An experimental study of the performance of a direct expansion evaporating coil

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A Thesis Presented to the Faculty of the College of Engineering The University of Southern California

In Partial Fulfillment of the Requirements for the Degree Master of Science in Mechanical Engineering

by Charles M. Daniels, Jr. August 1950

UMI Number: EP60503

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thesis, w r it t e n by

u n d e r the g u id a n c e o f h i®... F a c u l t y C o m m itte e , and ap p ro ve d

by a l l

its m em bers, has


presented to a n d accepted by the C o u n c i l on G r a d u a te S tu d y a n d R e s ea rch in p a r t i a l f u l f i l l ­ m ent o f the re q u ire m e n ts f o r the degree o f



Master of Science in Mechanical Engineering

n„„ August 1950

Faculty Committee



PAGE INTRODUCTION ...............................


The problem and object of s t u d y ..........


.II. DEFINITIONS OF TERMS U S E D ..................


Self-contained air conditioner ..........


Condensing units .......................


Absorption refrigeration unit Cooling coils



6 6

Log. mean temperature difference........


Mechanical cooling .....................


H e a t ...................................


- Temperature



Pressure ...............................


Boiling point



Ton of refrigeration....................


Coefficient of performance ..............


Apparent and actual mean coefficient of heat t r a n s f e r ....................... III.


REVIEW OF THELITERATURE................ . . .


Refrigerants .............................


Refrigeration cycle



Fundamental mechanisms of heat transfer Performance of air conditioning coils

. . ...

13 15




HEAT TRANSFER CALCULATION FOR EXTENDED SURFACES................................. Heat transfer of finned tubing ..........


Surface coefficients ...................


Coils operating with dehumidification


. .

Metal resistance....................... Empirical relations and formulas ..........


29 30

Fin effectiveness.................. . . .





Surface coefficients ...................




A p p a r a t u s ...............................


Self-contained air conditioner ..........


Anemometer or air-velocity meter ........





Thermocouple instrument Heaters







Experimental procedure ...................


The s e t - u p .............................


Operating conditions ...................


Test r u n s .............................





Data o b t a i n e d ...........................


Calculations................... ..



PAGE Determination of surface coefficients

. .


Determination of apparent and actual mean coefficients .........................


Relationship of surface coefficients to over-all expression Evaporator capacity Sample calculations







Determination of the over-all coefficient


of heat transfer.....................


SUMMARY AMD C O N C L U S I O N S ...................


BI B L I O G R A P H Y .......................


A P P E N D I X .........................................




I. The Effect of the Number of Rows of Coi.l Depth and Fin Spacing on the Amount of


Heat Transferred.........................


General Information on the Model 3~SCDUnit. .


Consolidated Tabulated Test Results



Observed D a t a ..............................




Fin Efficiency as a Function of Coil Geometry and Material .....................


Psychrometric Chart Showing Method of Predicting Coil Performance ................




Overall Dimensions Chrysler Airtemp Model 3 - S C D ...................


Typical Fin


Potentiometer with Cold-Junction Compensator. .



Sketch of Overall Test s e t - u p ................



Type of Heater Used to Load Conditioner . . . .



Thermocouple Locations on C o i l .............



Thermocouple Installation



Model 3 - S C D ...................



Graphical Method of Predicting Actual Mean Coefficient of Heat T r a n s f e r ..............




Relation of Overall Expression to Various C o m p o n e n t s .....................



Evaporator Capacity .........................



General View of the Test S e t - u p .............



A View of the Coil Showing Thermocouple Leads .



General View With Heaters and Anemometer in Test Position.............................




The purpose of any cooling system is to remove heat. This heat can be removed by means of a mechanical system, a medium or refrigerant, a function and a method or cycle of operation.

Since heat flows from a warmer to colder

substance, the transfer of heat forms the basis of the science of refrigeration. The Chrysler Airtemp Conditioner, model 3“SCD, is a self-contained air cooling unit which represents a mech­ anical heat removal system through the compression and ex­ pansion of a refrigerant.

This system consists of: (l) a

compressor, (2) combination condenser and receiver, (3) an evaporator with an air-side blower unit, and (4) a means of controlling the flow of the refrigerant.

This unit em­

ploys a direct air cooling system with the fluid or re­ frigerant inside the coil and the air passing normal to it. The extended surface coils consist of a primary surface and a secondary surface.

The external surface of the tubes

is known as primary and that of the fin as secondary.


primary surface consists of round tubes arranged staggered with respect to the air flow.

The secondary surface


consists of flat fins enclosing several rows of tubes. Cooling and dehumidifying coils are usually rated 1 within the following limits: Entering Air Dry-Bulb-------- 60 to 100 F. Entering Air Wet-Bulb-------- 50 to 80 F. Air Face Velocities----- ---- 300 to 800 fpm Volatile Refrigerant Temperatures--------------- 25 to 55 F, at coil suction outlet Water Temperatures----------- 40 to 65 F. Water Quantities------------- 2 to 6 gpm per ton. Water Velocity--------------- 2 to 6 fps. The ratio of total to sensible heat removed varies in practice from 1.00 to about 1 .65, i.e., sensible heat is from 60 to 100 per cent of total, depending on the applica* tion.


On usual comfort installations air face velocities

between 400 and 600 fpm are frequent, 500 being a common 3 value. Refrigerant temperatures ordinarily vary between 4-0° and 50°F where cooling is accompanied by dehumidification.

Water velocities range from 2 to about 6 fps. The operating conditions for a specific evaporator

coil are established by the designer, and are the results of weighing the contributions of all the components of the 1

Heating, Ventilating, and Air Conditioning Guide, 1949, V o l . 27 (New York: American Society of Heating and Ventilating Engineers, 19^9), P« 541• ^ Hoc. cit. 3 Loc. cit. 4- Loc. cit.

3 of the system in order to obtain a well balanced design.


varying such physical features as tube size, fin spacing, refrigerant, the depth of coil rows, and air volume over the coils, the designer is able to produce a wide range of performance.

For whatever conditions the designer selects,

a coil can be designed which will give the required per­ formance with an economical use of materials. The performance of a cooling unit depends in general upon: 1. The over-all coefficient of heat transfer from the fluid within the coil to the air passing over the coils. 2. The mean temperature difference between the fluid within the coil and the air flowing over the coil. 3. The physical dimensions of the coil such as face area, depth of coil rows, fin spacing, tube size, inside surface area, outside surface area and the gross area. Of all the factors affecting the performance of coils, the over-all coefficient of heat transfer is the most difficult to determine and data have been procured so far by experiment only.

In spite of the many test results

that have been published, no generally accepted value is given for a specific type of coil. This study will determine the over-all coefficient of heat transfer for the evaporator coil of the Chrysler Airtemp Conditioner, model 3-SCD.

It should be remembered

4 that the over-all coefficient of heat transfer U depends upon many variable factors and its value for various con­ ditions of operation cannot be predicted from the results of a test at one set of conditions.

In order, then, that

its value may be computed for various conditions, knowledge of inside surface coefficients hj_ and outside surface co­ efficients hQ, and the ratio of the external to the inter­ nal surface, R, is essential.

The plan of this study is to

determine the mean coefficient of heat transfer for the out­ side surface, the internal film coefficient by the use of equation (13), and the degree of fin efficiency with the use of equation (10).

Substitution of these values of hQ, hi

and R in equation (l) establishes the value of U. The data and results are based on the findings of the author in the gage laboratory of the Department of


anical Engineering of the University of Southern California where the self-contained air conditioner was installed.


complete outline of the procedure followed in the laboratory is given in Chapter V.

CHAPTER II DEFINITIONS OF TERMS USED Self-contained air conditioner.

According to the Air 1 Conditioning and Refrigerating Machinery Association a self-contained air conditioner is a factory made assembly of a condensing unit, or absorption refrigeration unit, a means for air circulation, ventilation, air cleaning, cool­ ing, dehumidification and control of temperature and may include or have provision for heating and dehumidifying. Condensing units.

These parts of a refrigerating

system are of great importance as it is within them that the heat acquired by the refrigerant in the system is disposed of at a comparatively high temperature level. In general, condensers may be divided into two groups, those cooled by air and those cooled by the action of water.

Water-cooled condensers are built as shell and

tube units with one or more water passes, or double tube condensers or tubular units upon which the cooling water is sprayed.

Double-tube condensers have inner tubes through

which water is circulated and outer tubes surrounding the inner tubes, refrigerant being confined in the spaces 1

Journal of American Society of Refrigerating En­ gineers, Refrigerating Engineers Magazine, Vol. 57. April,

19^ 9 -


between them.

Shell and tube types of condensers are those

in which the refrigerant to be cooled and liquefied is confined within the shell surrounding the tubes through which water is circulated. Absorption refrigeration unit. This method of re­ frigeration does not require a compressor for putting the refrigerant under pressure.

A source of heat (steam or hot

water) is necessary for production of the initial pressure of the refrigerant. Cooling coils.

According to A.S. H.V.E. Guide,



cooling and dehumidifying coils used in unit air condition­ ers are essentially the same as those used in central sta­ tions units.

The face area of the coil is essentially

fixed and the number of rows deep in the direction of air flow is the variable (variable prior to installation i.e., not variable in operation of unit) that determines capacity. Log, mean temperature difference. The temperature difference used in calculations must be a representative average temperature difference between the air stream and that of the surface with which it is in contact. 2

A straight

Heating. Ventilating, and Air Conditioning Guide, 19^9, Vol. 27 (New York: Society of Heating and Ventilating ‘ Engineers, 19^9), p. 699.

7 arithmetic average is satisfactory within + 5$


when the

entering temperature difference does not exceed twice the leaving temperature difference.

It is much more satisfac­

tory to use what Is termed the logarithmic mean temperature difference which is defined as:

■n ’ *

Initial diff. - Final diff. log^ (initial diff.) (final diff.)

Mechanical cooling.

A process, by means of mech­

anical equipment, of removing heat from a given space is called mechanical cooling. Heat.

A form of energy which cannot be destroyed

nor lost, but is transferable from one substance to another. Heat always flows from a warmer to colder substance. Sensible - heat transferred which can be measured by a change in temperature. Latent - A term used to express the energy involved in a change of state. Heat Quantity -Amount of heat required to raise the temperature of one pound of water 1 degree fahrenheit, or the amount of heat removed in cooling one pound of water 1 degree fahrenheit. Unit of heat is known as the British Thermal Unit expressed as B.T.U. 3 William H. McAdams, Heat Transmission (New York and London: McGraw-Hill Book Company, Inc., 1942), p. ^59.


Temperature. Degree of intensity of heat.

It may

be measured on either the fahrenheit or centigrade scale. Dry Bulb- May be measured by an ordinary thermometer and indicates sensible heat changes. Wet Bulb - May be measured by a thermometer equipped with a wick which is saturated with water. It indicates latent heat changes. Relative Humidity - Ratio of water content in air to the water content of saturated air at the same wet bulb temperature. Pressure. Force per unit area exerted by a liquid or gas. Gage pressure - Pressure above atmospheric pressure. Suction or back pressure - Pressure in the line of suction where it leaves the evaporator. Head pressure - Discharge pressure from the com­ pressor. Oil pressure - Pressure existing on the discharge side of a pressure pump used to lubricate moving parts of the compressor. Boiling point. Temperature at which evaporation takes place for any given pressure.

For example, water

boils at 212 deg. F. at atmospheric pressure and the boil­ ing point of Freon 12 is 40 deg. F. at 37 lbs. per sq. in. gage pressure. Ton of refrigeration.

Cooling effect produced by

mechanical refrigeration that is equivalent to the melting of one ton of ice from and at 32°F in twenty-four hours.

9 One ton of refrigeration is also expressed as cooling at the rate of 12,000 Btu per hour.

It is the measurement of

refrigerating equipment capacity. Coefficient of performance. The coefficient of performance of a refrigerating machine is analagous to the efficiency of a heat engine.

It is defined as the ratio of

the heat received hy the machine from the body that is being cooled to the net work received. Apparent and actual mean coefficient of heat trans­ fer.

Throughout this study h is called the apparent co­

efficient of heat transfer, based on the entire outside area, as opposed to he which represents the actual mean value over the entire surface comparable in magnitude to that prevailing at the prime surface.

Joseph H. Keenan, Thermodynamics (seventh print­ ing; New York: John Wiley and Sons, Inc., London: Chapman and Hall Limited, 1948), p. 245.

CHAPTER III REVIEW OF THE LITERATURE I . REFRIGERANTS Quite a number of media can be used as refrigerants, but those most practical for air conditioning are relative­ ly few in number.

Possible refrigerants are air; carbon

dioxide; ammonia; water vapor; dichlorodifluoromethane, commonly known as Freon, F-12; methylene chloride, which is Carrene No. 1; monofluorotriehioromethane, commonly called F-ll and Carrene No. 2; methyl chloride; and Zeon which is dichloromonofluoromethane.^

Air has sometimes been used

because it is free, abundant, and nonpoisonous; but

the co­

efficient of performance of air cycles Is not favorable. Constant pressure processes are usually substituted for the more impracticable isothermal processes of the Carnot cycle. All refrigerants have comparatively low boiling points have the ability to absorb heat.


The refrigerant thus

becomes the agent or means by which heat is transferred from

1 William H. Severns, Heating,Ventilating, and Air Conditioning Fundamentals (eleventh printing; New York: John Wiley and Sons. Inc., London: Chapman and Hall, Limited, 1949), p. 428. ^ Joseph H. Keenan, Thermodynamics (seventh printing; New York: John Wiley and Sons, Inc., London: Chapman and Hall, Limited, 1948), p. 246.



one part of the mechanical system to another.

A refriger­

ant can rapidly change its physical state, that is, absorb heat in changing from a liquid to a gas, and give up heat in changing from a gas to a liquid.

An ideal vapor for use

in a refrigerating machine would have among other character­ istics a moderate pressure at each limit of temperature in order that: (l) the volume of the desired amount of fluid shall not be excessively large and hence require large and expensive equipment, and (2) the stresses in the restrain3 ing walls shall be moderate. Water, in view of these requirements, is barely acceptable over a narrow range of temperatures, say between 35

where the density of the

vapor is only 0.00034 lb/cu ft., and atmospheric temperature.^ In this range it may be used, at least in small installa­ tions, if some cheap substitute can be found for the more expensive parts of the machine. Pressure has an effect on the boiling point of the liquid.

An increase in pressure causes a rise in the boil­

ing point of the refrigerant and the opposite occurs for a decrease in pressure. *2 J Keenan, loc. cit. 4 Loc. cit. ^ "Chrysler Airtemp 3 H.P. Packaged Air Conditioning Unit," Service Manual (Dayton, Ohio: Chrysler Corporation, Airtemp Division).




A compressor withdraws vapor from the evaporator, compresses the vapor and discharges it into the condenser. The condenser and evaporator function as heat exchangers through which refrigerant is passed to effect change in its physical state.

The amount of refrigerant necessary to

perform this operation is regulated by the expansion valve. When the fan and compressor switch and the tempera­ ture switch on the control panel of a 3 hp Chrysler Air Conditioning unit, model 3 SCD, are turned to their opera­ ting positions, the unit circuit is energized.


refrigerant, contained in the condenser under pressure, flows to the expansion valve.

Here the pressure is reduced

and the liquid fed into the evaporator or cooling coil.


liquid then absorbs heat from the.air circulated by the blower unit over the outside surfaces of the coil.


liquid refrigerant absorbs heat because it is held under a pressure at which its boiling point is lower than the tem­ perature of the circulated air.

The absorption of the heat

by the refrigerant has two results: (l) it lowers the tem­ perature and moisture content of the air circulated by the blower unit over the evaporator unit and changes the refrigerant from a liquid to a vapor.



c i t ., p. 433.

13 This vapor is withdrawn from the evaporator by the compressor and compressed to a pressure at which the re­ frigerant temperature is higher than that of the cooling medium employed in the condenser.

The vapor is discharged

from the compressor into the oil separator and then into the condenser where a cooling medium, which is water, is circulated through the spiral condenser coils.

The dis­

charge vapor, upon contacting the colder water coil, gives up heat and liquefies. The cycle is continuously repeated until the area served by this unit reaches the desired conditions.



Heat is transferred from points of high to points of lower temperature either by passing directly from one par­ ticle of matter to another, by bulk mixing of matter, or by the passage of energy waves directly from one point to another. The first two phenomena require the presence of matter in the space intervening between the points of differ­ ing temperature and are called conduction and convection,

^ Heat Exchanger Tube Manual (second edition; pub­ lished by the Scovill Manufacturing Company), p. 75*

Ir­ respectively.

The third manner in which heat is transferred

does not require the presence of matter between the points of temperature difference, it is spoken of as radiation. If one end of a eoppei* rod is heated and the other end is shielded from the passage of heat except through the rod itself, it is noticed that the shielded end also in­ creases in temperature because heat has been conducted through the rod from the higher to the lower temperature end. When a fluid such as air is heated, the warmer por­ tion because of its lower weight per unit volume, rises and carries heat to the colder portion.

Thus heat is trans­

ferred by convection, that is the bulk motion of matter from higher to lower temperature levels. Heat is transferred from an electrically heated filament placed in an evacuated steel cylinder even though there is no matter between the filament and the wall of the cylinder.

This transfer of heat without the aid of matter

is termed radiation. In nearly all practical cases the over-all mechanism of heat transfer is a combination of conduction, convection, and radiation.

For example, the heat from the filament in

the evacuated cylinder mentioned above is transferred to the wall by radiation; it flows through the steel wall by conduction and warms the air surrounding the outside of the

15 cylinder.

The warmer air rises because of its decreased

density and transfers the heat away from the steel wall by convection.

If the air moves only because of differences

in density, the motion is described as natural convection. The air might also be moved relative to the outside wall of the cylinder by a fan or other mechanical means.

In this

case the motion is called forced convection. Heat transfer is studied by breaking down the over­ all mechanism into the fundamental phenomena of conduction, convection and radiation; for each of these primary mech­ anisms there are fundamental quantitative relationships relating the rate of heat transfer to the temperature dif­ ferential that acts as the driving force for the flow of heat, and to the resistance to heat flow offered by each part of the system.



According to D. D. Wile


the performance of an air

conditioning coil depends upon many factors other than the entering wet and dry bulb temperatures.

For instance, a

shift of only the dry bulb temperature affects the o

D. D. Wile, "Air Conditioning Coils — Their Heat Transfer Problems," Refrigerating Engineering, 56:320-23, October, 19^8.

16 performance of a coil by changing the ratio of sensible heat to latent heat. In the same article the effect of the following factors on the performance of air conditioning coil has been shown: (l) the decrease in evaporating temperature on a six row coil operating at 500 fpm face velocity causes an in­ crease in rating and a decrease in sensible heat ratio, (2) increase in coil depth when operating at 40 deg. F. evaporating temperature and 500 fpm face velocity causes an increase in capacity and with it a decrease in sensible heat ratio, and (3) increase in air velocity in a six row coil operating at 40 deg. F. evaporating temperature causes an increase in rating.

It is for this reason that it some­

times becomes necessary to operate with low air velocity through the coil in order to obtain a sufficiently low sensible heat ratio. As a rule of thumb the following operating conditions and requirements have been proposed: 1. Coil face velocity 500 fpm 2. Airquantity----------- 300 to400 cfm per ton 3. Evaporating temperature- 40 to 45 deg. F. 4. Inside design conditions: humid climate-80 F. DB 50$ R.H. hot dry climate — 80 F. DB -- 40$R.H. cool dry climate - 78 F. DB--- 45$R. H.

9 Wile, loc. cit.

17 5

Coll Depth 4 6 6 8


row row row row


Latent Heat

dry dry humid humid

low high low high

T. C. Gleason^ in his article showed in a tabulated form (Tables la, lb) the effect of the number of rows of coil depth and fin spacing in the amount of heat trans­ ferred. The operating conditions at which the following tables were established were: Preon 12, refrigerating pres­ sure leaving coil— 37 lbs. per sq. in. gage, 40 deg.F. at saturation; refrigerant temperature leaving coil--42 deg.F. to 50 deg.F. (2 to 10 deg. superheat); air temperature entering coil--80 deg.F. D.B. to 67 deg.F. D.B.; air velo­ city entering coil— 800 fpm, and standard air density 0.075 lb./cu. ft.

no T. C. Gleason, "Refinement of Evaporator Coils for Package Units," Refrigerating Engineering, 55:549-51* June, 1948.



Wall thick


Spac­ ing



in line





in line




in line tube



Rows deep

Fins / in.

Sen­ sible













142 117 1 lb .cu/ton of re frigerat ion**


* 100$ indicates initial condition and subsequent percent­ ages indicate increase over initial condition. ** Weight of copper coil per ton of refrigeration. TABLE lb


Wall thick

Arrange­ ment

Spac­ ing

Rows deep

Fins / in.

Sen­ sible




0.375 0.022







0.375 0.022







0.375 0.022










lb .cu/ton of refri­ geration**

* 100$ indicates initial condition and subsequent percent­ ages indicate increase over initial condition. ** Weight of copper coil per ton of refrigeration.

CHAPTER IV HEAT TRANSFER CALCULATION FOR EXTENDED SURFACES The study of the heat transfer of finned surfaces and pipes will serve to illustrate the application of the principle of heat transfer to a specialized field. Fan coils of the fin-tube type are commonly employed in fan systems designed for supplying heated, cooled or dehumidified air as desired in various applications of airconditioning.

Fin-tube coils, as the name implies, are

generally made of tubes equipped with fins.

Generally the

tubes are made of copper or aluminum, and the fins are either copper, brass, or aluminum.

The fins are usually

rectangular or circular in shape or in the form of helical ribbon, either smooth or crimped, wound on the tube. Although there are many variations of design, the tubes are usually of small diameter, of the order of 3/8 to 3/4 inch.

The fins are generally spaced from four to eight

per inch of tube length.

If the assembled unit is dipped

in hot tin or zinc, as it is in many instances, good contact for the conduction of heat from the tubes to the fins is assured. Because of the marked difference between the extent and shape of the exterior and interior surfaces of the fin tube coils and that of the bare tube coils, the calculation


of heat transfer through the fin-tube coils differ from the calculation of heat transfer through bare tube coils. Furthermore, where a coil is used for cooling to a tempera­ ture below the dew point of the incoming air, the rate of heat transfer is affected by the film of the condensate that covers a part or all of the exterior surfaces. The exterior surface

of the tube is commonly referred

to as primary surface and the fin surface is called the secondary surface.

In most cases the extent of the fins is

such as to make the ratio of the outside surface to the in­ side surface in the range of from 10:1 to 30:1 * The tubes in any unit are of such number, length, and spacing as to provide the necessary area of opening for the desired quantity of air to flow through at a suitable velo­ city and pressure drop.

The average velocity of the air

flowing over the surface of the coil is commonly expressed as the average velocity of the entire face area, that is, the volume of air per unit of time divided by the area of the face of the coil.

In the studies of heat transfer,

however, it is often desirable to express the velocity in terms of net or free area opening between the fins and tubes, that is, the entire area minus the projected area of the fins and tubes, and quantity of air per unit time. In public buildings and other places where the noise that accompanies high velocity of air is objectionable,


face velodity for coils used are usually not more than 500 ft. per min., whereas for industrial installations, face velocities as high as 800 to 1000 ft. per min. are common.^ Units are made with one or more rows of tubes in depth in order to provide the desired degree of heating or cooling, the limiting factor being the increase in fric­ tional resistance to air flow that accompanies the in­ crease in depth. The basis for the use of fins on the outside of heat transfer tubing is apparent.

In the passage of heat through

the tube from the liquid or vapor inside to the surrounding air, the greatest thermal resistance encountered is on the outer, or air side.

It is logical to expect, then, that

this resistance can be decreased by increasing the exterior surface as by the addition of fins, provided other factors affecting the surface resistance are not generally altered.2 It is to be understood, however, that by decreasing the resistance is meant that the area is increased such that with the given resistance, more heat may be transferred.

Aubrey I. Brown, Salvatore M. Marco, Introduction to Heat Transfer (first edition; New York: McGraw-Hill Book Company, inc., 1942), p. 174. 2 Ibid., p. 175 .


Heat transfer of finned tubing;.3 The major factors affecting the heat transfer of finned tubing are as fol­ lows: 1. The mean velocity of air stream 2.

Thedegree of scrubbing action of air stream as influenced by the design of eddy currents or local turbulence


Thecondition of the exterior surface* wet or dry


Thedepth* thickness* and spacing of fins and the method of attaching them to the tube surface

3. The nature and velocity of the fluid inside the tube 6. Other factors of lesser influence are the dia­ meter of the tubes; the direction of heat flow; that is* for heating or for cooling; and the absolute temperature of the air Pin tube units are so compact that the enveloping or radiating surface per unit of heat capacity is small.


comparison with the flow of heat by convection* radiation can be of little consequence in any study of the perfor­ mance of finned tubes* and it is usually neglected.


wise* the resistance to heat flow through the thin metal tube itself is so small in comparison with the surface

3 Brown and Marco* loc. cit.

23 resistance that it may well be omitted from consideration and the problem may be considered to deal with the heat flow by convection only.

The over-all coefficient U may

then be expressed in terms of the inside and outside sur4 face conductances:

u =

(r A



(i A




where = inside (or refrigerant-side) surface coefficient hQ = outside (or air side) surface coefficient R

=ratio of the exterior to the interior surface


=over-all coefficient of heat transfer BTU per square foot of exterior or air-side surface per degree fahrenheit difference in temperatures of the fluid inside the•tube and of the air per hour The over-all coefficient U can be readily determined

experimentally for any certain set of conditions by measure­ ment of all remaining values expressed in the usual simple equation: Q = A U (tx - t2 ) where Q,

= rate of heat transfer, Btu per hr.


= area of exterior or air-side surface, sq. ft.

Ibid., p. 176.

24 (”^1 “ ^2 ) =



mean temperature difference (usually logarithmic between the fluid inside the tube and the air outside

over-all coefficient as described above

The total heat transferred may be determined as the product of the pounds of heating or cooling medium, which is condensing or evaporating within the tubes or flowing through the tubes per unit of time, and the amount of heat removed or added to each pound of fluid.

Otherwise, the

measurement may be made on the airside, and the total heat determined as the weight of air passed per unit of time multiplied by its specific heat and by its change in tem­ perature.

In the case where air is cooled and the conden­

sation of water vapor occurs, the heat removed from the air is the product of air per unit of time and the difference in enthalpy per pound at the initial and final wet-bulb temperatures. Surface coefficients.

It should be remembered that

the over-all coefficient U depends upon the variables al­ ready ennumerated and its value for various conditions of operation cannot be predicted from the result of the test at one set of conditions.

In order, then, that its value

may be computed for various conditions, knowledge of the surface coefficients h^ and hQ, and the ratio of external to the internal surface R, in equation (l) is essential.

25 The surface coefficients h^ and hQ are not easily determined experimentally, since their determinations in­ volve measurement of the temperature gradient, and that is difficult to accomplish.

The difficulty arises: (l) because

of the differences in temperature between the external sur­ face of the tubes and the various portions of the fins, and (2) because the inside-film coefficient is comparatively great and the drop in temperature across this film is small. Values of the over-all coefficient U, determined by test, however, can be broken down into the values of h^ and hQ and R, by test at several refrigerant velocities and several air velocities or by computations of hj_ and by measurement of the ratio R.

Substitution of these values of U, hi, and

R in equation (l) establishes the value of hQ . The logic of the latter procedure lies in the fact that fewer var­ iables affect

the inside film conductance than the outside

film conductance, and their influence is, in most cases, well established.

Further, since the thermal resistance

of the inside film is commonly much less than that of the outside film, the effect of inaccuracies in the computation of hi upon the outside coefficient hQ is minimized. Coils operating with dehumidification.

When coils

are used to cool air to temperatures below the dew point, the external surfaces of the coil are covered, partially


or completely* with a film of condensate and consequently a number of additional factors are involved in a study of the heat transfer. The designer of the coil to be used for cooling and dehumidification is concerned not only with the total amount of heat to be removed from the air but also with the sepa­ rate amounts of the sensible heat removed from the dry air and the latent heat removed from the water vapor. Knowledge of the proportionate amounts of sensible and latent heat is essential* since for certain applications the main function of the coil is to remove moisture from the air* and in other cases the removal of moisture is incidental to the lowering of air temperature.

These vary­

ing requirements govern the design of the coil and the condition under which it must be operated* particularly with regard to the temperature that must be maintained in the cooling medium. Many methods have been devised for predicting the performance of a dehumidifying coil.

These methods in

some cases are largely empirical or may be based upon heat transfer theory to a limited extent.

No attempt is made

here to describe the various methods in detail.

For such

a description the reader is referred to the paper by

27 A. P. Colburn and 0. A. H o u g e n ^ and the method devised by /r C. M. Ashley. In some oases calculations are based upon the over-all coefficient U for dry coils in accordance with the temperature differences or according to the ratio of the latent heat to the sensible heat


Several methods which appear more rational, have applied heat transfer principles to the extent of making use of surface coefficients or of basing calculations upon the average temperature of the external surface of the coil. In the methods that involve the average temperature of the external surface, that temperature is rarely found by test measurements] it is commonly a calculated value determined from other test data, or it is a fictitious value found graphically.

The graph is drawn on a non-logarithmic psy­

chometric chart and the surface temperature determined as the point of Intersection of the saturation line and a line through the point representing the condition of the air that

5 A. P. Colburn and 0. A. Hougen, "Design of Cooler Condensers for Mixtures of Vapors with Noncondensing Gases, 11 Ind. Eng. Chem., 26:1178-1182 (1934). ^ C. M. Ashley, "A Method of Analyzing Pinned Coil Heat Transfer Performance," Refrigerating Engineering, 51:529-32, June, 1946. 7 Brown and Marco, op. cit., p. 186.


enters the coll extending at an angle that corresponds to the ratio of the sensible to the total heat load.

Q G. L. Tuve

has developed a method by which the

final condition of the air as it leaves the coil can be predicted.

It is this method which will be applie.d here in

the further study of the performance of coils used for dehumidification. In Tuve’s method the outside-surface temperature is calculated and applied as if the outside surface were en­ tirely prime surface (i.e., without fins).

Surface coeffi­

cients are obtained from over-all coefficients in the same manner as in a dry coil. A feature of Tuve1s method, which is found in no other, is the manner of finding the relative humidity of the air leaving the coil when dehumidification is taking place.

Tuve has shown by mathematical analysis and by test

results that the relative humidity of the air leaving when it is wet is substantially the same as the relative humi­ dity of the air leaving the same coil when it is dry but with its surface at the same temperature as the dew point of the entering air.

® G. L. Tuve and L. J. Seigel, "Performance of Surface Coil Dehumidifiers," Transactions A.S.H.V.E., 44:523, 1938.

29 Values of he can be calculated from measured distri­ bution of temperatures but a more practical method is to measure the total amount of heat transfer over the entire area and the temperature difference between the prime sur­ face and the air.

For both finned plane surfaces and

finned tubing, the amount of total heat transferred by the entire area surface A is considered composed of heat transferred through the prime surface Ap and through the extended surface Ae . Metal resistance.

The fin resistance to heat trans­

fer is usually expressed in terms of fin effectiveness fac­ tor or degree of fin efficiency.

Ordinarily it is possible

to neglect the metal resistance through the tubes and con­ sider only the resistance of the fin where the fin effec­ tiveness is reasonable.

For determination of the degree of

fin efficiency, it is best to use as a starting point the differential equation of the temperature gradient in a fin as introduced by Harper and Brown.^

A detailed discussion

of this theory will foe given in a subsequent section of this chapter.

9 Harper and Brown, .1Mathematical Equations for Heat Conduction in the Fins of Air-Cooled Engines,11 Technical Report No. 158, National Advisory Committee for Aeronautics, Annual Report 1921-1922, pp. 680-684.

30 I.


Fin effectiveness>

For determination of the degree

of fin efficiency, it is best to use as a starting point the differential equation:10 d20

2 q0



Nomenclature. f

« fin effectiveness


= coefficient of heat transfer from surface


= true width of fin


= fin thickness


= fin thermal conductivity


= temperature of fin above air


astemperature of cylinder wall above air


= heat dissipated by fin per unit of time


- heat dissipated by equal area of wall surface per unit of time

w f as corrected fin width « w + t/2 dx

]~ ] w

10 Harper and Brown, loc. cit

31 General assumptions: four general assumptions of a physical nature are made that apply to every case to be considered. 1.

Quantity of heat transferred per unit of time from the metal surface to the air is proportional to the temperature difference between the metal and air.


The coefficient q, heat transferred in unit time from unit surface per unit temperature differ­ ence is constant over the fin surface.


The fin is symmetrical about a plane through its middle and approximately to its faces.


The temperature at a given point is independent of the time.

Simplifying assumptions: 1.

The temperature across the fin thickness is sensibly constant.


The fin is so long that the effect of the ex­ posed ends (of longitudinal fins) is negligible.


The heat loss from the exposed edge can be accounted for by imagining the width extended by a distance equal to 1/2 the fin thickness at the outer edge and assuming no heat loss from the end.


The base temperature is constant.

32 5.

The fin effectiveness is constant.


The fin is longitudinal.

Boundary conditions: dO / dx




when x = w


The difference in quantity of heat conducted into an elemental slab at coordinate x and that conducted out at x + dx is, in equilibrium state, the amount that escapes into the air through the surface dx.

This equality, when

common factors are cancelled, is the equation: d29 / dx2 = 2q9 / kt 9

m A cosh a (x - B)

(2 )

a A & B

= constants of integration

when A & B are determined to satisfy boundary conditions, (3)

cosh aw1 The rate of heat dissipation from a unit length of fin is computed by an integration with respect to x from 0 to w 1.

The heat dissipation from an equal area (2w!) of cylinder wall at temperature 0O is:

33 HQ = 2 q0ow'


therefore: f 1 = H/Hq =» l/w’cosh aw f 1 =a tanh aw1 / aw1 This function (tanh aw1 / aw1) for which the single letter f 1 will be used is the function which will serve as the basis of discussion for much of the following work. Under average conditions of practice it will be found suf­ ficiently exact to serve as the basis for all computations. In those cases where it does not fit with sufficient


curacy it will be found convenient to use it as a principal term plus necessary correcting terms. Since a straight fin is essentially a circular fin with infinitely long radius, then with a high degree of accuracy it is permissible to use the equation: ^ f

= tangh (BH1) / BH'



H' = H (1 + 0.35 In p)



= tangh (Br0) / Br(2
no temperature varied more than 1° in 10 minutes.

After this equilibrium condition

was reached a test run of 20 minutes duration was begun. During that time the following measurements were recorded every 5 minutes: 1.

Inlet air dry-bulb and wet-bulb temperatures


Outlet air dry-bulb and wet-bulb temperatures


Fin surface temperatures


Tube surface temperatures


Outlet air velocity


Room temperature


LABORATORY RESULTS AND ANALYSIS OF DATA The data obtained from the experimental procedure were used for determining the inside and outside surface heat transfer coefficients and the over-all heat transfer coefficient of the coil or h h Q, and U, respectively. From these results several plots were made which will be explained and analyzed later in this paper.

I . DATA OBTAINED The test set-up was placed in operation and the necessary observations of the instruments made on six separate occasions.

The conditioner was in operation for

approximately seven hours on each of these days, including the stabilization period and the actual tests. Approximately five runs were made each day.

Out of

the five runs, one or sometimes two of the readings of a particular instrument farthest from the mean of all the readings were discarded.

Most of those discarded were due

to inconsistencies in the thermocouple readings. The remaining data which were not discarded were averaged each day into one complete set of consolidated data for that day's runs.

The average data for these six

56 days are tabulated in Table III.

The complete original

data are tabulated in Table IV.



Determination of surface coefficients. For the in­ side surface heat transfer coefficient, the refrigerant temperature is determined from the suction pressure.1


total amount of heat transfer, qt, is taken as the product of the air per unit of time and the difference in enthalpy per pound of air at the initial and final wet-bulb tempera­ tures.

From the existing physical dimensions of the eva­

porator coil, R (the ratio of the air-side to refrigerantside surface) is calculated.

Substitution of these values,

q^, tr, tg, and R in the following equation establishes the value of hj_. ^r “ ts“(Rqt/hfAN) Nomenclature was given in Chapter III. For the outside surface heat transfer coefficient, the values of A (outside surface area), N (number of rows of coil depth), and G (mass velocity of the air) can be

1 Lionel S. Marks, Mechanical Engineers Handbook (New York and London: McGraw-Hill Book Company, Inc., 1941), p. 2142.


1st run

2nd run

3rd run

4th run

5th run

6th run

Temperature, inlet air Dry bulb Wet bulb

74.65 61.00

73-57 61.00

72.14 62.70

72.14 62.00

73-57 61.50

71.50 61.50

Temperature, outlet air Dry bulb Wet bulb

52.20 48.80

54.20 48.20

53.60 50.00

54.10 50.20

53-40 50.20

52.40 48.80

Fin temperature Thermocouple No. 1 Thermocouple No. 2 Thermocouple No. 5

40.00 40.00 41.67

41.00 41.00 43.34

39-64 41.67 41.00

38.57 38.93 41.67

41.00 43.34 44.00

41.67 43.34 45.00

Tube temperature Thermocouple No. 3 Thermocouple No. 6

38.20 36.43

40.00 36.43

38.57 35.38

38.22 37.86

39-28 38.21

41.00 38.57

Inlet air temperature Thermocouple No. 4







Air velocity, Ft/min







Room temperature







Enthalpy Entering air (hi) Leaving air (ho) ^ - h3

27.15 19.64 7.51

27.15 19.31 7.84

28.84 20.30 8.54

27.85 20.41 7.44

27.50 20.41 7.09

27.50 19.64 7.85


1st run

2nd run

3rd run

4th run

5th run

6th run

848 6.5 254 3-96

843 5-73 244 3.62

845 5.36 281 3-40

839 4.82 294 3.42

344 6.4 212 3.76

840 6.77 227.5 3.88







Latent heat







Total heat







Latent heat/sensible ht







Latent ht/total ht




OJ O -=*• +



Sensible ht/total ht







Air flow, ft3/min h0 hi U Sensible heat (Btu/hr)

VJI 00

TABLE IV OBSERVED DATA Pin Fin Fin Fin Pin Temp. Temp. Temp. Temp. Temp. Thermo. Thermo. Thermo. Thermo. Thermo. No. 5 No. 6 No.2 No.3 No.l

Temp. Inlet Air Dry Bulb

Temp. Outlet Air

Wet Bulb

Dry Bulb

Wet Bulb

Air Vel. fpm

61.00 61.00 61.00 61.00 61.00

52.75 52.00 51.75 52.00 52.50

50.00 49.25 48.75 49.00 48.00

2097 2054 2060 2045 2059

61.00 61.25 61.00 60.75 61.00

54.25 54.00 54.00 54.25 54.50

47.75 48.25 48.25 48.50 48.25

2042 2043 2040 2045 2050

62.75 63.25 63.OO 62.00 62.50

54.75 53.00 53.00 53.75 53.50

50.00 50.25 50.00 49.75 50.00

2047 2060 2043 2045 2050

First Day!s Runs 40.00 36.00 40.44 39.66 39.90

40.00 34.50 40.30 40.44 39.26

42.04 34.00 41.67 42.00 40.97

36.25 37-00 36.47 34.00 36.00

38.16 38.20 38.44 38.00 38.20

75.70 74.50 73.80 74.75 7^.50

Second Day's Runs 41.00 34.50 40.30 41.44 41.20

42.17 40.45 41.25 40.13 35-00

42.66 42.98 43.00 44.72 36.OO

36.05 37-00 36.00 36.67 40.00

40.82 39.18 4o.oo 40.00 40.00

73-10 73.50 73.75 74.00 73.50

Third Day1s Runs 39.76 34.00 40.13 39.00 39.67

42.56 42.00 41.12 36.13 41.00

41.90 41.13 41.00 34.00 39-97

34.97 35-17 36.00 41.00 35.38

39-33 38.98 38.40 37.57 38.57

71.00 72.00 73.00 72.50 72.25

^ VO

TABLE IV (continued)

OBSERVED DATA Fin Fin Temp. Inlet Air Fin Fin Fin Temp. Temp. Temp. Temp. Temp. Wet Thermo. Thermo. Thermo. Thermo. Thermo. Dry No.6 Bulb Bulb No. 1 No.2 No. 5 No.3

Temp. Outlet Air Dry Bulb

Wet Bulb

Air Vel. fpm

62.00 62.00 61.80 62.00 62.20

53.80 54.25 54.00 54.00 54.50

49.50 50.75 50.00 50.50 50.25

2015 2030 2050 2040 2035

61.50 61.50 61.50 61.50 61.50

53.50 53-00 53.25 53-75 53.50

50.20 50.20 50.80 49.60 50.20

2042 2043 2040 2045 2043

61.50 61.75 61.25 61.00 62.00

51.75 52.75 52.00 52.00 51.50

47.50 48.00 47.25 48.50 48.75

2015 2040 2050 2036 2029

Fourth Day* s Runs 37.38 39.33 38.57 42.00 39.00

38.30 39.63 38.66 42.00 39.13

41.68 41.33 42.67 35.00 41.00

37-97 38.00 37.97 37.97 37.50

38.77 38.40 38.00 37.71 38.22

72.00 71.00 72.25 72.50 73.00

Fifth Day's Runs 42.47 43.85 43.70 43.34 36.00

4i.oo 4i.oo 41.25 34.00 40.75

43.30 44.03 43.67 45.00 44.00

38.20 38.11 37.95 38.21 34.00

38.97 39.67 39.28 37.87 40.00

73-57 73.50 73.25 73-57 73.57

Sixth Day's Runs 42.61 43.25 44.00 43.50 35.OO

41.87 41.67 42.00 41.50 41.31

45.00 45.00 40.00 44.75 45.25

38.57 39.00 38.67 37.87 38.57

41.77 39.98 41.00 39.23 41.00

71.50 71.50 71.75 71.25 71.50

cr\ o

6l determined from the physical dimensions of the unit.


age values of fin surface temperature, incoming air dry-bulb temperature and exit dry-bulb temperature may be obtained from readings taken during the different runs.


tion of these values in the following equation establishes the value of hQ . hQ = (0.24 G/AN) loge (t1-ts)(t3-ts) For nomenclature see Chapter III. The over-all coefficient of heat transfer of the coil may be calculated from the values of R, hj_, and hQ (already found) being used in the following equation 1 U =

R/hi + l/hG

Determination of apparent and actual mean coeffi­ cients .

Plotting of h/h0 on he :

for both finned plane

surfaces and finned tubing, the total amount of heat trans­ ferred, Q,

by the entire surface, A, is considered composed

of heat transferred, Qp, through the primesurface, Ap, heat transferred, Qe, through the extended surface, Ae . Q

h A 0 suppose


— Qe


=b heAe0e+ a hp






= he (Ae/A)f + Ap/A)



= 9e/9


= mean difference of temperature between prime surface and ambient

0e s mean temperature difference between extended surface and ambient The factor h is called the apparent coefficient of heat transfer based on the entire external surface as opposed to he, which represents the actual mean value over the entire surface.

Values of he can be calculated from

measured distribution of temperatures but results are of little value due to unavoidable inaccuracy.

The practical

procedure is to measure Q and 0 for use in the preceding equation to help in finding he by calculation.

With the

degree of fin efficiency, f, also being dependent on he, values of h/he calculated for the existing dimensions may be plotted against he . determined by f.

The resulting curve is essentially

In addition, values of h/he with the value

of h found by test and suitably selected values of h0, are calculated.

A second curve is obtained which intersects

the first at the level of h/he to be found. Figure 10 shows that as the value of he varies from 2 to 16, the value of f varies from .99 to .91.

The corres­

ponding variation in the value of h/he is shown in Figure 10.

The small variation in the value of h/he as the value

63 of he increases from 2 to 16 is attributed to the small change in the value of f.

However, the second curve in

Figure 10 which is represented by the ratio of h (constant value) to he (variable arbitrary value), has a comparatively larger variation in the value of the ratio as he increases. This is attributed to the fact that the curve is affected by only one variable, he . Relationship of surface coefficients to over-all expression. By plotting the over-all coefficient formula for various values of hQ and hj_, curves are obtained and are presented in Figure 11, curves of hQ .

Equation (l) was used for these

The slope of the hQ curve depends on the

ratio of the exterior to the interior surface of the coil. Check points found by test are shown on the plot and also arbitrarily calculated points out of the hj_ range of the tests. For any given coil design the hQ remains fairly p constant during the normal operation of the coil. The main variable during the normal course of operation seems to lie in the changing of hj_ and the resulting change in the over-all U factor. o

The rate of change in U is dependent

*" Maurice R. Hull, An Analysis of Direct Expansion Refrigeration Evaporator Performance Versus Condensing Unit Capacity," (unpublished Master's thesis, University of Southern California, Los Angeles, 19^2), p. 19*




01 j

t“ .I

A t A

66 on the slope of the h Q curve.

If more change Is needed to

obtain a desired temperature difference, then an increase in hQ must be obtained. For a given change in hj_ the rate of change in U is dependent on hQ and R.

This is clearly shown in the curves

of Figure 10. Evaporator capacity.

A plot of evaporator capacity

in Btu/hr. versus difference between surface temperature and refrigerant temperature for various values of hj_ is shown in Figure 12.

Check points found by test are shown

on the plot and the (ts-tr ) factor checks very closely with that found by test.

Equation (13) was used for plotting

these curves of h^. III.


Determination of the over-all coefficient of heat transfer.

First the outside surface coefficient of sen­

sible heat is found through the substitution of test data in Equation (13), page 36. hG = 0.24G/AN loge (t1-ts)/(t3-ts) rain min (156.42 ft2 )(2 rows)(3.21 ft2)


(3.21 ft2 )(2 rows) = 6.5 Btu/(hr)(ft2)(°P)


74.65-40.56 5 2 .2 0 -4 0 .5 6

ir b



FIGURE /£ J! !
















Next, the inside surface coefficient is found using Equation (13), page 40.

hi = 7 , - ^ -- ---(tg—tjp)AN _

where qt is GChi-l^)

(25)(1225 lb/(hr)(ft2 )(27.15-19.643) (40.56-21.805)(24.3)(2 ) 254 Btu/(hr)(ft2 )(°F)

Finally, U the over-all coefficient can be found using the values of hj_ and h0 previously determined in Equation (l), page 23.

U = ----1----- = --------- = 3.89 Btu/(hr)(ft2 )(°F) R +1 2 5 +I hi h0 25^ 575 Determination of sensible heat, latent heat, and total heat. Sensible heat: qs = G(0.24)(t1-t3) = (1225)(0.24)(74.65-52.2) =

6600 Btu/ft2 of face area or (6600)(3.21) = 21200 Btu/(hr)

Total heat:

Qt = G(hl " h 3 )


= (1225)(7.507 ) - 9210 Btu/ft2 or (9210)(3.21)

= 29500 Btu/(hr)

Latent heat:

qx = qt-qs = 29500-21200 = 83OO Btu/(hr) Ratio of latent heat to total heat: R = _£Li— =




-100 = 28.1$


Method for determining the over-all coefficient as explained in summary on page 23. Average hj_ from tests = 252.1 he from Figure 9 = 6.3

U 25___ +




.099 + .158

« 3.89 Btu/(hr)(ft2 )(°F)



SUMMARY AND CONCLUSIONS A great deal of work has been done in the determina­ tion of the heat transfer performance of finned coils. Usually the over-all heat transfer performance has been determined by the use of steam inside the tubes and air flowing over the extended surface on the outside of the tubes.

At other times the over-all performance has been

determined using a refrigerant, water or brine in the tubes. For many problems, however, It is highly desirable to know the individual components of resistance to heat flow, in­ cluding the outside film resistance, the metal resistance, and inside film resistance. Knowledge of the proportionate amounts of sensible and latent heat is essential, since for certain applications the main function of the coil is to remove moisture from the air and in most cases the removal of moisture is inci­ dental to the lowering of temperature.

These varying re­

quirements greatly influence the design of the coil and the condition under which it must be operated, particularly with regard to the temperature that must be maintained In the cooling medium. Since information concerning such important factors is not available, a series of heat transfer measurements

71 were made by the author in a laboratory of the Department of Mechanical Engineering at the University of Southern California.

These measurements were made on the evaporating

coil of a 3 h.p, Chrysler Air-Temp Conditioner model 3-SCD, with air flowing outside the banks of finned tubes and Freon-12 inside.

A summary of the experimental data is

tabulated in Table III, and results are graphically repre­ sented in Figures 10, 11, and 12. For air flowing outside the tubes and Freon-12 in­ side at the conditions outlined in Table III, the coeffi­ cients of heat transfer for the outside surface and for the inside surface have been determined.

The over-all coeffi­

cient of heat transfer may be estimated for all coils by estimating the inside surface coefficient by the use of Equations (13) a^d (l4); by estimating the apparent outside surface coefficient with the use of Equation (12) and the actual outside surface coefficient with the use of Equations (11) and Figures 2 and 10, and neglecting the metal resis­ tance through the tubes. For a coil of the type embodied in the model tested, a value of 4 Btu/(hr)(sq.ft.)(deg.ft.) may be considered to be within 3% of the mean value for the over-all coefficient of heat transfer. In general, It is not to be expected that banks of finned tubes used for cooling and dehumidifying air will

72 have the same coefficients of heat transfer as single tubes. The coefficients of heat transfer for banks of finned tubes were found to be higher than for single tubes at the same maximum air velocity; an increase probably caused by greater turbulence.

However, the effect will be counter­

balanced to an undetermined extent by the decrease in co­ efficients caused by the dripping of condensate from one tube to another. Another factor affecting coefficients is the ratio between the sensible and the latent heat transferred. Sensible heat is the heat removed in cooling the air. Latent heat is the heat removed in condensing the moisture. Coefficients for condensation are much higher than for cooling and it is expected that the coefficient will de­ crease as the proportion of sensible heat to total heat transferred increases. Factors affecting sensible heat and total heat re­ moved from the air are the diameters of the tubes, number of rows deep in the direction of air flow, and fin spacing. When the number of rows deep is the only variable, it is expected that an increase in number of rows will bring an increase in the capacity of coil.

Nevertheless, with a

decrease in number of rows to a minimum of 2 rows together with an appropriate decrease in fin spacing, the capacity of the coil tends also to increase.

The extent to which

the fin spacing may be decreased is such that the space between fins is large enough to prevent the water from being entrained and held suspended within the fins thus hampering the continuous flow of air at the desired velo-


BIBLIOGRAPHY American Institute of Chemical Engineers Transactions, Volume 29, 1933. Pp. “174-210. Ashley, C. M., "A Method of Analyzing Pinned Coil Heat Transfer Performance," . ■> Refrigerating ..... SMf—,.— .1 ,I■,■■■■Smd Engineering, ^^1 . rmrf 51:529-32, June, 1946. Brown, Aubrey I., and Salvatore M. Marco, Introduction to Heat Transfer. New York and London: McGraw-Hill Book Company, Inc., 1942. 225 PP"Chrysler Air-temp 3 h.p. Packaged Air Conditioning Unit," Service Manual. Dayton, Ohio: Chrysler Corporation, Air-temp Division. Colburn, A. P., and 0. A. Hougen, "Design of Cooler Condensers for Mixtures of Vapors with Noncondensing Gases," Ind. Eng. Chem., 26:1178-1182, 1934. Gleason, T. C., "Refinement of Evaporator Coils for Pack­ aged Units," Refrigerating Engineering, 55:549-51, June, 1948. Harper, D. R., and W. B. Brown, "Mathematical Equations for Heat Conduction in the Fins of an Aircooled Engine," Technical Report No. 158, National Advisory Committee for Aeronautics Annual Report,1921-1922, pp. 680-84. Heat Exchanger Tube Manual. Waterbury, Conn.: second edition, Scovill Manufacturing Company, 169 pp. Heating, Ventilating, Air Conditioning Guide, Volume 27, 1949. New York: The American Society of Heating and Ventilating Engineers. 1384 pp. Hull, Maurice R., "An Analysis of Direct Expansion Refri­ geration Evaporator Performance Versus Condensing Unit Capacity." Unpublished Masterfs thesis, The University of Southern California, Los Angeles, 1942. 459 PP. Journal of American Society of Refrigerating Engineers, Refrigerating Engineers Magazine, Vol. 57, April, 1949* Keenan, Joseph H., Thermodynamics. Seventh Printing; New York: John Wiley and Sons, Inc., 1948. 499 PP.

75 Marks, Lionel S., Mechanical Engineers Handbook, New York and London: McGraw-Hill Book Company, Inc., 1941. McAdams, William H., Heat Transmission. New York and London: McGraw-Hill Book Company, Inc., 1942. 459 pp. Miller, Morton A., "Forced-Air Refrigeration Evaporator Test Method," Refrigerating Engineering, 49:381-82, May, 1945. Severns, William H., Heating, Ventilating and Air-Condition­ ing Fundamentals. Eleventh printing; New York: John Wiley and Sons, Inc., 1949. 467 PP* Schmidt, Theodore E., "Heat Transfer Calculations for Extended Surfaces," Refrigerating Engineering, 56:32023, October, 1948. Shoop, Charles F., and George L. Tuve, Mechanical Engineer­ ing Practice. New York and London: McGraw-Hill Book Company, Inc., 1949. 513 PP. Tuve, G. L., and L. J. Seigel, "Performance of Surface Coil Dehumidifiers," Transactions A .S .H .V.E., 44:523, 1938. Wile, D. D., "Air Conditioning Coils -- Their Heat Transfer Problems," Refrigerating Engineering, 56:320-23* October, 19T O


APPENDIX The following appendix contains several photographs of the test set-up.

They are included in this paper only

to give a better idea of the equipment and test set-up previously described. In figure 13, the top panel, discharge grill, and return grill of the unit are removed (Refer to Figure 3> page 42), exposing the coil and attached thermocouple leads.

The potentiometer, switch, and cold junction are

seen on the left. Figure 14 is similar to Figure 13 but viewed at a slightly different angle.

The tubes may be seen entering

the coil in the right of the photo. Figure 15 does not truly represent the method In which the tests were rim but was the only way in which everything could be shown in one view.

First, the heaters

placed near the air intake were covered by a sheet metal shroud in order to better direct the heat into the unit. Secondly, the anemometer readings were taken at the dis­ charge end of the ducting as it would have been too hot to take readings in the position shown.

The psychrometer

shown directly behind the anemometer was read by shining a flashlight beam on it, as the shroud over the intake shut out most of the light.