter loading are indications of uneven airflow through the coil. Normal corrective procedure is to install inlet air straighteners, or an air blender if several airstreams converge at the coil inlet face. Additionally, condensate water should never be allowed to saturate the duct liner or stand in the drain pan (trough). The coil frame (particularly its bottom sheet metal member) should not be allowed to sit in a pool of water, to prevent rusting.

When selecting a coil, consider the following factors:

Chapters 17 and 18 of the 2021 ASHRAE Handbook—Fundamentals provide information on load calculation.

Air quantity is affected by factors such as design parameters, codes, space, and equipment. Resistance through the air circuit influences fan power and speed. This resistance may be limited to allow the use of a given size fan motor, to keep operating expense low, or because of sound-level requirements. Air friction loss across the cooling coil (in summation with other series air-pressure drops for elements such as air filters, water sprays, heating coils, air grilles, and ductwork) determines the static pressure requirement for the complete airway system. The static pressure requirement is used in selecting fans and drives to obtain the design air quantity under operating conditions. See Chapter 21 for a description of fan selection.

The conditioned-air face velocity is determined by economic evaluation of initial and operating costs for the complete installation as influenced by (1) heat transfer performance of the specific coil surface type for various combinations of face areas and row depths as a function of air velocity; (2) air-side frictional resistance for the complete air circuit (including coils), which affects fan size, power, and sound-level requirements; and (3) condensate water carryover considerations. Allowable friction through the water or brine coil circuitry may be dictated by the pressure available from a given size pump and pump motor, as well as the same economic factors governing the air side made applicable to the water side. Additionally, the adverse effect of high cooling-water velocities on erosion/corrosion of tube walls is a major factor in sizing and circuitry to keep tube velocity below the recommended maximums. On larger coils, water pressure drop limits of 45 to 60 kPa usually keep such velocities within acceptable limits of 0.6 to 1.2 m/s, depending on circuit design.

Coil ratings are based on a uniform velocity. Design interference with uniform airflow through the coil makes predicting coil performance difficult as well as inaccurate. Such airflow interference may be caused by air entering at odd angles or by inadvertent partial blocking of the coil face. To obtain rated performance, the volumetric airflow quantity must be adjusted on the job to that at which the coil was rated and must be kept at that value. At start-up for air balance, the most common causes of incorrect airflow are the lack of altitude correction to standard air (where applicable) and ductwork problems. At commissioning, the most common causes of an air quantity deficiency are filter fouling and dirt or frost collection on the coils. These difficulties can be avoided through proper design, start-up checkout, and regular servicing.

The required total heat capacity of the cooling coil should be in balance with the capacity of other refrigerant system components such as the compressor, water chiller, condenser, and refrigerant liquid metering device. Chapter 5 of the 2022 ASHRAE Handbook—Refrigeration describes methods of estimating balanced system capacity under various operating conditions when using direct-expansion coils for both factory- and field-assembled systems.

For dehumidifying coils, it is important that the proper amount of surface area be installed to obtain the ratio of air-side sensible-to-total heat required to maintain air dry-bulb and wet-bulb temperatures in the conditioned space. This is an important consideration when preconditioning is done by reheat arrangement. The method for calculating the sensible and total heat loads and leaving-air conditions at the coil to satisfy the sensible-to-total heat ratio required for the conditioned space is covered in Appendix D of Cooling and Heating Load Calculation Principles (Pedersen et al. 1998).

The same room air conditions can be maintained with different air quantities (including outdoor and return air) through a coil. However, for a given total air quantity with fixed percentages of outdoor and return air, there is only one set of air conditions leaving the coil that will precisely maintain room design air conditions. Once air quantity and leaving-air conditions at the coil have been selected, there is usually only one combination of face area, row depth, and air face velocity for a given coil surface that will precisely maintain the required room ambient conditions. Therefore, in making final coil selections it is necessary to recheck the initial selection to ensure that the leaving-air conditions, as calculated by a coil selection computer program or other procedure, match those determined from the cooling load estimate.

Coil ratings and selections can be obtained from manufacturers’ catalogs. Most catalogs contain extensive tables giving the performance of coils at various air and water velocities and entering humidity and temperatures. Most manufacturers provide computerized coil selection programs to potential customers. The final choice can then be made based on system performance and economic requirements.

A cooling coil’s airflow resistance (air friction) depends on the tube pattern and fin geometry (tube size and spacing, fin configuration, and number of in-line or staggered rows), coil face velocity, and amount of moisture on the coil. The coil air friction may also be affected by the degree of aerodynamic cleanliness of the coil core; burrs on fin edges may increase coil friction and increase the tendency to pocket dirt or lint on the faces. A completely dry coil, removing only sensible heat, offers approximately one-third less resistance to airflow than a dehumidifying coil removing both sensible and latent heat.

For a given surface and airflow, increasing the number of rows or fins increases airflow resistance. Therefore, final selection involves economic balancing of the initial cost of the coil against the operating costs of the coil geometry combinations available to adequately meet the performance requirements.

The aluminum fin surfaces of new dehumidifying coils tend to inhibit condensate sheeting action until they have aged for a year. Hydrophilic aluminum fin surface coatings reduce water droplet surface tension, producing a more evenly dispersed wetted surface action at initial start-up. Manufacturers have tried different methods of applying such coatings, including dipping the coil into a tank, coating the fin stock material, or subjecting the material to a chemical etching process. Tests have shown as much as a 30% reduction in air pressure drop across a hydrophilic coil as opposed a new untreated coil.

The heat transmission rate of air passing over a clean tube (with or without extended surface) to a fluid flowing in it is impeded principally by three thermal resistances: (1) surface air-side film thermal resistance from the air to the surface of the exterior fin and tube assembly; (2) metal thermal resistance to heat conductance through the exterior fin and tube assembly; and (3) in-tube fluid-side film thermal resistance, which impedes heat flow between the internal surface of the metal and the fluid flowing within the tube. For some applications, an additional thermal resistance is factored in to account for external and/or internal surface fouling. Usually, the combination of metal and tube-side film resistance is considerably lower than the air-side surface resistance.

For a reduction in thermal resistance, the fin surface is fabricated with die-formed corrugations instead of the traditional flat design. At low airflows or wide fin spacing, the air-side transfer coefficient is virtually the same for flat and corrugated fins. Under normal comfort conditioning operation, the corrugated fin surface is designed to reduce the boundary air film thickness by undulating the passing airstream within the coil; this produces a marked improvement in heat transfer without much airflow penalty. Further fin enhancements, including louvered and lanced fin designs, have been driven by the desire to duplicate throughout the coil depth the thin boundary air film characteristic of the fin’s leading edge. Louvered fin design maximizes the number of fin surface leading edges throughout the entire secondary surface area and increases the external secondary surface area A_s through the multiplicity of edges.

Where an application allows economical use of coil construction materials, the mass and size of the coil can be reduced when boundary air and water films are lessened. For example, the exterior surface resistance can be reduced to nearly the same as the fluid-side resistance by using lanced and/or louvered fins. External as well as internal tube fins (or internal turbulators) can economically decrease overall heat transfer surface resistances. Also, water sprays applied to a flat fin coil surface may increase overall heat transfer slightly, although they may better serve other purposes such as air and coil cleaning.

Heat transfer between the cooling medium and the airstream across a coil is influenced by the following variables:

With water coils, only the water temperature rises. With coils of volatile refrigerants, an appreciable pressure drop and a corresponding change in evaporating temperature through the refrigerant circuit often occur. Alternative refrigerants to R-22, such as R-407C, which has a temperature glide, have an evaporation temperature rise of 4 to 7 K through the evaporator. This must be considered in design and performance calculation of the coil. A compensating pressure drop in the coil may partially, or even totally, compensate for the low-side temperature glide of a zeotropic refrigerant blend. Rating direct-expansion coils is further complicated by the refrigerant evaporating in part of the circuit and superheating in the remainder. Thus, for halocarbon refrigerants, a cooling coil is tested and rated with a specific distributing and liquid-metering device, and the capacities are stated with the superheat condition of the leaving vapor.

At a given air mass velocity, performance depends on the turbulence of airflow into the coil and the uniformity of air distribution over the coil face. The latter is necessary to obtain reliable test ratings and realize rated performance in actual installations. Air resistance through the coils helps distribute air properly, but the effect is frequently inadequate where inlet duct connections are brought in at sharp angles to the coil face. Reverse air currents may pass through a portion of the coils. These currents reduce capacity but can be avoided with proper inlet air vanes or baffles. Air blades may also be required. Remember that coil performance ratings (AHRI Standard 410) represent optimum conditions resulting from adequate and reliable laboratory tests (ASHRAE Standard 33).

For cases when available data must be extended, for arriving at general design criteria for a single, unique installation, or for understanding the calculation progression, the following material and illustrative examples for calculating cooling coil performance are useful guides.

The performance of sensible cooling coils depends on the following factors. See the section on Symbols for an explanation of the variables.

The sensible heat cooling capacity q_t d of a given coil is expressed by the following equation:

with

Assuming no extraneous heat losses, the same amount of sensible heat is lost from the airstream:

with

The same amount of sensible heat is absorbed by the coolant; for a nonvolatile type, it is

For a nonvolatile coolant in thermal counterflow with the air, the mean temperature difference in Equation (1a) is expressed as

Proper temperature differences for various cross-flow situations are given in many texts, including Mueller (1973). These calculations are based on various assumptions, among them that U for the total external surface is constant. Although this assumption is generally not valid for multirow coils, using cross-flow temperature differences from Mueller (1973) or other texts should be preferable to Equation (4), which applies only to counterflow. However, using the log mean temperature difference is widespread.

The overall heat transfer coefficient U_o for a given coil design, whether bare-pipe or finned-type, with clean, nonfouled surfaces, consists of the combined effect of three individual heat transfer coefficients:

These three individual coefficients acting in series form an overall coefficient of heat transfer in accordance with the material given in Chapters 4 and 25 to 27 of the 2021 ASHRAE Handbook—Fundamentals.

For a bare-pipe coil, the overall coefficient of heat transfer for sensible cooling (without dehumidification) can be expressed by a simplified basic equation:

When pipe or tube walls are thin and of high-conductivity material (as in typical heating and cooling coils), the term (D_o – D_i)/2000k in Equation (5a) frequently becomes negligible and is generally disregarded. (This effect in typical bare-pipe cooling coils seldom exceeds 1 to 2% of the overall coefficient.) Thus, the overall coefficient for bare pipe in its simplest form is

For finned coils, the equation for the overall coefficient of heat transfer can be written

where the fin effectiveness η allows for the resistance to heat flow encountered in the fins. It is defined as

For typical cooling surface designs, the surface ratio B ranges from about 1.03 to 1.15 for bare-pipe coils and from 10 to 30 for finned coils. Chapter 4 of the 2021 ASHRAE Handbook—Fundamentals describes how to estimate fin efficiency and calculate the tube-side heat transfer coefficient f_r for nonvolatile fluids. Table 3 in AHRI Standard 410 lists thermal conductivity k of standard coil materials.

Estimating the air-side heat transfer coefficient f_a is more difficult because well-verified general predictive techniques are not available. Hence, direct use of experimental data is usually necessary. For plate fin coils, some correlations that satisfy several data sets are available (Kusuda 1970; McQuiston 1981). Webb (1980) reviewed air-side heat transfer and pressure drop correlations for various geometries. Mueller (1973) and Chapter 4 of the 2021 ASHRAE Handbook—Fundamentals provide guidance on this subject.

For analyzing a given heat exchanger, the concept of effectiveness is useful. Expressions for effectiveness have been derived for various flow configurations and can be found in Kusuda (1970) and Mueller (1973). The cooling coils covered in this chapter actually involve various forms of cross-flow. However, the case of counterflow is addressed here to illustrate the value of this concept. The air-side effectiveness E_a for counterflow heat exchangers is given by the following equations:

with

or

with

and

Note the following two special conditions:

With a given design and arrangement of heat transfer surface used as cooling coil core material for which basic physical and heat transfer data are available to determine U_o from Equations (5a), (5b), and (5c), the selection, sizing, and performance calculation of sensible cooling coils for a particular application generally fall into either of two categories:

The two categories of application problems are shown in Examples 1 and 2, respectively:

The preceding equations and examples demonstrate the method for calculating thermal performance of sensible cooling coils that operate with a dry surface. However, when cooling coils operate wet or act as dehumidifying coils, performance cannot be predicted without including the effect of air-side moisture (latent heat) removal.

A dehumidifying coil normally removes both moisture and sensible heat from entering air. In most air-conditioning processes, the air to be cooled is a mixture of water vapor and dry air gases. Both lose sensible heat when in contact with a surface cooler than the air. Latent heat is removed through condensation only on the parts of the coil where the surface temperature is lower than the dew point of the air passing over it. Figure 2 in Chapter 4 shows the assumed psychrometric conditions of this process. As the leaving dry-bulb temperature drops below the entering dew-point temperature, the difference between leaving dry-bulb temperature and leaving dew point for a given coil, airflow, and entering air condition is lessened.

When the coil starts to remove moisture, the cooling surfaces carry both the sensible and latent heat load. As the air approaches saturation, each degree of sensible cooling is nearly matched by a corresponding degree of dew-point decrease. The latent heat removal per degree of dew-point change is considerably greater. The following table compares the amount of moisture removed from air at standard barometric pressure that is cooled from 16 to 15°C at both wet and dry conditions.

For volatile refrigerant coils, the refrigerant distributor assembly must be tested at the higher and lower capacities of its rated range. Testing at lower capacities checks whether the refrigerant distributor provides equal distribution and whether the control is able to modulate without hunting. Testing at higher capacities checks the maximum feeding capacity of the flow control device at the greater pressure drop that occurs in the coil system.

Most manufacturers develop and produce their own performance rating tables using data obtained from suitable tests. ASHRAE Standard 33 specifies the acceptable method of lab-testing coils. AHRI Standard 410 gives a method for rating thermal performance of dehumidifying coils by extending data from laboratory tests on prototypes to other operating conditions, coil sizes, and row depths. To account for simultaneous transfer of both sensible and latent heat from the airstream to the surface, AHRI Standard 410 uses essentially the same method for arriving at cooling and dehumidifying coil thermal performance as determined by McElgin and Wiley (1940) and described in the context of Standard 410 by Anderson (1970). In systems operating at partial flow such as in thermal storage, accurate performance predictions for 2300 < Re < 4000 flows have been obtained by using the Gnielinski correlation. This work was presented in detail comparable to Standard 410 by Mirth et al. (1993).

The potential or driving force for transferring total heat q_t from the airstream to the tube-side coolant is composed of two components in series heat flow: (1) an air-to-surface air enthalpy difference (h_a − h_s) and (2) a surface-to-coolant temperature difference (t_s − t_r).

Figure 7 is a typical thermal diagram for a coil in which the air and a nonvolatile coolant are arranged in counterflow. The top and bottom lines in the diagram indicate, respectively, changes across the coil in the airstream enthalpy h_a and the coolant temperature t_r. To illustrate continuity, the single middle line in Figure 7 represents both surface temperature t_s and the corresponding saturated air enthalpy h_s, although the temperature and air enthalpy scales do not actually coincide as shown. The differential surface area dA_w represents any specific location within the coil thermal diagram where operating conditions are such that the air-surface interface temperature t_s is lower than the local air dew-point temperature. Under these conditions, both sensible and latent heat are removed from the airstream, and the cooler surface actively condenses water vapor.

Neglecting the enthalpy of condensed water vapor leaving the surface and any radiation and convection losses, the total heat lost from the airstream in flowing over dA_w is

This same total heat is transferred from the airstream to the surface interface. According to McElgin and Wiley (1940),

The total heat transferred from the air-surface interface across the surface elements and into the coolant is equal to that given in Equations (8) and (9):

The same quantity of total heat is also gained by the nonvolatile coolant in passing across dA_w:

If Equations (9) and (10) are equated and the terms rearranged, an expression for the coil characteristic C is obtained:

Equation (12) shows the basic relationship of the two components of the driving force between air and coolant in terms of three principal thermal resistances. For a given coil, these three resistances of air, metal, and in-tube fluid (R_aw, R_mw, and R_r) are usually known or can be determined for the particular application, which gives a fixed value for C. Equation (12) can then be used to determine point conditions for the interrelated values of airstream enthalpy h_a, coolant temperature t_r, surface temperature t_s, and enthalpy h_s of saturated air corresponding to the surface temperature. When both t_s and h_s are unknown, a trial-and-error solution is necessary; however, this can be solved graphically by a surface temperature chart such as Figure 8.

Figure 9 shows a typical thermal diagram for a portion of the coil surface when it is operating dry. The illustration is for counterflow with a halocarbon refrigerant. The diagram at the top of the figure shows a typical coil installation in an air duct with tube passes circuited countercurrent to airflow. Locations of the entering and leaving boundary conditions for both air and coolant are shown.

The thermal diagram in Figure 9 is the same type as in Figure 7, showing three lines to illustrate local conditions for the air, surface, and coolant throughout a coil. The dry/wet boundary conditions are located where the coil surface temperature t_sb equals the entering air dew-point temperature Thus, the surface area A_d to the left of this boundary is dry, with the remainder A_w of the coil surface area operating wet.

When using fluids or halocarbon refrigerants in a thermal counterflow arrangement as shown in Figure 9, the dry/wet boundary conditions can be determined from the following relationships:

The value of h_ab from Equation (14) serves as an index of whether the coil surface is operating fully wetted, partially dry, or completely dry, according to the following three limits:

Other dry/wet boundary properties are then determined:

The dry surface area A_d required and capacity q_td are calculated by conventional sensible heat transfer relationships, as follows.

The overall thermal resistance R_o comprises three basic elements:

with

The mean difference between air dry-bulb temperature and coolant temperature, using symbols from Figure 9, is

The dry surface area required is

The air-side total heat capacity is

From the coolant side,

The wet surface area A_w and capacity q_tw are determined by the following relationships, using terminology in Figure 9.

For a given coil size, design, and arrangement, the fixed value of the coil characteristic C can be determined from the ratio of the three prime thermal resistances for the job conditions:

Knowing coil characteristic C for point conditions, the interrelations between airstream enthalpy h_a, coolant temperature t_r, and surface temperature t_s and its corresponding enthalpy of saturated air h_s can be determined by using a surface temperature chart (Figure 8) or by a trial-and-error procedure using Equation (24):

The mean effective difference in air enthalpy between airstream and surface from Figure 9 is

Similarly, the mean temperature difference between surface and coolant is

The wet surface area required, calculated from air-side enthalpy difference, is

Calculated from the coolant-side temperature difference,

The air-side total heat capacity is

The enthalpy h_fw of condensate removed is

where c_pw = specific heat of water = 4.18 kJ/(kg · K).

Note that h_fw for normal air-conditioning applications is about 0.5% of the airstream enthalpy difference (h_a1 − h_a2) and is usually neglected.

The coolant-side heat capacity is

The total surface area requirement of the coil is

The total heat capacity for the coil is

The leaving-air dry-bulb temperature is found by the method shown in Figure 10, which represents part of a psychrometric chart showing the air saturation curve and lines of constant air enthalpy closely corresponding to constant-wet-bulb temperature lines.

For a given coil and air quantity, a straight line projected through the entering and leaving-air conditions intersects the air saturation curve at a point denoted as the effective coil surface temperature . Thus, for fixed entering air conditions t_a1 and h_a1 and a given effective surface temperature t_s, leaving air dry bulb t_a2 increases but is still located on this straight line if air quantity is increased or coil depth is reduced. Conversely, a decrease in air quantity or an increase in coil depth produces a lower t_a2 that is still located on the same straight-line segment.

An index of the air-side effectiveness is the heat transfer exponent c, defined as

This exponent c, sometimes called the number of air-side transfer units NTU_a, is also defined as

The temperature drop (t_a1 − t_a2) of the airstream and mean temperature difference between air and effective surface in Equation (32) are shown at the top of Figure 10.

Knowing the exponent c and entering and leaving enthalpies h_a1 and h_a2 for the airstream, the enthalpy of saturated air corresponding to effective surface temperature is calculated as follows:

After finding the value of that corresponds to from the saturated air enthalpy tables, the leaving-air dry-bulb temperature can be determined:

The air-side sensible heat ratio SHR can then be calculated:

For thermal performance of a coil to be determined from the foregoing relationships, values of the following three principal resistances to heat flow between air and coolant must be known:

In AHRI Standard 410, metal thermal resistance R_m is calculated based on the physical data, material, and arrangement of the fin and tube elements, together with the fin efficiency E for the specific fin configuration. R_m is variable as a weak function of the effective air-side heat transfer coefficient f_a for a specific coil geometry, as shown in Figure 11.

For wetted-surface application, Brown (1954), with certain simplifying assumptions, showed that f_a is directly proportional to the rate of change m″ of saturated air enthalpy h_s with the corresponding surface temperature t_s. This slope m″ of the air enthalpy saturation curve is shown in the small inset graph at the top of Figure 11.

The abscissa for f_a in the main graph of Figure 11 is an effective value, which, for a dry surface, is the simple thermal resistance reciprocal 1/R_ad. For a wet surface, f_a is the product of the thermal resistance reciprocal 1/R_aw and the multiplying factor m″/c_p. AHRI Standard 410 outlines a method for obtaining a mean value of m″/c_p for a given coil and job condition. The total metal resistance R_m in Figure 11 includes the resistance R_t across the tube wall. For most coil designs, R_t is quite small compared to the resistance R_f through the fin metal.

The air-side thermal resistances R_ad and R_aw for dry and wet surfaces, together with their respective air-side pressure drops Δp_st/N_r and Δp_sw/N_r, are determined from tests on a representative coil model over the full range in the rated airflow. Typical plots of experimental data for these four performance variables versus coil air face velocity V_a at 20°C are shown in Figure 12.

If water is used as the tube-side coolant, the heat transfer coefficient f_r is calculated from Equation (8) in AHRI Standard 410. For evaporating refrigerants, many predictive techniques for calculating coefficients of evaporation are listed in Table 2 in Chapter 5 of the 2021 ASHRAE Handbook—Fundamentals. The most verified predictive technique is the Shah correlation (Shah 1976, 1982). A series of tests is specified in AHRI Standard 410 for obtaining heat transfer data for direct-expansion refrigerants inside tubes of a given diameter.

ASHRAE Standard 33 specifies laboratory apparatus and instrumentation, including procedure and operating criteria for conducting tests on representative coil prototypes to obtain basic performance data. Procedures are available in AHRI Standard 410 for reducing these test data to the performance parameters necessary to rate a line or lines of various air coils. This information is available from various coil manufacturers for use in selecting AHRI Standard 410 certified coils.

The following example shows a method for selecting coil size, row depth, and performance data to satisfy specified job requirements. The application is for typical cooling and dehumidifying coil selection under conditions in which a part of the coil surface on the entering air side operates dry, with the remaining surface wet with condensing moisture. Figure 9 shows the thermal diagram, dry/wet boundary conditions, and terminology used in the problem solution.

The following calculation of refrigeration load distinguishes between the true sensible and latent heat loss of the air, which is accurate within the data’s limitations. This division will not correspond to load determination obtained from approximate factors or constants.

The total refrigeration load q_t of a cooling and dehumidifying coil (or air washer) per unit mass of dry air is indicated in Figure 13 and consists of the following components:

The preceding components are related by the following equation:

If only the total heat value is desired, it may be computed by

where

h1 and h2 = enthalpy of air at points 1 and 2, respectively

W1 and W2 = humidity ratio at points 1 and 2, respectively

hw3 = enthalpy of saturated liquid at final temperature t3

If a breakdown into latent and sensible heat components is desired, the following relations may be used.

Latent heat may be found from

where

hfg4 = enthalpy representing latent heat of water vapor at condensing temperature t4

Sensible heat may be shown to be

or

where

hg4 = hfg4 + hw4 = enthalpy of saturated water vapor at condensing temperature t4

hw4 = enthalpy of saturated liquid at condensing temperature t4

The last term in Equation (39a) is the heat of subcooling the condensate from t₄ to its final temperature t₃. Then,

The final condensate temperature t₃ leaving the system is subject to substantial variations, depending on the method of coil installation, as affected by coil face orientation, airflow direction, and air duct insulation. In practice, t₃ is frequently the same as the leaving wet-bulb temperature. Within the normal air-conditioning range, precise values of t₃ are not necessary because the heat q_t of condensate removed from the air usually represents about 0.5 to 1.5% of the total refrigeration cooling load.

If the coil is to deliver its full cooling capacity, both its internal and external surfaces must be clean. The tubes generally stay clean in pressurized water or brine systems. Tube surfaces can be cleaned in a number of ways, but are often washed with low-pressure water spray and mild detergent. Water coils should be completely drained if freezing is possible. When coils use built-up system refrigerant evaporators, oil can accumulate. Check and drain oil occasionally, and check for leaks and refrigerant dryness.

Air Side. The best maintenance for the outside finned area is consistent inspection and service of inlet air filters. Surface cleaning of the coil with pressurized hot water and a mild detergent should be done only when necessary (primarily when a blockage occurs under severe fin-surface-fouling service conditions, or bacterial growth is seen or suspected). Pressurized cleaning is more thorough if done first from the air exit side of the coil and then from the air entry side. Foaming chemical sprays and washes should be used instead of high pressure on fragile fins, or when fin density is too restrictive to allow proper in-depth cleaning with pressurized water spray. In all cases, limit spray water temperature to below 65°C on evaporator coils containing refrigerant. In cases of marked neglect or heavy-duty use (especially in restaurants where grease and dirt have accumulated) coils must sometimes be removed and the accumulation washed off with steam, compressed air and water, or hot water.

The surfaces can also be brushed and vacuumed. Best practice is to inspect and service the filters frequently. Also, condensate drain pan(s) and their drain lines, including open drain areas, should be kept clean and clear at all times.

Water Side. The best service for the inside tube coil surface is keeping the circulated fluid (water or glycol) free of sediment, corrosive products, and biological growth. Maintaining proper circulated water chemistry and velocity and filtering out solids should minimize the water-side fouling factor. If large amounts of scale form when untreated water is used as coolant, chemical or mechanical (rod) cleaning of internal surfaces at frequent intervals is necessary. A properly maintained chilled-water system using a glycol solution as the circulated fluid is not considered to ever have water-side fouling, as such, but glycol solutions must be analyzed seasonally to determine alkalinity, percent concentration, and corrosion inhibitor condition. Consult a glycol expert or the manufacturer’s agent for detailed recommendations on proper use and control of glycol solutions.

Refrigerant Side. Moisture content of the refrigerant should be checked yearly, and acidity of the compressor oil as often as monthly. For built-up direct-expansion (DX) systems, normal is ≤50 mg/kg of moisture for systems with mineral oil, and ≤100 mg/kg for some refrigerant types in systems with polyol ester (POE) compressor oil. The actual values are set by the compressor manufacturer, and should be checked and verified during operation by moisture-indicating sight-glass viewing, and yearly by laboratory sample analysis reports. Depending on temperature and velocity, excessive moisture coming to the cooling coil through refrigerant might cause internal freeze-up around the coil’s expansion valve. Generally, moisture in a system contributes to formation of acids, sludge, copper plating, and corrosion. Oil breakdown (by excessive compressor overheat) can form organic, hydrofluoric, and hydrochloric acids in the lubricant oil, all of which can corrode copper.

The refrigerant must be of high quality and meet AHRI Standard 700 purity requirements. Application design ratings of the coil or coil bank are based on purity and dryness. This is a primary requirement of refrigerant evaporator coil maintenance.

DX coil replacement often coincides with refrigerant change-out to a chlorine-free refrigerant. Special care should be taken to ensure that the system is retrofitted in accordance with the compressor and refrigerant manufacturers’ written procedures.

Aa = coil face or frontal area, m2

Ad = dry external surface area, m2

Ao = total external surface area, m2

Ap = exposed external prime surface area, m2

As = external secondary surface area, m2

Aw = wet external surface area, m2

B = ratio of external to internal surface area, dimensionless

C = coil characteristic as defined in Equations (12) and (23), (kg · K)/kJ

c = heat transfer exponent, or NTUa, as defined in Equations (31) and (32), dimensionless

co = heat transfer exponent, as defined in Equation (7d), dimensionless

cp = specific heat of humid air = 1.019 kJ/(kg · K) for cooling coils

cpw = specific heat of water = 4.18 kJ/(kg · K)

cr = specific heat of nonvolatile coolant, kJ/(kg · K)

Di = tube inside diameter, mm

Do = tube outside diameter, mm

Ea = air-side effectiveness defined in Equation (7b), dimensionless

Fs = coil core surface area parameter = (external surface area)/(face area) (no. of rows deep)

f = convection heat transfer coefficient, W/(m2 · K)

h = air enthalpy (actual in airstream or saturation value at surface temperature), kJ/kg

Δhm = mean effective difference of air enthalpy, as defined in Equation (25), kJ/kg

k = thermal conductivity of tube material, W/(m · K)

M = ratio of nonvolatile coolant-to-air temperature changes for sensible heat cooling coils, as defined in Equation (7e), dimensionless

m″ = rate of change of air enthalpy at saturation with air temperature, kJ/(kg · K)

Nr = number of coil rows deep in airflow direction, dimensionless

Δpsd = isothermal dry surface air-side at standard conditions (20°C, 101.325 kPa), Pa

Δpsw = wet surface air-side at standard conditions (20°C, 101.325 kPa), Pa

q = heat transfer capacity, W

qe = latent heat removed from entering air to condense moisture, kJ/kg

qs = sensible heat removed from entering air, kJ/kg

qt = total refrigeration load of cooling and dehumidifying coil, kJ/kg

qw = sensible heat removed from condensate to cool it to leaving temperature, kJ/kg

R = thermal resistance, referred to external area Ao, (m2 · K)/W

SHR = ratio of air sensible heat to air total heat, dimensionless

t = temperature, °C

Δtm = mean effective temperature difference, air dry bulb to coolant temperature, K

Δtms = mean effective temperature difference, surface-to-coolant, K

= mean effective temperature difference, air dry bulb to effective surface temperature , K

Uo = overall sensible heat transfer coefficient, W/(m2 · K)

Va = coil air face velocity at 20°C, m/s

W = air humidity ratio, kilograms of water per kilogram of air

w = mass flow rate, kg/s

y = ratio of nonvolatile coolant temperature rise to airstream enthalpy drop, as defined in Equation (13), (kg · K)/kJ

η = fin effectiveness, as defined in Equation (6), dimensionless

ρa = air density = 1.20 kg/m3 at 20°C at sea level

Superscripts

′ = wet bulb

″ = dew point

Subscripts

1 = condition entering coil

2 = condition leaving coil

a = airstream

ab = air, dry/wet boundary

ad = dry air

aw = wet air

b = dry/wet surface boundary

d = dry surface

e = latent

f = fin (with R); saturated liquid water (with h)

g = saturated water vapor

i = installed, selected (with Ao, Nr)

m = metal (with R) and mean (with other symbols)

md = dry metal

mw = wet metal

o = overall (except for A)

r = coolant

rb = coolant dry/wet boundary

s = surface (with d, t, and w) and saturated (with h)

= effective surface

sb = surface dry/wet boundary

t = tube (with R) and total (with s and q)

td = total heat capacity, dry surface

tw = total heat capacity, wet surface

w = water (with ρ), condensate (with h and subscript number), and wet surface (with other symbols)

AHRI. 2001. Forced-circulation air-cooling and air-heating coils. ANSI/AHRI Standard 410-01. Air-Conditioning, Heating, and Refrigeration Institute, Arlington, VA.

AHRI. 2015. Specification for fluorocarbon refrigerants. Standard 700-2015. Air-Conditioning, Heating, and Refrigeration Institute, Arlington, VA.

Anderson, S.W. 1970. Air-cooling and dehumidifying coil performance based on ARI Industrial Standard 410-64. In Heat and mass transfer to extended surfaces, ASHRAE Symposium CH-69-3, pp. 22-28.

ASHRAE. 2016. Methods of testing forced circulation air-cooling and air heating coils. Standard 33-2016.

ASME. 2015. Code on nuclear air and gas treatment. ANSI/ASME Standard AG-1-2015. American Society of Mechanical Engineers, New York.

ASME. 2015. Boiler and pressure vessel code. American Society of Mechanical Engineers, New York.

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The preparation of this chapter is assigned to TC 8.4, Air-to-Refrigerant Heat Transfer Equipment.