NATIONAL BUREAU OF STANDARDS REPORT

1^697

EXPERIMENTAL RADIANT HEATING AND COOLING SYSTEM;

REFLECTION POINT, CINCINNATI, OHIO

by

R. S. Dill

P. R. Achenbach

0. N. McDorman

Report to

Federal Housing Administration

Washington 25, 9. C* .

U. S. DEPARTMENT OF COMMERCE

NATIONAL BUREAU OF STANDARDS
U. S. DEPARTMENT OF COMMERCE

Sinclair Weeks, Secretary NATIONAL BUREAU OF STANDARDS

A. V. Asl in. Director THE NATIONAL BUREAU OF STANDARDS

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this report.

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NATIONAL BUREAU OF STANDARDS REPORT

ms PROJECT MBS REPORT

1000-20-11830 October 3, 1958 ^&97

EXPERIMENTAL RADIANT HEATING AND COOLING SYSTEM;

REFLECTION POINT, CINCINNATI, OHIO

by

R. S. Dill

P, R. Achenbach

0. N. McDorman

to

Federal Housing Administration

Washington 25, D. C.

<^BS>

U. S. DEPARTMENT OF COMMERCE

NATIONAL BUREAU OF STANDARDS

The publication, repr

unless permission Iso

25, D. C. Such perm

cal ly prepared If th«

Approved for public release by the

Director of the National Institute of

Standards and Technology (NIST) on October 9, 2015.

part, is prohibited

idards, Washington

irt has been specifiort for its own use,

EXPERIMENTAL RADIANT HEATING AND COOLING SYBTEM

;

REFLECTION POINT, CINCINNATI, OHIO

by

R„ S. Dill, P. R. Achenbach, and 0. N. McDorman

Abstract

The performance of the heating and air conditioning

system in a house of novel design at "Reflection

Point," in Cincinnati, Ohio, was observed by members

of the Bureau staff at the request of the Federal

Housing Administration,, A special electric radiant

heating system in the same house was observed in

1953, and is covered in National Bureau of Standards

Report No. 2929« During the tests of 1953> heat was

supplied by means of electric resistance elements

in "coves" around the walls of the rooms near the

ceiling. Reflective wallpaper was used on the ceiling

and walls, and reflective drapes were provided for

the windows to conserve heat and to direct heat

radiated from the elements to the contents and occupants of the rooms. No other heat insulating means

was provided in the house.

Prior to the observations covered herein, the roof

was built up to include an air space and insulated

with a reflective material, the heating system was

changed to a gas-burning type, and a cooling system

was installed. Aluminum tubes in the coves, in

lieu of the electric heaters, are used to warm the

house in winter and cool it in summer. For this

purpose, an ethylene-glycol solution is circulated

through the tubes and the solution is warmed in

winter by a conventional boiler and cooled in summer by a refrigerating machine. The reflective

surfaces on walls, ceiling and windows were retained.

Ventilating air, supplied by a fan and duct system,

is "pre-conditioned" before entering the house by

a coil and elec trostaticc cleaner. This coil utilized

about 60 percent of the total cooling effect of the

refrigerating machine under the summer test condition

and about I4 O percent of the total boiler heat output

under the winter test condition. The effect of the

coil was not totally available for either heating

-2 -

or cooling the house because the ducts between the

coil and the ceiling grille were uninsulated. The

heat transmission coefficient for the cove coils,

when used for cooling, was found to be about

0.65 Btu per (hr„, sq. ft.) of coil surface for

each degree of temperature difference between coil

surface and room air during summer operation. The

heat transmission coefficient of these coils during

winter heating was nearly twice as great, probably

because a significant amount of convection heating

occurred. The tests of the heating system showed

that a satisfactory temperature distribution was

produced in the house at outdoor temperatures near

the winter design temperature in Cincinnati

although water temperatures in excess of 200 F

would have been required for adequate heating at

design temperatures. The tests of the cooling

system revealed that the condensing unit was

loaded in excess of the rated full load of the

three-horsepower motor for outdoor conditions less

severe than design summer donditions in Cincinnati

although the motor was not dangerously overheated.

The amount of cooling surface provided and the

method of heat transfer incorporated in the design

did not permit the condensing unit to produce a ton of refrigeration per horsepower motor load as

is conventionally claimed for air conditioning

units.

The observations indicate that the arrangement

functioned approximately in accord with the design

concept in that heat exchange by radiation occurred

between the coils in the coves and the floor and

contents of the rooms, assisted by the reflective

surfaces on the ceiling, walls, and over the windows. A small increase in heating and cooling

capacity appears desirable and can be obtained by

increasing the amount of heat transfer surface in

the cove and preconditioning coil, or possibly by

improving the cove coil arrangement to increase

its heat transfer. Ultimate utilization of the

design depends, of course, on desirability and

economic factors.
'
-3-

1. INTRODUCTION

At the request of the Federal Housing Administration,

the performance of the heating and air conditioning system

was observed in a house of unusual design, at "Reflection

Point," Cincinnati, Ohio. The system in this house is unconventional in that aluminum tubes, in coves in the walls near

the ceiling, are used for heating in winter and cooling in

summer. Great reliance is placed on radiant heat transfer to

or from the tubes in the design concept of the system. The

tubes are blackened outside, and reflective wall and ceiling

surfacing materials and reflective drapes over the windows

are used to promote heat transfer by radiation between persons

or contents of the house and the tubes. Ethylene-glycol

solution, called "brine," is circulated through the tubes to

serve as a heating medium in winter and a cooling medium in

summer. Air for ventilation is drawn from outside the house,

passes through an air cleaner and a tempering coil, and is

delivered through the hall ceiling, which is a perforated type

selected for the purpose. Data on the performances of the

system were obtained under both winter and summer conditions.

Results of previous work on this house, when it was equipped

with an electric radiant heating system, are contained in NBS

Report 2929, dated November 19, 19^3*

2. DESCRIPTION OF THE HOUSE

The house is located on a bluff and overlooks the city

of Cincinnati from the North. It Is a one-story structure of

about 2000 square feet floor area, with a flat roof, a ninefoot ceiling height, and with a concrete slab floor. The

floor is on grade, except under the bedrooms and study which

were built over a garage. Some dimensional data are given in

table 1.

The coves containing the heating-and-cooling coils are

gutter-like structures extending around the interior walls

about one foot down from the ceiling. They support^ partly

enclose, and conceal the tubes or coils that serve for both

heating and cooling the house. The surfaces facing the coils,

a s emi -cylinder of metal and a baffle, are polished, while

the coils are blackened, to promote heat exchange with the

rooms by radiation. The semi -cylinder or trough beneath the

coils Is insulated underneath and is concealed by a metal

enclosure, of rectangular cross section, which can be seen In

figure 3. The coils are composed of about 23OO feet of 3/4“

inchO.D. aluminum tubing with an aggregate surface area of

45l square feet.

TABLE 1 - ROOM SIZES AND VOLUMES AT REFLECTION POINT

Length

of Cold

Wall,

Ft

Ceiling

Height,

Ft

Gross

Cold Wall

Area,

Ft2

Window

Area,

Ft2

Door

Area,

Ft2

Net Cold

Wall

Area,

Ft 2

Living Room 39.6 9 356 100 0 256

Dining Room 27,8 9 250 78 0 172

Kitchen & Utility 29.0 8 232 68 18 II46

Foyer 11.2 9 101 51 20 30

Interior Hall 5*5 7.9 k3 8 20 a 15

North Bedroom 26.2 9 236 58 0 178

Soutli Bedroom 28.9 9 260 68 0 192

Study llj.6 9 131 53 0 78

Large Bath 7.0 8 56 6 0 50

Small Bath 0.0 10 21 3 0 18°

Whole House 189.8 - 1686 1493 58 113s

Other Heat

Transmitting Ceiling Floor Crack Room

Partitions Area, Area, Length Volume,

Ft 53 Ft 2 Ft2 Ft Ft 3

Living Room 0 377 377 0 3393

Dining Room 0 189 189 21 1701

Kitchen & Utility 0 25l| 25I4 51 2032

Foyer °a 176h 176 19

a

1581*

Interior Hall 113 ob 233 37 181*1

North Bedroom 0 160 160 56 11*1*0

South Bedroom 0 238 238 56 211*2

Study 0 a 129 129 h3 1161

Large Bath k9 57 57 11 1456

Small Bath 0 28 21* 0 21*0

Whole House 162 1608 l837d29I4 15990

a. Partitions adjacent to basement stairwell

b. Fresh air inlet

c. Walls for skylight

d. Total enclosed floor area for whole house was 1986 ft2

-4

- Embossed aluminum foil paper, lacquered in various

colors, was used for interior coverings on most of the wall

and ceiling area. Previous measurements of the emissivities

of samples of the various colors showed that the emissivity

ranged from a low value of 0,26 for the bronze color to a high value of O.I4O for the red color with an average of O.32

for eight samples of different papers.

Fabrics with moderate reflective properties were used

for draperies over many of the large windows. The emissivities previously measured for these fabrics ranged from 0.5b

to 0 . 69 . Some of these draperies are visible in figure 3.

The floors are covered with aluminum foil with the

reflective surface upward, a waffle-design foam rubber mat,

and wall-to-wall nylon carpeting in the living room, dining

room, den, and one bedroom. Cork floor covering is used for

the portion of the floor over the basement and garage.

Forced ventilation with cleaned outdoor air is provided

by means of a blower and an electrostatic precipitator to

reduce random infiltration at doors and windows, A finned

coil in the ventilation duct tempers the outdoor air in winter,

and dehumidifies and cools it during the summer. The fresh

air is introduced into the house through a perforated ceiling

in the hall.

Large picture windows, in every room of the house, total

about 1/3 of the gross exterior wall area. Most of the windows are of fixed plate glass with no provision for opening.

The heating system consists of a gas-fired boiler with

a rated output of 166,000 Btu per hour, a ^--horsepower circulating pump, a tempering coil in the inlet of the ventilation

duct, and the series-parallel circuits in the ceiling coves.

The tempering coil and cove coils are in parallel. There is

a by-pass line between the inlet and outlet of the boiler and

a three-way valve to control the mixing of heated boiler

water and water by-passed around the boiler in response to an

electronic control which senses the room temperature. An

aquastat on the outlet piping of the boiler controls the

operation of the gas burner.

The air conditioning system consists of a nominally

rated 3”hor sepower water chilling unit and circulating pump

which forces water-glycol mixture through the same coil in

the ventilation duct and the same cove coils that are employed

-5

-

in the heating circuit. The coil in the ventilation duct is

intended to do all of the dehumidification and a limited

amount of sensible cooling while the cove coils are intended

to do sensible cooling only. An evaporator pressure regulator

is used to prevent excessively low temperatures in the water

chiller. A cooling tower is employed for cooling the condenser water.

Inspection of the house construction, and published

descriptions, indicate that the following materials were

used:

Walls: Embossed aluminum foil wallpaper on 3/8-inch gypsum

board on 2 x Ij-inch studs with 3/8-inch plywood

sheathing and 25/32-inch vertical redwood tongueand-groove siding. There was no insulation in the

stud spaces.

Roof: The original roof consisted of 7 x 12 -inch wooden

beams spaced eight feet on centers, carrying l/lj-inch

plywood ceiling and 2 x 6-inch tongue-and-groove

planking covered with six layers of roofing felt.

The ceiling was covered with embossed aluminum foil

paper. The beamed ceiling and embossed foil ceiling

paper are visible in figure I4. Prior to the observations reported herein, this

original roof was "built-up'* by laying joists on the

top surface of the roof to support a layer of

Infra 6S insulation, one-inch sheathing boards, and

roofing felt. The joists were tapered to provide an

air space of two inches at the center and t en inches

at the edges, between the original roof and the

reflective insulation. This air space was closed at

the edges, but eight 6-inch gravity roof ventilators

installed along the two higher edges permitted air

exchange with the out-of-doors. These roof ventilators are shown in figure 5°

Floor: On grade, there was a layer of gravel, a concrete

slab floor with aluminum foil covering, waffle-design

sponge rubber mat, and nylon carpeting. Over the
• ' • *
-6

basement and garage, a 5/l^-inch layer of cork tile

covers the concrete floor. Orlan tile is used in

the kitchen, and quarry tile in the entry and hallway.

As shown in figure 1, there is a roof overhang to shade

the south windows from the midday sun, and vertical extensions

from the walls on the exterior of the south wall at two places

shaded the south windows from the late afternoon sun. Vertically-slatted radiation shields were attached to the exterior

of the east and west walls to decrease the solar absorption

in the early morning and late afternoon. Those on the west

wall are shown in figure 6.

3. TESTING EQUIPMENT

For measuring air temperatures, unshielded thermocouples

with polished, soldered junctions were used. Such couples

were installed at the 30-inch level at the middle of each

room. In addition, an eight-inch black globe thermometer

was used to evaluate the radiation-convection temperatures

near the middle of the living room. Wall and ceiling surface

temperatures were measured with 30-gage thermocouples

inserted beneath the wall paper. Wall heat transfer was

measured with three heat flow meters in the living room, two

on the walls at the 60-inch level, and one at the middle of

the ceiling, as shown in figure 3« Wet and dry bulb temperatures were measured with a sling psychrome ter , and air

velocities with a vane anemometer.

In addition, during the heating tests, the temperature

of the ethylene-glycol solution was measured at the boiler

inlet and outlet, at the tempering cdII inlet and .outlefi , hand

the flow was measured by k( wabetu meter* Air temperature was

measured at the inlet and outlet of the tempering coils and

at the ceiling in the hall where the air entered the house.

Gas consumption was measured with the house gas meter.

Running time of the burner was observed and recorded in percent.

During the cooling tests the capacity of the water

chiller was determined by measuring the rate of circulation

of the glycol-water mixture with a water meter and its temperature change as it passed through the chiller. The

enthalpy change of the ventilating air, as it passed through
.

.
-7-

the dehumidif ying coil, was determined by measuring the

entering wet bulb and dry bulb conditions, the change in dry

bulb temperature, and the amount of condensate drained from

the coil. The enthalpy of the ventilating air was also determined at the hall ceiling. Observations of the radiant energy

emitted at various areas on the wall and ceiling were made

with a radiometer. The power consumption of the condensing

unit was observed with a watthour meter. The indoor and outdoor dry bulb and wet bulb temperatures were determined with

a sling psychrome ter

.

[|. SYSTEM PERFORMANCE; COOLING

The cooling equipment was observed in operation during

three days in August 195^4 » and the cooling coil in the ventilating air duct was operated alone during one trial. This

coil is the same that serves for tempering the ventilating

air under winter conditions. In summer, its main function

is considered to be dehumidification of the fresh or ventilating air drawn in from outside. In a second trial, cooling

was effected by both the duct coil and the coils in the coves.

5. DUCT COIL; AIR CONDITIONING CAPACITY

During the first day of the tests, the dehumidifying

coil in the ventilating duct was operated for about three

hours, from 1:50 P» M. to l\:b,5 P« M. as the sole air conditioning means for the house. The outdoor conditions were,

initial; 85 . 5 P, D. B.; 67 percent R. H., and final; 83 F,

D. B. ; 63 percent R. H. The sky w as partly cloudy. Under

this condition the coil reduced the inside temperature from

83.5 F to 8l.7 F and the relative humidity from 70 percent

to 58 percent in about three hours.

Performance data on the dehumidif ying coil at 14:30 P.M.

are summarized in table 2. The total cooling effect of the

coil was about 17,000 Btu per hour, based on the coolant

circulation rate and temperature change, or 18,600 Btu per

hour, based on the air flow rate and enthalpy change. There

is, of course, a discrepancy of about 10 percent between

these figures, but inaccuracies of this order are expected

in field tests of this kind. Corrected values of specific

heat and density of the ethylene-glycol solution were used

in the calculation.
' .

’* ' '
TABLE 2

PERFORMANCE OF DEHUMIDIFYING COIL, 4:30 P.M.

Outdoor Conditions

Dry Bulb Temp. °F 83

Wet Bulb Temp. °F 73

Relative Humidity % 63

Enthalpy Btu/lb 38.

6

Sky Partly cloudy

Indoor Conditions, Room Center

30-inch Level

Living Room, Dry Bulb °F 82. 14 Living Room, 8-inch Globe °F 82.6

Dining Room, Dry Bulb °F 61.6

Kitchen, ” M °F 81.0

Foyer, " n °F 82.1

Study, " ” °F 81.5

North Bedroom, Dry Bulb °F 80.9

South. Bedroom, 11 11 °F 82.1

Average, " " °F 81.7

Living Room, R. H. 1° 59

North Bedroom, R. H. % 57

Coolant Circulation Rate gph 285

Chiller Inlet Temp. eF 66.2

Chiller Outlet Temp. °F 37.9

Useful Cooling Capacity Btu/hr 17,070

Power Used by Compressor Watts 2,760

Ventilating Air Rate cfm 286

Air Inlet D.B. Temp. °F 86.8

D.B. Temp, at Dehumidifier Outlet °F 69.6

D.B. Temp, at Hall Ceiling °F 70.1

Moisture Removal Rate lb/hr 9.27

Enthalpy Change in Vent. Air Btu/hr 16,660

Sensible Heat Removal Btu/hr 8,620

Latent Heat Removal Btu/hr 10,020

Sensible Heat Gain Between

Dehumidifier & Hall Ceiling Btu/hr 5,11*0

Potential Sensible Cooling

Capacity in Living Space Btu/hr 2,910

- 6-

The condensate collected from this duct coil was 9.27

pounds per hour, which accounts for about half of the total

cooling effect of the coil. The resulting drop in absolute

humidity was thus from 0.0152 to O.OO 63 pounds of water vapor

per pound of dry air. The average absolute humidity in the

house was O.OI 33 pounds of water per pound of air.

The air cooled by the duct coil warmed considerably in

the duct and plenums on the way to the hall ceiling. The

measured temperature in the coil discharge was about 50 F

and at the hall ceiling it was about 70 F. Therefore the

net cooling effect of the coil was somewhat less than 3000

Btu per hour under the existing condition.

The capacity of the refrigerating machine was greater

than the coil could utilize under the existing condition.

Evidence of this is the fact that the suction pressure was

reduced considerably below the setting of the evaporator

pressure regulator, which was 30 pounds per square inch.

6 . PERFORMANCE OF COOLING SYSTEM; ALL COILS OPERATING

During the second day, the performance of the -whole system, with both the duct coil and the cove coils in operation,

was observed from 9 A. M. to l\ P. M. The observations from

noon to I4 P. M. , when the conditions were most severe, are

reported in table 3 »

In the forenoon, a dry bulb temperature of about GO F

was maintained, and the humidity was reduced from 72 to 61

percent in the house. A mangle and an electric oven were

used part of the time.

In the afternoon, there were alternate periods of

cloudiness and of bright sunshine. The outside temperature

ranged from 8 I4 to 68 F, and the relative humidity averaged

about 71 percent. The difference between inside and outside

temperatures was about 6 F. The indoor temperature increased

about 1.2 F, while the relative humidity decreased about

three percent in the four-hour period. The indoor dry bulb

temperature averaged 79.5 F, and the relative humidity in

the living room and one of the bedrooms averaged 60 .

5

percent and 58 percent, respectively.
,* .
TABLE 3

OBSERVED PERFORMANCE OF AIR CONDITIONING SYSTEM

ALL COILS OPERATING: 12 M. to 4:00 P.M.

Outdoor Conditions

Dry Bulb Temp.

Wet Bulb Temp.

Relative Humidity

Enthalpy

Sky

Indoor Conditions, Room Center

30-inch Level

Living Room, Dry Bulb

Living Room, 8-in. Globe

Dining Room, Dry Bulb

Kitchen, '* " Foyer, 11 u Study, w " North Bedroom, 1 * 11 South Bedroom,” ” Average, n n Living Room, R. H.

North Bedroom, R. H.

Coolant Circulation Rate

Chiller Inlet Temp.

Chiller Outlet Temp.

Total Cooling Effect

Power Used by Compressor

Ventilating Air Rate

Coolant Temp, at Dehumidifier

Inlet

Coolant Temp, at Dehumidifier

Outlet

Air Temp, at Dehumidifier

Inlet

Air Temp, at Dehumidifier

Outlet

Moisture Removal Rate of

Dehumidifier

°F 85-14

°F 77.7

% 71

Btu/lb 41.2

Partly Cloudy

°F 79.4

°F 79.8

°F 79.3

°F 80.0

°F 79.5

°F 79.0

°F 79.1

°F 80.2

°F 79.5

% 60.5

% 58.o

gph 313

°F 52.1

°F I|0 . 6

Btu/hr 26,000

Watts 3,585

cfm 244

Op 40.4

°F 52.1

°F 85.4

°F 56.2

lb/hr 8.16

TABLE 3 (Cont'd)

Enthalpy Change in Vent. Air Btu/lb of air 15

Enthalpy Change in Vent. Air Btu/hr 16, II4 O

Sensible Heat Removal in Vent.

Air- Btu/hr 7,320

Latent Heat Removal in Vent.

Air Btu/hr 6,820

Sensible Heat Gain between Dehumidifier & Hall Ceiling Btu/hr 3,-460

Potential Sensible Cooling

Capacity in Living Space

(Vent. Air) Btu/hr 2,330

Net Cooling from Cove Coils by

Difference Btu/hr 9,860

Coolant Temp, at Cove Coil

Inlet °F 41

Coolant Temp, at Cove Coil

Outlet °F 50

Temp. Diff. Room Air to Cove

Coils °F 33

Heat Transmission Factor of

Cove Coils

Btu/hr (°F) (linear ft) 0

Btu/hr (°F) (sq ft) •0 Ceiling Surface Temp. Room

Center °F 79

Wall Surface Temp. 60 in above

Floor

South Wall, Living Room °F 81

West Wall, Living Room °F 77

5

to

2

6

13

65

2

8

8

-9-

Table 3 shows that the brine chiller was producing a

useful cooling effect of 26,000 Btu per hour under steady

operation. The pressures on either side of the evaporator

pressure regulator indicated that this valve was wide open,

thus permitting the condensing unit to operate at maximum

capacity for the existing conditions. The compressor motor

pox^/er averaged 3,585 watts. This corresponds to about four

horsepower at an estimated efficiency of 6

percent.

The condensate collected from the duct or dehumidffying

coil and the inlet and outlet dry bulb temperatures of this

coil indicated a total enthalpy change of 16,1140 Btu per

hour in the ventilating air. The average condensate rate

was 8.16 pounds per hour corresponding to a latent heat

removal of 6,820 Btu per hour, or 55 percent of the total

cooling capacity of the coil. The ventilating air was warmed

from 58 F to 70 P between the coil outlet and the perforated

discharge panel in the hall ceiling. This corresponded to

a sensible heat gain of 3,^60 Btu per hour in the supply

ducts and left a potential sensible cooling capacity in the

ventilating air of 2,336 Btu P er dour.

Subtracting the cooling capacity of the dehumidifying

coil from the total* cooling capacity of the chiller indicates

that the cooling effect of the cove coils was probably about

9,660 Btu per hour. This capacity represents a heat transmission of 0.13 Btu per (hr., linear ft.) of coil or 0.65 Btu

per (hr., sq. ft.) of coil surface for each degree F of temperature difference between brine and room air.

It should be noted in analyzing the performance of the

duct coil that the dry bulb temperature observed at the coil

outlet is not entirely consistent with the moisture content

at the same station derived from the moisture content of the

outdoor air and the condensate collected from the coil. This

introduces an uncertainty of about ten percent in the capacity

of this coil, but does not affect significantly the major

conclusions derived from the tests.

7. COOLING LOAD COMPUTATIONS

Computed cooling loads for the house, summarized in

table are based on heat transmission data contained in

the ASHAF ’'Guide .' 1 Table I4 indicates a cooling load of

25,680 Btuper hour for the conditions that prevailed indoors
,* '

••- *
TABLE I4 COMPUTED COOLING LOAD, BTU/HR AT

OBSERVED CONDITIONS AND DESIGN CONDITIONS

Observed Desi gn

Conditions Conditions

Outdoor D. B. Temp. °F 87 95

Outdoor W. Bo Temp. °F 78.5 78

Wind Velocity mph Calm 6

Indoor D. B. Temp. °F 80.1 80

Indoor W. B. Temp. °F 69. 14 67

Indoor R.H. % 58.3 51

Cooling Loads

Outside Walls Btu/hr 3680 7970

Windows & Doors Btu/hr 2790 7890

Inside Partitions Btu/lir 0 600

Ceiling Btu/hr 5890 7050

Floor over Basement Btu/hr 1050 2250

Slab on Ground Btu/hr 0 0

Infiltration Btu/hr 9390 10 860

Sub Total Btu/hr 23000 36620

People, I4 (light activity) Btu/hr 2000 2000

Electrical (estimated 200w. ) Btu/hr 680 680

Total Load Btu/hr 25680 39300

Computed Transmission Coefficients

,

u. Btu/hr ( sq f t )

( °F ) for Air Conditioning

Wind Velocity mph 0 6

Outside Walls U= 0.21|6 O.26I4

Ceiling and Roof U= 0.06)4 0 0 066

Windows and Doors U= 0.75 0.99

Floor over Basement U= 0.33 0.33

Inside Partitions u= 0.37 0.37

-loan d outdoors at I4 P. M„ on the second day of the tests. It

also indicates that the average cooling effect of 26,000 Btu

per hour observed for the brine chiller should have been just

adequate for the existing weather on the day of the tests.

This equality of cooling effect and cooling load was corroborated, in general, by the near ly steady state conditions

maintained indoors from noon to I4 P. M. while the condensing

unit operated continuously. The computed cooling load for

outdoor design conditions in Cincinnati was 39,300 Btu per

hour, assuming indoor conditions of 60 F dry bulb and 67 F wet

bulb, and outside conditions of 95 F dry bulb and 76 F wet

bulb.

The equivalent temperature method described in the

’'Guide" was used for these computations. The shading of the

exterior walls and windows by the roof overhangs, the wall

extensions on the south side, and the slatted radiation

shields on east and west sides were taken into account. The

air space between the built-up roof and the original roof

was considered to be a dead air space since the arrangement

of the ventilation hoods did not assure positive air circulation through this space. The access of the chilled ventilating air to the underside of the hall and kitchen roof

was taken into account in computing the roof load of these

areas. The reflective surfaces of the paper on the walls and

ceiling were taken into account in computing the heat transmission coefficients listed at the bottom of table l\. An

occupancy of four people was assumed for both computations

since four people were present in the house during the test

of the air conditioning system. The air infiltration rate

was assumed equal to the observed air delivery of the ventilating blower.

The computations in table l\ indicate that the outside

walls, windows, ceiling, floor, and infiltration represented

15.1, 10 . 6 , 22.9, and 36.5 percent of the total cooling

load, respectively, at the observed conditions. Observations

of the heat exchange at the wall and ceiling surfaces made

with heat flow meters and with a radiometer do not, in some

cases, agree with the values obtained when the recommended

ASHAE computation methods were applied, as will be seen later.

Table 5 summarizes the radiant heat transfer from

selected areas of the walls, ceiling, and drapes, and

through the undraped windows in the living room and dining

TABLE 5

RADIANT HEAT TRANSFER AS MEASURED WITH A

RADIOMETER , 1:30 P.M. , August 28

Heat Transfer Rat

Source Btu/hr (sq ft)

Heat Flow Meter, South Wall, Black 3*7

Heat Flow Meter, South Wall, Reflective 1.9

South W

r all, After Washing 0.7

Celling, Reflective O.I4 West Wall, Reflective 0 . Lj Heat Flow Meter, West Wall, Reflective 0 .I4 West Wall Above Window, Reflective 2.0

West Drape 12.2

South Drape 12.2

North Drape 13.0

South Window Above Horizon 161.

North Window Below Horizon 31.

North Window Above Horizon 1I4I4

Indoor-Outdoor Temp. Difference at

1:30 P. M., °F (5.0)

- 11 -

room of the house at i :30 £• M-* on the second day of the

testa as measured by a radiometer. These values do not

include the convection component of heat transfer.

Figure 7 shows the heat transfer through the south and

west walls and the ceiling of the living room as measured

by heat flow meters attached to the inner wall surface at

the 60-inch level. The surfaces of the meters exposed to

the room were covered with the same reflective paper used

on the walls and ceilings of the house except for one of the

two meters on the south wall, which was left uncovered.

Figure 7 shows that heat was being lost through all of these

surfaces at 9 A. M. when the observations began. Heat gain

commenced through the south wall soon after 9 A. M, , through

the west wall at noon, and through the ceiling between 1 and

2 P. M.

A few comparisons can be made between the datain figure

7 and table 5» Figure 7 shows that the total transfer

between the south wall and room at 1:30 P. M. was 6,1 Btu

per (hr., sq. ft.) for the black heat flow meter and I4.5 Btu

per (hr., sq. ft.) for the reflective heat flow meter.

Table 5 indicates that the radiant components were 3.7 and

1.9 Btu per (hr., sq. ft.), respectively, for these two surfaces, leaving convection components of 2 .I4 and 2.6 Btu per

(hr., sq, ft.) in the two cases.

These comparisons of the heat transfer through the

black and reflective heat flow meters do not mean that the

heat transfer for the entire wall could be changed in the

same ratio by changing the emissivity of the inner surface,

but only that portion of the wall inside of the stud spaces.

The stud space acted as a plenum chamber behind both heat flow

meters because the heat flow meters were of small area. If

large sections of the wall had been covered with wallpapers of

different emissivi ties, the air temperatures in the stud

spaces would be different behind the two coverings. The temperature

in the stud space back of the r ef lective wallpaper would probably

be higher than that behind the non-ref lective wallpaper, and the

difference in heat transfer through the entire wall would be

less than the data in figure 7 and table 5 Indicate. Table

shows that washing the r eflective paper on the south wall significantly reduced the radiation heat transfer. The owner advised

that this paper had been on the wall for several years.
'

.

.
- 12 -

Figure 7 shows that the total heat transfer of the west

wall at 1:30 P. M. was 0.9 Btu per (hr., sq. ft.), whereas

the radiometer measurement of the radiant component was 0 .I4 Btu per (hr., sq. ft.) at the same time and place. The

radiant heat transfer of the west wall above the window where

it was not externally shaded was five times as high as at the

heat flow meter where the wall was shaded by the slatted

radiation shield outdoors. The heat flow meter on the ceiling

showed no heat transfer at 1:30 P. M. , whereas the radiometer

indicated a heat gain of O.J4 Btu per (hr., sq. ft.).

Some comparisons are also possible between the heat flow

meter data in figure 7 , and the computed heat transfer through

the walls and ceiling of the living room, based on data in

the ASHAE "Guide.” Figure 7 shows that the heat transfer

through the south and west walls of the living room at 14 P. M.

was 5.2 and 2.1 Btu per (hr., sq. ft.), respectively. The

"Guide" data indicate that the heat flow through a shaded

wall at I4 P. M. would be about l.^d Btu per (hr., sq. ft.),

based on the computed heat transmission coefficients in

table I4 . This comparison indicates that the south and west

walls were not the equivalent of a shaded wall. Even though

they received no direct insolation, they were receiving

some diffuse radiation from the sky and from the patio in

the case of the south wall. Figure 7 shows a heat transfer

through the ceiling at I4 P. M. of 0.75 Btu per (hr., sq. ft.),

whereas the "Guide*4 data indicate a value of 2 . 9)4 Btu per

(hr., sq. ft.), indicating that the ceiling was a better heat

barrier for downward heat flow than the computed heat transmission coefficient indicates.

Table 5 shows that the drapes on the north, south, and

west windows were radiating from 12 to 13 Btu per (hr., sq. ft.)

to the rooms even though shaded from direct sunlight and,

further, that sky radiation ranging from II4

I4 to 161 Btu per

(hr., sq. ft.) entered the north and south windows when

they were uncovered. A completely shaded window would have

transmitted only about 5 Btu per (hr., sq. ft.) for the

existing temperature difference and air motion. The "Guide"

does not list heat transmission data for windows shaded on

the outside and covered with reflective draperies on the

inside, but suggests that a shaded window should be treated

as though its heat gain resulted primarily from the temperature

difference existing on the two sides. The radiometer data in

table 5 indicate that the heat transfer through the shaded,

draped windows was more than twice that based on the air

temperature difference alone.

-13-

The data obtained with the heat flow meters and radiometer, though grossly incomplete, indicate that the cooling

loaa computations summarized in table I4 do not accurately

represent the actual conditions that prevailed, and that it

was perhaps fortuitous that the total computed cooling load

was equal to the observed cooling capacity of the air conditioning system for the period under consideration. The

foregoing analysis of the heat flow meter and radiometer

data indicates that the actual heat flow through the south

and west walls and through the draped windows was probably

greater than the computed values shown in table I4, and that

the actual heat flow through the ceiling was probably less

than is shown by table L|

The capacity of the refrigerating machine was shown, by

the best indications, to be sufficient to maintain almost

half the design inside-outside temperature difference usually

assumed for Cincinnati when the ventilating air load was about

12 percent below that assumed for design conditions, and when

the sun did not shine continuously. More system capacity

would be desirable under the existing conditions and it may

be considered inadequate for more severe conditions, including

the design condition for Cincinnati.

0. SURFACE TEMPERATURES

Table 3 shows the average temperatures observed on the

wall and ceiling surfaces in the living room during the

period from noon to I4 P. M. The temperatures were measured

with 30-gage thermocouples inserted just underneath the wallpaper, through openings made with a darning needle. The

average surface temperature at the center of the ceiling in

the living room was 79.2 F, differing 0.2 F from the average

dry bulb temperature of the air at the 30-inch level in the

room during the same period. The average surface temperatures

of the south and west walls of the living room at the station

60 inches above the floor were 81.8 F and 77.8 F, respectively. The latter figure is probably inaccurate since it

would indicate that the wall was gaining heat from both

indoors and outdoors, whereas figure 7 indicates an outward

heat flow during the afternoon. The temperature indicated

by the globe thermometer in the living room was only O.I4 F

higher than the air temperature measured with small thermocouple wire in the same room. This indicates that the globe

was receiving a relatively small amount of radiated heat

from the surrounding surfaces.
> •
Figure 8 shows the ceiling surface t emper atur es measured

in the living room from the center to the edge of the

cooling cove by inserting a thermocouple underneath the

reflective paper. These data reveal ceiling surface temperatures ranging from 79 F to 6l F at distances greater

than 15 inches from the edge of the cove. There was a decrease in surface temperature near the cove which was

caused, presumably, by the greater radiation to the cove

coils. When the radiometer was sighted on the surface of

the cove coils, it indicated a heat transfer to the coil

surface of 20.7 Btu per (hr., sq. ft.). This agrees approximately with the value of 21.9 Btu per (hr., sq. ft.) that

can be derived from the data in table 3 Dor total heat transfer to the cove coils and their total surface area.

9. SUMMER COMFORT CONDITIONS

The average indoor conditions maintained during the

second day of tests were 79.5 F dry bulb and 59 percent relative humidity, corresponding to an effective temperature of

about 74*6 F in the ASHAE effective temperature chart. It

is recognized that the ASHAE chart does not necessarily

reflect comfort conditions accurately when there is a greater

or lesser than normal amount of radiation present in the

environment. The reflective papers on walls and ceiling

affect the radiant heat exchange between the occupants and

these surfaces. The two observers engaged in these tests

were not at optimum comfort. They were engaged in intermittent light activity while making observations; they noted

slight perspiration on their foreheads and were aware of the

low air motion in the living space.

10. SYSTEM PERFORMANCE; HEATING

The operation of the heating system was observed

throughout one night while the outdoor temperature decreased

gradually from 32 F at 9 P. M. to 9 F at 9 A, M. This

decrease in temperature was accompanied by snowfall and a

wind ranging in velocity from 0 to 9 miles per hour.

As the weather became colder during the night, the

operating time of the gas burner increased, the by-pass

valve around the boiler remained in a closed position more

of the time, and finally, the aquastat setting of the boiler
'

,

‘
-15

had to be raised. Between 1 A. M. and 2 A. M. , with an outdoor temperature of 26.5 F, the gas burner operated 35 percent of the time, the boiler outlet temperature averaged

I 63 F, and the by-pass valve cycled between 5/& closed and

fully closed. Between 6 A. M. and 9 A. M. , when the outdoor temperature averaged 10 F, the gas burner operated 63.7

percent of the time, the boiler outlet temperature averaged

188 F, and the by-pass was closed at all times. During this

latter period the aquastat setting of the boiler had to be

raised to 200 F to provide satisfactory room temperatures.

The differential of the aquastat was about 25 F. The average

gas consumption rate between 8 A. M. and 9 A. M. was I 3 I4 cubic feet per hour, corresponding to a boiler output of

98,500 Btu per hour if the boiler were assumed to be 70 percent efficient. . The performance of the heating system is summarized in

table 6 for the hour between 8 A. M. and 9 A. M. Air was

passed through the preheater at a rate of 305 cubic feet per

minute and was heated approximately 121 F above outdoor temperature. This preheater is the same coil that served to

dehumidify the fresh air brought in for ventilation under

summer conditions. The above figures indicate a heat transfer of 39,860 Btu per hour in the preheater coil. This

value of heat transfer may be in error because the outlet

air temperature increased and decreased as the boiler outlet temperature increased and decreased, so an accurate

average could not readily be determined. The air decreased

in temperature nearly 30 F between the preheater outlet and

the discharge grille in the hall ceiling. This temperature

drop corresponded to a heat loss of 9,350 Btu per hour in

the ducts and hall plenum. Thus, the ventilating air at the

hall ceiling could provide approximately 11,800 Btu per hour

to offset transmission heat loss. By difference, the cove

coils were computed to be supplying heat at a rate of about

53,800 Btu per hour after correcting the difference between

boiler output and preheater output for the computed piping

loss.

11. TEMPERATURE DISTRIBUTION

The air temperatures at the 30-inch level at the center

of the several rooms averaged 66.7 F, whereas the black

globe thermometer at the center of the living room indicated

a temperature J4.7 F above air temperature at the same

TABLE 6

PERFORMANCE OF THE HEATING SYSTEM,

8jOO to 9:00 A. PI. , February 11

Outdoor Temp.

Indoor Temp., Room Center, 30-in.

Living Room, 8 -in. Globe

Living Room, Air

Dining Room, Air

Kitchen, Air

Foyer, Air

North Bedroom, Air

Study, Air

South Bedroom, Air

Average, Air 30-in. Level

°F

Level

°F

°F

°F

°F

°F

°F

°F

°F

10.0

72.5

67.8

66.6

66.0

67 .6

65 c 8

66 .

66.2

66.

Glycol-Water Circulation Rate

Average Boiler Inlet Temp.

Average Boiler Outlet Temp.

Instantaneous Gas Consumption Rate

Percent Running Time

Average Gas Consumption Rate

Computed Heat Output Rate (70/ eff.)

gal /hr 236 . °F 151.6

°F 186 .

cu ft /hr 210.3

63.7

cu ft /hr I 3 I1 .O

Btu/hr 96,490.

Ventilating Air Rate Through Preheater cfm

Air Temp, at Preheater Outlet °F

Air Temp, at Hall Ceiling °F

Heat Transfer in Preheater Btu/hr

Heat Loss Between Preheater &

Hall Ceiling Btu/hr

Sensible Heat Available at Hall

Ceiling for House Heating Btu/hr

305c

131.1

102.7

39,860.

9,350.

11 , 800 . Net Heat from Cove Coils, by Difference Btu/hr 58,600.#

Ceiling Surface Temp., Center of

Living Room °F 81.

Living Room Temp., 60 in. Above Floor

South Wall Surface, Reflective °F 63.7

South Wall Surface, Black °F 66.6

Air 2. in. from South Wall °F 68.

4

#This value includes piping loss from the boiler, which was

computed to be approximately 5,000 Btu/hr.

-16-

location. These temperatures are summarized in table 6. The

maximum variation in air temperature at the 30-inch level for

the seven room centers was 2 F. The difference between globe

temperature and air temperature in the living room indicates

a considerable radiant flux in that room and shows that the

mean radiant temperature would be several degrees above the

globe temperature. The observers were comfortable during

most of the test, but some feeling of coolness was experienced

about 6 A. M. This could have been the result of long hours

of duty and lowered metabolism.

The surface temperature of a heat flow meter on the

south wall covered with reflective wallpaper was 63. 7 F,

whereas the temperature of a similar heat flow meter with

a black surface, also mounted 60 inches above the floor,

was 66.6 F. The ceiling surface temperature at the center

of the living room was 61. 5 F during the coldest period of

the test. The ceiling surface temperature at mid-length of the

living room, from the edge of the south heating cove to the

edge of the north heating cove, is shown graphically in

Figure 9 for an outdoor temperature of 15 F. The rise in

temperature near the cove edges indicates greater heat transfer from the coves in this area by radiation or convection,

or both. The heat loss through the ceiling would also be

greater near the cove edges. The arithmetical average

ceiling temperature for the traverse shown in figure 9 was

67.7 F. The temperatures on the aluminum reflector shield

in the cove ranged from 111 F to 112.5 F.

The air temperatures in a vertical line from floor to

ceiling at the center of the living room, and at a station

36 inches from the south picture window in the living room,

are plotted in figures 10 and 11, respectively. Figure 10

shows that the rug surface in the living room was 3 to I4 F

warmer than the air one inch above the surface as a result

of the radiant energy absorbed by the rug. The minimum air

temperature was observed in the zone from 6 to 2l\ inches

above the floor, and the floor surface temperature was

approximately equal to the air temperature b0 inches above

the floor. The air temperature increased rapidly in the

zone from 80 to about 100 inches above the floor and dropped

sharply in the last inch below the ceiling surface. The

ceiling surface was 8 or 9 F cooler than the air one inch

below the surface. The shape of the air temperature curve

near the ceiling indicates considerable heat transfer from

the coves by convection. If only radiant heat were striking

the ceiling surface and being reflected, the air temperature

-17-

beneath the ceiling surface would be cooler than the surface# The coves were 12 inches high and open at the front

face, so convection currents could be set up in the coves

that would direct warm air outward over the ceiling surface

as cooler air entered at the lower edge of the cove opening#

The presence of convection from the coves is also indicated by a comparison of the maximum air temperature shown

in figure 11 with that for the corresponding outdoor temperature in figure 10. This comparison shows that the air

temperature one inch below the ceiling was 95 F at a distance 2lj inches from the cove edge, whereas the temperature

at the same level was only 86 F at the room center. This

indicates an outward flow of warm air near the ceiling from

the cove toward the center of the room. This movement of

warm air from the cove must contribute to the difference in

ceiling surface temperature between perimeter and room center

as shown in figure 9. These convection currents increase

the amount of heat that can be transferred from the cove

coils and do not adversely affect the comfort in the living

zone because the stratum of warm air is above the 80-inch

level

Figure 10 shows that the minimum air temperature near

the floor was lowered about 2 F and the maximum air temperature near the ceiling raised about 7 F, whereas the

rug surface temperature was affected only one-half degree

when the outdoor temperature decreased from 30 F to 10 F#

A comparison of figures 10 and 11 shows little difference

in air temperatures below the 80-inch level between the

room center and a station 36 inches from the south picture

window, which was only 15 inches in front of the bookcases

beneath the window.

12. HEATING LOAD

The outdoor weather conditions and the computed heating

loads of the various components of the house for the observed

conditions are summarized in table 7o The heat transmission

coefficients used for these computations are listed at the

bottom of the table. Table 7 indicates that the heat loss

of the house was 86,900 Btu per hour exclusive of duct

losses and piping losses for an indoor -outdoor temperature

difference of 57 F. Of this total, 20 percent is attributed

to the exterior walls, 35»5 percent to the windows and doors.
.
TABLE 7

COMPUTED HEATING LOAD AT OBSERVED

CONDITIONS, BTU/HR, 8:00 to 9:00 A.M., February 11

Outdoor Temp. °F 10.0

Wind Velocity mph 7.U

Indoor-Outdoor Temp. !Diff

Heating Loads

°F 56.9

Outside Walls Btu/hr 17,1(00

Windows and Doors Btu/hr 30,800

Inside Partitions Btu/hr 900

Ceiling Btu/hr 11,600

Floor over Basement Btu/hr 14,160

Slab on Ground Btu/hr 3,290

Infiltration Btu/hr 18,71(0

Total Heating Load of Living Space Btu/hr 86,890

Loss from Ventilating Air Ducts Btu/hr 9,350

Total System Load (excluding Piping Loss) Btu/hr 96,250

Computed Transmission Coefficients,

Btu/hr (sq ft) (°F) for Heating

Ceiling and Roof 0.092

Outside Walls 0.270

Windows and Doors 1.02

Floors over Basement O .33

Inside Partitions O .37

-18-

13 * U percent to the ceiling, 8.6 percent to the floors, and

21.6 percent to infiltration of outdoor air. The air infiltration was assumed to be equal to the forced ventilation

provided by the ventilation blower. It represented about 1.15

air changes per hour for the entire house.

The heat loss from the ventilating ducts was computed

to be 9,350 Btu per hour, based on the observed temperature

drop between the preheater outlet and the hall ceiling

grille. Thus the total system load, exclusive of loss from

the hot water piping, was computed to be 96,250 Btu per hour.

If 5,000 Btu per hour were added for piping loss, the total

boiler load would be 101,250 Btu per hour compared with a boiler output of 98,590 Btu per hour derived from the

observed gas consumption and an assumed boiler efficiency

of 70 percent.

Some inaccuracy in field observations is usually expected

and the heat flow meter readings taken on the south and

west walls and ceiling of the living room are not concordant

with the computed heat transmission coefficients in table 7

in some respects. The heat flow indicated by the heat flow

meters is plotted against time for the entire test in

figure 12. Figure 12 shows an average heat flow for the

south and west walls, using the value obtained with reflective wallpaper for the south wall, of about 10 Btu per

(hr., sq. ft.) for the time between 8 A. M. and 9 A. M.

The computed value, using the U-f actor in table 7 and the

observed temperature difference, would be 15»5 Btu per

(hr., sq. ft.). The corresponding observed value for the

ceiling averaged about 5*5 Btu per (hr., aq. ft.) in figure

12 for the same time interval, whereas the computed value

was 7.1 Btu per (hr., sq. ft.). Probably the heat capacity

of the walls and ceiling construction caused the instantaneous values observed with the heat flow meters to be too

low, but the lag of the exterior walls was only one to two

hours so heat lag would hardly account for the disparity

between observed and computed values in the case of the

walls

Thus the heat flow meter data indicate that table 7

provides only an approximate accounting for the heat loss

in the structure and they also create some uncertainties

about the accuracy of the computed heat transmission coefficients. It has been shown by test and reported in the

literature that the heat transfer through a cavity wall is

not uniform from floor to ceiling.

Although the heat flow meters were near mid-height of the

wall, they might not have been indicating an average heat

flow.

The two heat flow meters on the south wall, with black

and reflective surfaces, indicated a considerable difference

in heat flow as shown in figure 12. As was the case for the

air conditioning test, the stud space of the wall provided

a plenum of equal temperature behind both meters. Thus the

difference in indicated heat flow shown in figure 12 is the

effect of placing a reflectively-covered and a black heat

flow meter on a sheet of ^/Q-inch gypsum board over this

constant temperature plenum, and does not provide an adequate

evaluation of the effect of using reflective wallpaper on the

heat transfer through the whole wall. In this case, the

reflectively-covered meter provides the better value of heat

flow because the black meter was only about 16 square inches

in area on an otherwise reflective wall.

13. DISCUSSION AND CONCLUSIONS

Accurate data are difficult to obtain under field conditions and some discrepancies are apparent in that contained

in this report. However, conclusions concerning this type

of house and its heating and cooling system are supported

by the data as follows?

The results show that considerable heat can be transferred by radiation from cove coils to the floor and contents

of a house when the wall and ceiling surfaces and window

drapes are made of reflective materials in accord with the

design concept. Under the observed winter condition, the

ceiling was warmer than the floor so that heat reached the

floor in part by radiation from the ceiling. However, the

ceiling was not as warm nor as emissive as that ordinarily

required by a ceiling radiant heating system, and the

difference must be attributed to radiation by the coils.

Use of more highly reflective material on ceiling, walls,

and drapes would be expected to enhance the radiant transfer by the coils and lessen that from the ceiling.

• * Economy of heating equal to that expected of other

typical systems probably can be attained with systems of the

type investigated. A criterion for house heat requirements

-20 -

in use by FHA is 55 Btu per hour for each square foot of

floor area at design conditions. The observed total heat

loss of the house investigated was about 100,000 Btu per

hour for an indoor-outdoor temperature difference of about

57 F. This amounts to 50 Btu per hour- for each square foot

of floor area, since the house had a floor area of about

2000 square feet. At an indoor-outdoor difference of 70 F,

the adjusted heat loss is

(70/57)5° = 61 Btu per hour for each square foot.

It is plausible to credit the reflective interior surfaces of the house with saving since the black bulb thermometer indicated a temperature nearly five degrees higher

than the air temperature under the most severe test condition. Assuming an indoor-outdoor temperature of 65,

instead of 70 F, the adjusted heat loss becomes

(65/57)5° = 57 Btu per hour for each square foot.

Both these adjusted heat losses exceed the criterion,

but better economy can be attained by such changes as

reducing the ventilating air, use of more highly reflective

surfaces on walls and ceiling, insulation in exposed walls,

insulation of the fresh air duct and plenum, etc.

The observed heat transfer coefficient for the cove

coils when used for cooling was about 0.65 Btu per (hr., sq

,

ft.) for each degree F of temperature difference between

coil surface and room air. It would be more logical to

base the heat transmission coefficient for summer cooling

on the difference between the mean radiant temperature of

the room and the coil surf ace temperature since there was

practically no convection during summer operation. Howe'ver,

the difference between mean radiant temperature and air

temperature was so small during summer operation as to make

this refinement unimportant.

The heat transfer coefficient of the cove coils for

heating was about twice as great as for cooling. This

difference is attributable to the more pronounced convection currents around these coils when they were used for

heating. This convection is regarded as responsible for

the relatively high air temperatures near the ceiling.

When the coils were cold, the air probably stratified

around them in the gutter-like structure used to conceal
.

’

-
-21 -

them from below. This suggests that this structure might

be slotted or perforated in future designs to permit passage

of air and thus improve heat transfer by convection and

increase their capacity for cooling.

The coil in the intake duct, used for tempering the

outside air drawn in for ventilation in winter and for

cooling and dehumidifying that air in summer, accounted

for about I4 O percent of the heating load in winter and

about 60 percent of the cooling load in summer under the

test conditions. This can be construed to mean that the

ventilating rate is greater than necessary and that it

can be reduced in future designs. It is possible to arrange

for some recirculation through this coil to effect some

economy of operation which is particularly desirable under

the summer condition.

The cooling capacity required, 26,000 Btu per hour,

agreed well with the computed cooling load, 25,680 Btu per

hour, under the observed condition. This is regarded as fortuitous because the heat flow meters did not agree in

all cases with heat transfer rates computed by conventional

methods. Indications are that actual heat gains through

the south and west walls are greater than computed and that

the gain through the roof is less than that computed. This

emphasized the complexity of the heat gain problem and the

need for further research concerning it.

The observers reported a feeling of coolness during

part of the most severe heating test and perspiration on

the forehead during the most severe cooling test. Neither

of these tests was conducted under as severe a condition

as the design condition for Cincinnati. The tests showed

that a satisfactory temperature distribution was produced

in the house at an outdoor temperature of 10 F although it

was indicated that boiler water temperatures in excess of

200 F would have been required to adequately warm the house

at the design winter temperature of 0 F in Cincinnati.

Summer operation of the cooling system at conditions

less severe than design summer conditions showed that the

proportions of the components were such that the three

- 22 -

horsepower condensing unit was loaded to approximately four

horsepower while producing a cooling capacity of about

26,000 Btu per hour. This performance was associated with

an average dry bulb temperature of 79.5 F and an average

relative humidity of about 59 percent in the house during

the period from noon to I4 P. M. on the day of the test.

These data on the operation of the system and the comfort sensations cited above suggest that a greater heat

transfer surface area in the cove coils would have been

advantageous in both heating and cooling. However, the

same result might possibly be obtained with the existing

coil area by an arrangement that would increase the

radiant and convection heat transfer rates.

THE NATIONAL BUREAU OF STANDARDS

Functions and Activities

The functions of the National Bureau of Standards are set forth in the Act of Congress, March

3, 1901, as amended by Congress in Public Law 619, 1950. These include the development and

maintenance of the national standards of measurement and the provision of means and methods

for making measurements consistent with these standards; the determination of physical constants and properties of materials; the development of methods and instruments for testing materials,

devices, and structures; advisory services to Government Agencies on scientific and technical

problems; invention and development of devices to serve special needs of the Government; and the

development of standard practices, codes, and specifications. The work includes bas*c and applied

research, development,' engineering, instrumentation, testing, evaluation, calibration services, and

various consultation and information services. A major portion of the Bureau’s work is performed

for other Government Agencies, particularly the Department of Defense and the Atomic Energy

Commission. The scope of activities is suggested by the listing of divisions and sections on the

inside of the front cover.

Reports and Publications

The results of the Bureau’s work take the form of either actual equipment and devices or published papers and reports. Reports are issued to the sponsoring agency of a particular project

or program. Published papers appear either in the Bureau’s own series of publications or in the

journals of professional and scientific societies. The Bureau itself publishes three monthly periodicals, available from the Government Printing Office: The Journal of Research, which presents

complete papers reporting technical investigations; the Technical News Bulletin, which presents

summary and preliminary reports on work in progress; and Basic Radio Propagation Predictions,

which provides data for determining the best frequencies to use for radio communications throughout

the world. There are also five series of nonperiodical publications: The Applied Mathematics

Series, Circulars, Handbooks, Building Materials and Structures Reports, and Miscellaneous

Publications.

Information on the Bureau’s publications can be found in NBS Circular 460, Publications of

the National Bureau of Standards ($1.25) and its Supplement ($0.75), available from the Superintendent of Documents, Government Printing Office. Inquiries regarding the Bureau’s reports and

publications should be addressed to the Office of Scientific Publications, National Bureau of

Standards, Washington 25, D. C.
NBS