Thermal Design of Cooling and Dehumidifying Coils

The cooling and dehumidifying coil is a critical component of air conditioning. Its performance has a strong bearing on the ultimate indoor environmental conditions, which in turn, has a significant impact on the indoor air quality. Decisions made to select a cooling coil influence the initial investment as well as the costs of installing, providing, and maintaining thermal comfort. The efficient thermal design of the cooling coil leads to a crucial reduction in the coil surface heat transfer area and of course, its capital cost and its weight. On the other hand, the enhancement in the coil thermal performance will usually be established at expense of the hydraulic performance of the cooling coil and in turn, its running cost. Because the cooling coil is an integral part of the air distribution system, its geometry — size, number of rows, fin spacing, and fin profile — contributes to the airside pressure drop and affects the sound power level of the fans. (Fan power needed to circulate air through the duct system may warrant extra sound attenuation at the air handler.) Cooling coils are an integral part of the chilled water system or the refrigeration unit, too. The extent to which coils raise the chilled water temperature or the evaporation temperature dramatically affects both capital investment in the cooling coil or the pumping power. Coil performance can even influence the efficiency of the chiller or Dx-unit. The focus of this chapter is on the description of the methodology should be used in thermal design of the cooling coil either chilled water coil or Dx-coil.


Introduction
The cooling and dehumidifying coil is a critical component of air conditioning. Its performance has a strong bearing on the ultimate indoor environmental conditions, which in turn, has a significant impact on the indoor air quality. Decisions made to select a cooling coil influence the initial investment as well as the costs of installing, providing, and maintaining thermal comfort. The efficient thermal design of the cooling coil leads to a crucial reduction in the coil surface heat transfer area and of course, its capital cost and its weight. On the other hand, the enhancement in the coil thermal performance will usually be established at expense of the hydraulic performance of the cooling coil and in turn, its running cost. Because the cooling coil is an integral part of the air distribution system, its geometry -size, number of rows, fin spacing, and fin profile -contributes to the airside pressure drop and affects the sound power level of the fans. (Fan power needed to circulate air through the duct system may warrant extra sound attenuation at the air handler.) Cooling coils are an integral part of the chilled water system or the refrigeration unit, too. The extent to which coils raise the chilled water temperature or the evaporation temperature dramatically affects both capital investment in the cooling coil or the pumping power. Coil performance can even influence the efficiency of the chiller or Dx-unit. The focus of this chapter is on the description of the methodology should be used in thermal design of the cooling coil either chilled water coil or Dx-coil.
Methods to design the cooling and dehumidifying coil either chilled water coil or Dx evaporator coil are usually based on log mean enthalpy or log equivalent dry-bulb temperature difference [1]. In both methods, the cooling coil is treated as a single zone/region and hence the required surface area is determined [2]. This manner of the cooling coil design could lead to an imprecise design particularly when the cooling coil is partially wet. In this chapter, the numerical calculation using a discrete technique "row-byrow method" will be presented to calculate the detailed design of the cooling coil in order to enhance the calculation accuracy and trace the air and coil surface temperature locally.

Types of cooling coils
Cooling coils are classified to direct-expansion (DX) coils and chilled water coils as shown in Figure 1. Some coil manufacturers fabricate coils from 5/8 inch OD copper tubes, others from 1/2 inch copper tube and still others use 3/8 inch tubes. Selection of the tube size is a matter of manufacturer's choice and market demand. Price, as always, plays a major part in the tube size selection.

Coil construction and geometry
In a coil, copper tubes are arranged parallel to one another, either in staggered pattern or non-staggered pattern, along the length L of the coil. A staggered pattern is more commonly used. For 5/8 inch tubes, the triangular pitch is 1.75 inch or 1.5 inch. For 1/2 inch tubes it is 1.25 inch. Plate or ripple fins are used to enhance the heat transfer area. Thus the primary surface area (outside area of bare copper tubes) is enhanced greatly by adding a secondary area of fins. The total area including fins is called outside surface area. The cross-section (L × H) which the air flows is called the face area or the finned area. Thus L is finned length and H is fin height (see Figure 2). Fins are arranged perpendicular to the tubes. Where, the fin spacing varies between 8 and 16 fins per inch of tube.  The average air velocity across the face area is called the coil face/frontal velocity and it is calculated as follows [3]: The number of rows of tubes in the direction of air flow is termed as depth of coil (rows deep, D). Coils with 3, 4, 6 or 8 rows are commonly used. Refrigerant or chilled water enters the first row and leaves the coil from the last row. A coil in which chilled water or refrigerant is supplied to all the tubes in the first row (also referred to as tubes high or tubes in face) is called a maximum or full circuit coil (see Figure 3). Thus a typical coil of 17.5 inch (0.44 m) height which has 10 tubes in face (based on 1.75 inch (0.044 m) pitch) will have a maximum of 10 circuits. If the supply is given to alternate tubes in face, we get a half-circuit coil with 5 circuits as against 10 circuits. The U-bends at the end of the tubes can be arranged, at the time of manufacturing, to obtain the number of circuits desired. See Figure  4 for full and half circuit coils with 4 tube face.
Face velocity is restricted to 500 fpm (2.5 m/s) to avoid carryover of condensate from the coil. The value of 500 fpm (2.5 m/s) is very commonly used for coil sizing and it works very well for cfm/ton in the range of 500 to 600 (2.5 to 3 m 3 /s per ton). If cfm/ton ratio falls below 500 (2.5 m 3 /s per ton), this generally happens when room sensible heat factor goes below 0.8 due to high room latent load, a 4-row coil at 500 fpm (2.5 m/s) becomes inadequate. A 5-row coil is not very common. Hence by lowering face velocity, a 4-row deep coil can be selected at 400 fpm (2 m/s), when cfm/ton is about 400 (2 m 3 /s per ton).. As cfm/ton ratio reduces further, 6-row or 8-row coils have to be selected. This situation is encountered when the occupancy and/or fresh air components are high.

Fin patterns
There are three standard plate fin patterns that are usually used in the cooling coil: flatplate, wavy-plate, and star-plate fin patterns, as shown in Figure 5. They are made of Aluminum, copper, and stainless steel or carbon steel. The fins are permanently attached to the tubes by expansion of each tube. Full fin collars allow for both precise fin spacing and maximum fin-to-tube contact. The flat-plate fin type has no corrugation, which results in the lowest possible air friction drop and lowest fan horsepower demands while the wavy-plate fin corrugation across the fin provides the maximum heat transfer for a given surface area, and is the standard fin configuration used. The star-plate fin pattern corrugation around the tubes provides lower air friction. This pattern is used when lower air friction is desired without a large decrease in heat transfer capacity.

Simultaneous heat and mass transfer in cooling and dehumidifying coils
In the cooling coil, the coolant fluid "chilled water or refrigerant" flows inside the tubes and the air passes across the tube bundle. Since the coolant fluid temperature is less than the dew point temperature to ensure the dehumidification process there is possibility of heat www.intechopen.com and moisture transfer between them. The directions of heat and moisture transfer depend upon the temperature and vapor pressure differences between air and wetted surface. As a result, the direction of the total heat transfer rate, which is a sum of sensible heat transfer and latent heat transfers. The concept of enthalpy potential [4] is very useful in quantifying the total heat transfer in these processes and its direction.
The sensible (Q S ) and latent (Q L ) heat transfer rates are given by: Q L = h mass A S (W i -W a ) h fg the total heat transfer Q T is given by: Since the transport mechanism that controls the convective heat transfer between air and water also controls the moisture transfer between air and water, there exists a relation between heat and mass transfer coefficients, h C and h D as discussed in an earlier chapter. It has been shown that for air-water vapor mixtures, H mass ≈h o /c pm or h o /h mass .c pm = Lewis number ≈ 1.0 Where c pm is the humid air specific heat ≈ 1.0216 kJ/kg.K. Hence the total heat transfer is given by: by manipulating the term in the parenthesis of RHS, it can be shown that: The air heat transfer coefficient, h o has been computed from the experimental correlations derived in [3]. The heat transfer parameter is written as Stanton number, St times Prandtl number, Pr to the 2/3 power. It is given as a function of Reynolds number, Re where the function was established through curve-fitting of a set of the experimental data as follow: Where these three dimensionless parameters are defined as:

Governing equations and methodology
The sizing of cooling coil requires solving the two energy equations of the air-side and coolant sides coupling with the heat and mass transfer equations. The design is accomplished through discretizing the cooling coil into N segments according to the number of the coil rows. The three governing equations are applied to each segment. By knowing the process data, coil geometry, and the design cooling load imposed on the coil the required surface area can be computed. The coil sizing is expressed by the face area and number of rows of a finned-tube coil for satisfying the design coil cooling load. Air-Side www.intechopen.com Here, Eliminate ha i+1 and Tw i+1 from Equation (1) & (3) respectively, the energy equations can be formulated; Eliminate ha between equations (2) & (7), it is yielded: Similarly, eliminate Tw between equations (4) and (8): Now, by dividing equation (9) over equation (10): Where, Relation between hs and Ts: a. Dry-Surface (Ts  Tdew point) When the coil is wet the enthalpy of saturated air hs is a function of the temperature of the wetted surface Ts , by curve fitting for psychometric chart [2] Summary of final solution: The final solutions for the coil capacity per row and for the states of air and water at the exit of any row within a chilled-water coil are given, in terms of the mean outer surface temperature of this row, as: The calculations of (Tw i+1 , ha i+1 , Ta i+1 ,and ∆Q ) are started from the first row until reaching the row number Nr at which its outlet water temperature is nearly equal to the given inlet water temperature to the coil, i.e. Tw Nr+1 Tw in .
Procedure of cooling coil design at a given cooling load Q : 1. The condition of the air leaving a chilled-water coil is nearly saturated, therefore, the relative humidity of the outlet air, ϕ out from the coil can be assumed as 95 %. 2. Knowing [inlet air state, CSHF= Q S /Q ,and ϕ out , the enthalpy of the outlet air ha out from the coil can then be determined from the Psychometric Chart. 3. Knowing [Q , ha in , and ha out ], then the air flow rate can be determined as: a. The coil dimensions (tube length, finned width and coil depth). b. The number of coil rows and the total number of tubes. c. The exit air temperature.

Calculation Procedures
From psychometric chart at inlet air conditions the inlet air properties are obtained represented by ha in =54 kJ/kg, Wa in =0.011 kg v /kg a . and dew point temperature, dpt = 15.5 o C. By knowing Q C =60 kW, CSHF=0.75 (=1-Q L /Q c ), and ϕ = % using information from inlet point, the exit conditions can be determined as ha out =33 kJ/kg, Ta o =10. . .
The total calculated cooling load for 6-rows coil is: Q =64.31 kW And coil sensible heat factor, = = .
The calculated unknowns are listed row-by-row in the next Table; and the psychometric process for the cooling and dehumidification process is represented by Figure 6.  = ℎ= * = .

Row number Surface condition Ts mi o C ∆Q
c. Exit air temperature Ta out =11.14 o C

Design of the cooling coil as single Region
In calculating the surface area of the cooling coil, the heat and mass transfer equations are applied on the entire coil surface. This approximation will greatly simplify the analysis. The obtained results ( A o , Ta out ) for one-section coil will be compared with the corresponding results obtained for N r -sections coil. Water-side Applying the heat transfer equations for the air and water at the inlet and exit sections of the coil, this leads to the following equation for Ts at these sections: For an entire wet-surface, the saturated air temperature at the inlet and exit of the coil surfaces Ts 1 and Ts 2 are obtained, in a similar manner as done before for N-sections coil, as: . . * * . .
Where, Tw 1 = inlet water temperature Tw 2 = exit water temperature = (8) Knowing (Ts 1 & Tw 1 ) and (Ts 2 & Tw 2 ), the mean temperature difference between the chilled water and the coil surface can be assumed equal to the logarithmic mean temperature difference. ∆ can be determined from: The area of the coil can now be determined from equation (4) as: The outer coil surface area A o is determined from The volume of the cooling coil is given as:

Calculation of exit air Temperature:
The temperature difference between the air stream and the coil surface is approximated as arithmetic mean temperature difference as shown from the heat transfer equation for the dry air.

Worked Example
We will solve the previous worked problem using principal of treating the coil as single zone/section instead of multi-sections and compare the two results.   The results presented in Table-1 indicate that cooling coil analyzed as only one-section gives results with good agreement with those obtained with the coil analyzed as 6-sections. The maximum error is 7%.

Worked example of partially dry chilled-water coils
Cross-counter flow chilled water cooling coil using corrugated plate-fins, has the flowing construction and operating design parameters: a. The coil dimensions (tube length, finned width and coil depth). b. The number of coil rows and the total number of tubes. c. The exit air temperature.

Calculation Procedures
From psychometric chart at inlet air conditions the inlet air properties are obtained represented by ha in =48 kJ/kg, Wa in =0.0081 kg v /kg a, dew point temperature, dpt = 10 o C. By knowing Q C =60 kW, CSHF=0.75 (=1-Q L /Q c ), and ϕ = % using information from inlet point, the exit conditions can be determined as ha out =30.6 kJ/kg, Ta o =10. Since the mean coil surface temperature at the 1 st row is 13.5 and it is larger than the inlet dew point temperature of the entering air, dpt = 10 o C the coil will be partially dry until the coil surface temperature reaches at least the dew point temperature. Therefore, the dry coil equations will be used here. Tsm < dpt therefore, the coil will act as a wet coil ℎ = 1.001xTa 5 +Wa 5 *(2501+1.8* Ta  = .
The calculated unknowns are listed row-by-row in the next Table; and the psychometric process for the cooling and dehumidification process is represented by Figure 7.  = ℎ= * = .
c. Exit air temperature Treating the cooling coil as a single zone "Worked Example" We will solve the previous worked problem using principal treating the coil as single zone/section instead of multi-sections and compare the two results.  = .98 The results presented in Table-2 indicate that cooling coil analyzed as only one-section gives results with good agreement with those obtained with the coil analyzed as 6-sections. The maximum error is 12%.

Worked problem on the thermal design of Dx-coils
Cross-counter flow Dx-evaporator coil using corrugated plate-fins, has the flowing construction and operating design parameters:  Air is saturate at this temperature with ha 2 = 31.5 kJ/kg

Conclusion
In this chapter, simulation of the cooling coil using a discrete technique "row-by-row method" has been presented. The main advantage of this method is to trace the air and coil surface temperature locally. In addition, this method gives more accurate results for the cooling coil design or simulation compared with those given by ordinary method such as log mean enthalpy method.
Step-by-step procedure has been introduced and worked examples are presented. The deviation between the two methods "numerical discrete method and treating the coil as a single zone" is around of 12%. Selecting and bringing together matter provided by specialists, this project offers comprehensive information on particular cases of heat exchangers. The selection was guided by actual and future demands of applied research and industry, mainly focusing on the efficient use and conversion energy in changing environment. Beside the questions of thermodynamic basics, the book addresses several important issues, such as conceptions, design, operations, fouling and cleaning of heat exchangers. It includes also storage of thermal energy and geothermal energy use, directly or by application of heat pumps. The contributions are thematically grouped in sections and the content of each section is introduced by summarising the main objectives of the encompassed chapters. The book is not necessarily intended to be an elementary source of the knowledge in the area it covers, but rather a mentor while pursuing detailed solutions of specific technical problems which face engineers and technicians engaged in research and development in the fields of heat transfer and heat exchangers.