Embed Size (px)
Citation preview
ASHRAE Research Project Report RP-394
A Study to Determine Methods for Designing Radiant Heating and Cooling Systems
Approval: May 1987
Contractor: University of Missouri-Rolla Rolla, MO 65401 Principal Investigator: Ronald Howell Authors: N/A Author Affiliations, Sponsoring Committee: TC 6.5, Radiant Space Heating and Cooling
Co-Sponsoring Committee: N/A
Co-Sponsoring Organizations: N/A
©2012 ASHRAE www.ashrae.org. This material may not be copied nor distributed in either paper or digital form without ASHRAE’s permission. Requests for this report should be directed to the ASHRAE Manager of Research and Technical Services.
FINAL REPORT
ASHRAE RP - 394
A STUDY TO DETERMINE METHODS FOR
DESIGNING RADIANT HEATING AND COOLING
SYSTEMS
Ronald H. Howell
Department of Mechanical & Aerospace Engineering
University of Missouri-Rolla
Rolla, MO 65401
May 1987
I ^«SWC>WS0C»ETV<S^"GHT
TABLE OF CONTENTS
SUMMARY 1
1.0 - OBJECTIVES AND SCOPE 3
2.0 - INTRODUCTION 4
3.0 - DEFINITIONS AND TERMINOLOGY 16
3.1 - Radiation, Convection and Conduction 16 3.2 - Infrared Ranges 17 3.3 - Low, Medium, and High Temperature Radiant Sources 19 3.4 - Mean Radiant Temperature, MRT 20 3.5 - Radiant Temperature Asymmetry 20 3.6 - Operative Temperature 21 3.7 - Effective Radiant Flux 22 3.8 - Average Unheated Surface Temperature 22 3.9 - Comfort Conditions 22 3.10 - Design Heat Loss Values 23
3.10.1 - ASHRAE Standard Heat Loss (HLD) 23 3.10.2 - Actual Design Heat Loss (HLA) 24 3.10.3 - Conduction Design Heat Loss (HLC) 24 3.10.4 - Conduction Design Heat Loss with Room Air 24
Temperature Gradient (HLCG)
4.0 - BACKGROUND 25
4.1 - Relationship Between Radiant Heating/Cooling and Comfort 25 Conditions.
4.1.1 - Fanger's Comfort Equation 25 4.1.2 - Changes in Air Temperature with Changes in MRT for Equal 31
Comfort 4.1.3 - Asymmetric Radiation and Comfort 32
4.2. Descriptions of Common Types of Radiant Systems 34
4.2.1 - Hydronic Floor Panels 35 4.2.2 - Electric Floor Panels 35 4.2.3 - Air Floors 36 4.2.4 - Hydronic Wall Panels 36 4.2.5 - Electric Wall Panels 37 4.2.6 - Hydronic Ceiling Panels (Metals or Plaster) 37 4.2.7 - Electric Ceiling Panels 38 4.2.8 - Miscellaneous Electric Systems 39 4.2.9 - Gas-Fired Radiant Ceramic Surface Infrared Units 4.2.10 - Gas-Fired Radiant Tube Infrared Units
39 40
4.2.11 - Electric Infrared Units 40
5.0 - CALCULATION OF DESIGN HEATING LOADS 43
5.1 - Standard ASHRAE Design Procedure 43
5.1.1 - Design Inside Air Temperature 45 5.1.2 - Room Temperature Gradients 45 5.1.3 - Wall, Ceiling, Floor Convection Coefficients 46
5.2 - Development of a Design Heat Loss Procedure for Radiant Systems 47
5.2.1 - Heat Balance on Room Surfaces 50
5.2.1.1 - qr - Radiant Exchange Rate 51
5.2.1.2 - qcv - Convective Heat Transfer 51 5.2.1.3 - qCQ< - Conductive Heat Transfer 53
5.2.2 - Heat Balance on Complete Room 54 5.2.3 - Comfort Equations 55 5.2.4 - Other Parameters Evaluated 56
5.3 - Comparison of Calculated Design Radiant Loads With the Standard 59
ASHRAE Design Load Calculation
5.4 - Test Case Calculation 60
5.5 - Radiant Panel Heating Systems Calculations 69
5.5.1 - Single Panel Radiant Heating Cases 69 5.5.2 - Effect Due to Infiltration for Radiant Panels 79 5.5.3 - Effect of Glass Distribution 79 5.5.4 - Changes in Wall, Floor, and Ceiling U-Factors 86 5.5.5 - Effect of Changes in Room Length and Width 91 5.5.6 - Changes in Room Height 91 5.5.7 - Changes in Outside Design Temperature 91 5.5.8 - Changes in Number of Panels 96 5.5.9 - Perimeter Panel System 96
5.6 - Comparison of Forced Air and Radiant Ceiling Panels 99 5.7 - Heated Floor Cases 99 5.8 - Infrared Heating Cases 109 5.9- U-Tube Infrared Cases 121 5.10 - Summary of Design Heating Calculations 127
6.0- DESIGN PROCEDURES 133
6.1 - Radiant Ceiling Panel Heating Systems 135 6.2 - Radiant Ceiling Panel Cooling Systems 136 6.3 - Heated Floor Systems 139 6.4 - High and Medium Temperature (Infrared Systems) 141 6.5 - Other Design Procedures 143
7.0 - SUMMARY OF MANUFACTURER SURVEY 145
8.0 - SYSTEM DYNAMICS 146
9.0 - RESEARCH NEEDS 147
9.1 - Convection Coefficients 147
9.2 - Air Temperature Stratification 147
'9.3 - Surface Emissivities 9.4 - Comfort During Radiant Temperature Asymmetry 9.5 - Radiant System Dynamics 9.6 - Heated Floor Systems
10.0 - REFERENCES
APPENDIX A - BIBLIOGRAPHY
APPENDIX B - ANNOTATED BIBLIOGRAPHY
147 147 148 148
149
A-l
B-l
A; B; c: D: E: F; G; H) i] J] K: L; M; N:
PENDIX C - LIS
Load Analysis and Modeling Convection Coefficients General Comfort Conditions
1 Thermal Comfort-Radiant 1 Floor Panels 1 Panel Heating and Cooling
Infrared Heating Design Procedures Energy Consumption
) Transient Effects Instruments
> Controls I Spot Heating and Cooling
5TING OF COMPUTER PROGRAM
APPENDIX D
Program Listing
Data Input File Listing
REPRODUCTION OF CHAPTER 8 FROM 1984 ASHRAE SYSTEMS HANDBOOK
C-l
C-2
D-l
SUMMARY
The goal of this study was to obtain design data and relevant manufactur
ers data concerning the design procedures for radiant heating and cooling sys
tems. A comprehensive literature search was conducted which resulted in an
annotated bibliography with over 250 entries. This bibliography was sub
divided into the following sections: load analysis and modeling, convection
coefficients, comfort conditions, radiant thermal comfort, floor panels, panel
heating and cooling, infrared heating, design procedures, energy consumption,
transient effects, controls, and spot heating and cooling.
The manufacturers survey resulted in identifying three commonly used cate
gories of radiant heating/cooling surface temperature ranges. The low surface
temperature range is 8O0F to 200oF for heating and 50oF to 70oF for cooling.
The medium surface temperature range is from 700 to HOOoF and the high
surface temperature range is from 1200oF to 2000oF. These surface temperature
ranges identify the four commonly used systems for radiant heating and cool
ing: ceiling panel heating and cooling and floor heated panels operate in the
low temperature range, U-tube infrared units operate in the medium temperature
range, and modular gas-fired or electric infrared units operator in the high
temperature range.
Analysis of the above information indicated that the only reliable or
appropriate design consideration would involve looking at the surface-to-air
design process and not the means which is used to obtain the heated surface
temperature. There are many variations or schemes used to obtain appropriate
surface temperatures and it was not the object of this study to evaluate all
of these schemes. Each manufacturer or designer has their unique method for
obtaining a specific surface temperature. Descriptions and applications are
provided for eleven of the most common configurations. These are: hydronic
floor panels, electric floor panels, air floors, hydronic wall panels, elec-
-1-
trie wall panels, hydronic ceiling panels, electric ceiling panels, gas-fired
radiant ceramic tube surface infrared units, gas-fired radiant ceramic tube
surface infrared units, gas-fired radiant tube infrared units, electric
infrared units, and miscellaneous electric systems.
A computerized technique was developed to relate heater surface tempera
ture to the space heating requirements while maintaining the Fanger comfort
constraints. For each variation, the required area of heater surface was
calculated and the actual design heat loss for the radiant heating system was
calculated and compared to the ASHRAE standard design procedure. Calculations
were made for the four types of radiant systems (ceiling panel heating and
cooling, heated floor panels, U-tube infrared, and modular infrared) for
typical ranges of many of the variables. The variables considered were:
U-factors, quantity of glass, heater surface temperature, surface emissivi-
ties, convection coefficients, outside air design temperature, room size,
ceiling height, infiltration rate, number of heating surfaces, heater place
ment, and use of reflectors or deflectors on infrared units.
The only variable which was found to have a significant effect on the
difference between the actual design heat loss and the ASHRAE standard heat
loss was the infiltration rate. The percent difference in these two design
heating loads varied from -4% at 0.5 ACH to -16% at 4 ACH, The actual design
heat loss is less than the ASHRAE standard design heat loss.
Design methods considering techniques for calculating loads, sizing
equipment and positioning equipment are presented for each of the common types
of radiant heating systems. The design procedure for radiant cooling which is
presented in the ASHRAE Systems Handbook was found to be adequate and is
recommended for use.
-2-
1.0 - OBJECTIVES AND SCOPE
The goal of this project was to obtain a body of accurate and relevant
data on methods of designing radiant heating and cooling systems. The data
includes methods of calculating loads, sizing equipment, and positioning
equipment.
The study has focused on identifying all significant types of radiant
heating and cooling systems by means of a literature search and analysis of
appropriate available data and technical material. From this material, a
procedure for designing radiant heating and cooling systems has been
developed. This procedure includes methods of calculating loads, sizing
equipment, and positioning equipment.
A major effort of the project has been the preparation of an annotated
bibliography of published sources of information for radiant heating/cooling
systems. As a result of the preparation of this annotated bibliography,
additional research needed in order to improve the recommended methods for
calculating loads, sizing equipment, positioning equipment, and system
dynamics has been provided.
-3-
2.0 - INTRODUCTION
Convective and radiant heating and cooling systems have been used for many
years in providing comfort systems in rooms occupied by people and/or materi
als. These two types of systems produce different comfort environments due to
their nature of heat delivery or removal, and thus there is no fundamental
reason to expect them to be sized by the same technique or to require the same
energy to produce identical levels of comfort. Proponents of radiant heating
systems assert that these types of systems offer the potential for reduced
heating unit sizes and reduced energy consumption. They claim that the room
may be operated at a lower air temperature than if it is heated by a convec
tive system because the radiant heat from the heater falls directly on the
occupants, producing comfort conditions. However, there is also the opposite
factor, that radiant heating systems produce higher floor, wall and glass tem
peratures due to the radiant heaters heating these surfaces and not the air,
and thus producing greater heat losses to the surroundings.
The thermal environment within a room and its rate of heat loss are deter
mined by the configuration and structural materials used in the walls, floor
and ceiling; the amount of infiltration air forced through the room; and the
nature of the heat suppliers. A convective type of system such as shown in
Figure 1 produces an environment where the air temperature is greater than the
mean radiant temperature in the space. A radiant heating system such as
shown in Figure 2, on the other hand, produces an environment in which the
mean radiant temperature (or "average" room surface temperature) is higher
than the air temperature. For this reason, the infiltration air losses will
be greater in convective than radiant heating systems. Convective types of
systems using fans for delivering heated air which cause slight air pressure
differences will tend to increase the air infiltration loss. On the other
hand, radiant heating systems increase the room surface temperature causing
-4-
FIGURE 1. CONVECTIVE TYPE OF HEATING SYSTEM
1 m J n L
Heating And/Or Cooling Panel
< EflH] I3LT15.
FIGURE 2. RADIANT CEILING PANEL HEATING/COOLING SYSTEM
-5-
increased heat loss to the surroundings. Radiant heating systems also have the
advantage of increasing the mean radiant temperature to which occupants are
exposed and thereby allowing comfort at lower air temperatures.
There are two fundamentally different characteristics to be considered.
First is the concept of sizing of radiant heating systems and second is esti
mating the energy required by radiant systems for providing comfort conditions
over a heating season. For sizing, design calculations are made to indicate
what is the expected maximum rate of total heat delivery which is to be
expected from the heater. Along with the total size of the heating system is
the positioning of the individual heating units so that they provide uniform
comfort conditions throughout the space.. In Chapter 25 of the ASHRAE Handbook
of Fundamentals (1) a procedure is presented for determining the design heating
load for a structure. The fundamental objective of this project has been to
determine if this ASHRAE Design Heating Load Procedure is applicable to
radiant heating systems.
The estimation of the energy used by a radiant heating system over a heat
ing season is a separate and more complicated problem. System dynamics and
thermal storage characteristics of the structure are important factors in ans
wering this question. It is questionable whether some of the simpler proce
dures presented in Chapter 28 of the ASHRAE Handbook of Fundamentals (HBF) (1)
such as degree day method, full load hours, or BIN method are applicable to
radiant heating systems. This project does not address the energy requirement
calculation.
There are three general categories of radiant heating/cooling systems, and
these can be identified by the temperature range in which they operate. One
category is that of panel heating and cooling systems where the surface tem
perature can be called low and is in the range of 80 F to 200 F for heating
and 50 F to 70 F for cooling. In these systems, the surface temperature is
-6-
controlled in order to vary the quantity of heat being delivered or absorbed.
The controlled temperature surfaces may be in the floor, walls or ceiling, and
the temperature is maintained by circulating water, air, or electric current.
Chapter 8 in Reference 2 presents a good description of these types of radiant
heating or cooling systems as well as some design procedures for installing
the systems. Figure 2 illustrates a ceiling panel system and Figure 3 depicts
a heated floor type of system.
The second type of radiant system comprises the medium temperature range
units which operate from about 700 F. to 1100 F and consist of radiant tubes
through which the products of combustion from a gas burner are circulated and
then exhausted to the outside. Descriptions of these types of units are given
in Chapter 30 in Reference 3. The use of these types of radiant heaters is
presented in Chapter 18 in Ref. 2. These units come in integral lengths which
can be placed in specific patterns or in U-tube shaped units of different
lengths. They have the advantage of exhausting the exhaust products to the
outdoors rather than inside the structure. These are illustrated in Figure 4.
The third type of radiant unit is the modular high temperature infrared
unit operating in the range of 1200 F to 2000 F surface temperature. They
consist of gas or electric operated units placed at various locations
throughout the space and are generally used for spot heating applications, or
in many cases, for full area comfort heating. Use of these units is depicted
in Figure 5. The gas fired units have the disadvantage of discharging the
products of combustion inside the conditioned space. Descriptions of these
units are given in Chapter 30 in Reference 3, and the application of the units
is discussed in Chapter 18 in Reference 2.
One of the advantages, and in some cases a disadvantage, is the
maintenance of comfort conditions when using radiant heating. The advantage
occurs when the units are properly sized and located, providing a higher mean
-7-
I 00 I
FIGURE 3 HEATED FLOOR TYPE OF RADIANT HEATING SYSTEM FOR BEDROOM AND BATH
Tube Heaters Medium Temperature
U - Tube Heaters Medium Temperature
feaJ
FIGURE 4 MEDIUM TEMPERATURE RANGE RADIANT TUBE TYPE OF HEATERS
-9-
CO u CD
4-1 Cfl CD
a CU U K) U
Q O
<
o
IT)
UJ
a: cs
- 1 0 -
radiant temperature for the occupants, which then permits a lower air
temperature for equal comfort conditions. The disadvantage can occur if the
radiant heat is concentrated to such a condition that the asymmetric
temperature felt by the occupant is such that discomfort occurs in the space.
Any design procedure that is specified must account for maintaining comfort
and not creating severe asymmetric temperature conditions. Typically, by
satisfying the Fanger comfort equations [4] and limiting the asymmetric
temperature to 9 F, no discomfort should be experienced by the occupants.
Another advantage claimed for radiant heating systems is the negligible
air temperature gradient experienced by spaces using radiant sources rather
than convective sources for heating. This occurs due to the fact that radiant
systems provide higher surface temperatures than experienced in convective
systems with very little air motion resulting in a more uniform air temperature
distribution. Convective heating systems will generally have air temperature
gradients due to the higher temperature of the air brought into the space for
heating purposes with a resultant higher air temperature at the ceiling than
at the floor.
In Figure 6, a schematic is given of room air temperature gradients for
forced air heating, heated ceiling panels, and heat floors. These schematics
were prepared from data such as that shown in Figure 7 which are some results
from 1953 data at the ASHVE Laboratory in Cleveland, Ohio. Articles contain
ing this type of data are listed under G-Panel Heating and Cooling in the
ANNOTATED BIBLIOGRAPHY. For convection or forced air heating systems room,
air temperature gradients from 1/2 to 2°F per foot could be experienced
depending on room size, insulation levels, and the air distribution system
design and operation. For radiant heated and cooled rooms, the room air tem
perature gradient is negligible. It should be kept in mind, however, that
application of the available room air temperature gradient requires careful
-11-
61° 68° 75° 82
61° 68° 75° 82
61° 68° 75° 82c
FIGURE 6. SCHEMATIC OF AIR TEMPERATURE GRADIENTS FOR FORCED AIR HEATING, RADIANT CEILING HEATING, AND RADIANT FLOOR HEATING
-12-
UJ I
cu fa
4-1 .C 60
•H ID
a
I T
Heated Floor Floor a t 85FJ AUST = 65 F
65 70 75
\
\
T — r
Ceiling Panel -Cooling,Ceiling at 65F,
AUST = 85 F
80 85
Ceiling
Floor
ROOM AIR TEMPERATURE, °F
FIGURE 7 MEASURED ROOM AIR TEMPERATURES FOR RADIANT HEATING AND COOLING SITUATIONS. ASHVE LABORATO RY DATA, JULY 1953
consideration of all of the parameters under which the data was measured and
collected. All data cannot arbitrarily be applied to any situation.
With a temperature gradient in the room, the infiltration heat loss is
greater than when a gradient does not exist. Because of this, it is expected
that convection types of heating systems will have larger design infiltration
losses than radiant systems. During the heating mode, infiltration air will
enter at the bottom of the space and exfiltration will occur at the top of the
space. This exfiltrated air will be at a higher temperature in the convection
type systems thereby creating a larger heat loss than experienced in radiant
types of systems. This is another expected benefit of radiant heating types
of systems.
Radiant heating systems are used in many types of applications such as
offices, hospitals, homes, warehouses, and manufacturing or industrial situa
tions . For hospitals and offices ceiling panel radiant systems are typically
used. For homes, offices, and warehouses, very often radiant floor panels are
used. The medium and high temperature infrared systems are generally found in
warehouses, manufacturing, and industrial situations. These general types of
applications are not meant to be restrictive since each application should be
addressed individually by weighing the advantages and disadvantages of each
type of system. Additional details on applicability are given later on in
this report in Section 4.2.
What has been investigated in this project is a system design procedure
for radiant heating and cooling. The evaluation of the energy requirements
for radiant heating systems has not been considered. The system design proce
dure involves the estimation of the design heating or cooling load, the selec
tion of the type of radiant system to be used (ceiling panels, floor panels,
U-tube modular units or infrared modular units) which is partially based on
the allowable heater surface temperatures for the application considered, and
-14-
the positioning of the heaters in the space. In addition, a literature survey
has been conducted and is included as Appendix A and an Annotated Bibliography
is presented as Appendix B.
&-.: -15-
3.0 - DEFINITIONS AND TERMINOLOGY
This section of the report presents some definitions and terminology that
are used throughout the remainder of the report. Some of the terminology
concepts are important because names have been commonly associated with
various phenomena or items which are restrictive when they need not be.
3.1 - Radiation. Convection and Conduction
3.1.1 - Radiation.
The radiation energy transfer process is the consequence of
energy carrying electromagnetic waves emitted by atoms and
molecules resulting from changes in their energy content.
Amount and characteristics of radiant energy emitted by a
quantity of material depends on the nature of the material, its
microscopic arrangement and its absolute temperature. Although
rate of energy emission is independent of the surroundings, net
energy transfer rate depends on the temperatures and spatial
relationships of the surface and its surroundings and can be
expressed for two black surfaces as:
qi_2 = orAF (T4 - T4) (1)
where:
a = Stefan-Boltzman Constant, 0.1713x10"8 Btuh/ft2 R4
A = area of one surface
F = geometrical factor relating shape and orientation of the
surfaces
^1-2 n e t exchange of radiant heat between the two surfaces
3.1.2 - Convection.
Convection involves the transfer of heat by mixing one portion
of fluid with another. The motion of the fluid may be entirely
the result of differences of density resulting from the
-16-
temperature difference, as in natural convection, or the motion
may be produced by mechanical means such in forced
convection.[5] The general equation for convection heat transfer
is
q = h A (T-L - T2) (2)
where:
q = exchange of convective heat between two surfaces
h = convection heat transfer coefficient
A = surface area
3.1.3 - Conduction.
Conduction in a homogeneous opaque solid is the transfer of heat
from one part to another under the influence of a temperature
gradient without appreciable displacement of the particles.
Conduction involves the transfer of kinetic energy from one
molecule to an adjacent molecule. For steady state
one-dimensional conduction heat transfer, the following equation
applies.
K A (TX - T2) T]_ - T2
q = C (T]_ - T2) (3)
X R
where:
q = exchange of heat by conduction from one surface to another
K = thermal conductivity of the material
X = thickness of the material
A = area perpendicular to the flow of heat
R = thermal resistance of the material
C = thermal conductance of the material
3.2 - Infrared Ranses
The thermal radiation emitted by a surface encompasses a range
-17-
of wavelengths. An example of this is shown in Figure 8
where the magnitude of the radiation varies with wavelength.
c o
2 c o o ~ w> ro vt
Wavelength
Figure 8 Spectral Distribution of Thermal Radiation [6].
Emitted radiation consists of a continuous, nonuniform
distribution of monochromatic (single-wavelength) components.
The magnitude of the radiation at any wavelength and the
spectral distribution vary with the nature and temperature of
the emitting surface. It is also important to understand that a
surface may emit preferentially in certain directions creating a
directional distribution of emitted radiation as illustrated
in Figure 9 [6].
Directional distribution
Figure 9 Directional Distribution of Emitted Radiation [6],
The relationship between temperature and wavelength for the peak
radiated energy is illustrated in Figure 10. Also
delineated there are the various wavelength regions of thermal
radiation.
-18-
Spectral distribution
WAVE LENGTH (IN MICRONS)
30
20
IS
10
8
6 5
4
3
a •
i.s
i .8
FAR
INFRA
RED
MIDDLE
INFRA
RED
NEAR INFRARED
VISIBLE
LIGHT
ULTRA VIOLET
-O'F
SOO'F
-|0O0oF
-3OO0°F
6000°F
Figure 10 Temperature - Wavelength Relationship
3.3
Radiant heaters used in comfort heating applications operate in
what is generally classified as the far-and middle-infrared
region: 2 to 8 microns wavelength and 85 F to 2000 F surface
temperature.
Low. Medium and High Temperature Radiant Sources
Commonly available radiant heating systems for HVAC applications
exist in three general temperature ranges. The low ranpe is
from 80 F to 200 F and consists of hydronically or electrically
-19-
heated panels or surfaces usually placed in the ceiling or walls
or as part of the floor: The medium range units operate between
700 F and 1100 F and usually consist of vented gas-fired tubes
placed near the ceiling or roof in a structure. The hiph
temperature radiant units operate in the range of 1200 F to 2000
F and usually are nonvented gas-fired units or electrically
heated units.
3.4 - Mean Radiant Temperature (MRT)
The temperature of a theoretically conceived isothermal black
enclosure in which an occupant would exchange the same amount
of heat by radiation as in an actual nonuniform surface
temperature environment. The MRT can be determined from the
following equation.
MRT4 = T Fp..! + T4Fp_2 + --- + T4 Fp.n (4)
where:
T^, T2, -- Tn = surface temperatures surrounding the occupant
in a room
*p-l» *p-2» "" Fp-n = geometrical factor relating shape and
orientation between a person and
the surrounding surfaces.
3.5 - Radiant Temperature Asymmetry
The difference between the plane radiant temperature of the two
opposite sides of a small plane element. This is a new term
used to describe the asymmetry of a radiant environment. For
vertical heating or cooling panels it refers to a small vertical
element 3.6 feet above the floor at the person's position. For
-20-
horizontal heating or cooling surfaces it refers to a small
horizontal element at the same position. The plane radiant
temperature (T_r) can be calculated [7] from
Tpr = Fe-lT4 + Fe.2T4 + --- + Fe.n T
4 (5)
where:
T^, T2, -- Tn = temperatures of surfaces surrounding the
element
Fe_^, Fe_2, -- Fe_n = geometrical factors from the plane
element to the specified surface.
Radiant temperature asymmetry is then defined as
Tprl " Tpr2 <6>
where the subscript 1 refers to one side of the plane element
and the subscript 2 refers to the opposite side.
3.6 - Operative Temperature (tQ)
The uniform temperature of a theoretically conceived enclosure
in which an occupant would exchange the same amount of heat
by radiation and convection as in the actual nonuniform
surface temperature environment. It is given
by the following equation.
(hc x ta) + (hr x MRT)
tG = = (7) hc + hr
where: hc <•= the convective heat transfer coefficient for the occupant [1]
hr = the radiant heat transfer coefficient for the occupant [1]
ta = ambient air temperature
-21-
3.7 - Effective Radiant Flux (ERF)
This is the net radiant >heat exchanged per unit area
at the ambient temperature ta between an occupant represented
by a hypothetical surface and all enclosing surfaces and
directional heat sources and sinks. ERF is the net radiant
energy per unit area received by the occupant from all surfaces
and sources whose temperatures differ from the ambient air ta.
It is given by the following equation [1].
ERF - hr (MRT-ta) - hc (tQ - ta) (8)
3.8 - Average Unheated Surface Temperature (AUST)
The area-weighted temperature of the surfaces in a room which
are not acting as suppliers of external heat to the room. It is
given by
AlTi + A2T2 + -- + ANTN
AUST (9)
Al + A2 + •• + AN
where:
Al> A2> "" n = a r e a s °f surfaces not supplying external heat
to a room
T^, T2, -- Tn «= temperature of surfaces not supplying external
heat to a room
3.9 - Comfort Conditions
Several parameters are used to identify when a human occupant is
exposed to what are commonly called comfort conditions. The
frequently used parameters are:
ta = ambient air temperature
RH = relative humidity
-22-
V — relative air velocity
MRT = mean radiant temperature
There are two other items which affect the values of these
parameters when comfort is concerned.
MET = the metabolic rate for the occupant which is a
function of their activity level
CLO = the thermal resistance of the clothing being
worn by the occupant
The above variables have been combined into a set of equations
• [4] which can be solved yielding values for the comfort range.
In addition, sets of charts have been developed [1] which are
applicable for the most common range of conditions. These are
discussed later in Section 5.2.3.
3.10 - Design Heat Loss Values
There are four design heat loss values which will be calculated
during this analysis. Three of these are different from what is
taken as the "standard ASHRAE design heat loss". These are
defined as follows.
3.10.1 - ASHRAE Standard Heat Loss (HLD)
HLD - SiUjAi (75 - toa) +1.1 CFMI (75 - toa) (10)
where:
V-i '«=» design U value for each component of the room (walls,
glass, ceiling, floor)
AJL = area of each of these individual components
-23-
toa «= outside design air temperature
CFMI = estimated design value of infiltration air
3.10.2 - Actual Design Heat Loss (HLA)
HLA - S Ut k± (ta - toa) +1.1 CFMI <ta - toa) (11)
where:
ta = design room air temperature based on comfort conditions
being met at the center of the room
3.10.3 - Conduction Design Heat Loss (HLC)
HLC ~ S ± Ct A£ (tsi - toa) + 1.1 CFMI <ta - toa) (12)
where:
C^ = The conductance for the room component (wall, glass,
ceiling, floor) from the inside surface to the outside
air. It excludes the inside convection coefficient.
tsi = *-he surface temperature of the room component
3.10.4 - Conduction Design Heat Loss with Room Air Temperature
Gradient (HLCG)
HLCG - S ± Ci Ai (tsi - toa) +1.1 CFMI (tag - toa) (13)
where:
ta„ = The room air temperature at the ceiling level where
exfiltration occurs. It is evaluated using a specified
air temperature gradient based on a reference height of
5 ft.
-24-
4.0 - BACKGROUND
4.1 - Relationship Between Radiant Heating/Cooling and Comfort Conditions
When studying the use and performance of radiant heating and cooling sys
tems it is important to understand the inter-relation between the various par
ameters of comfort and their influence on the sizing and location of radiant
units. Many comfort studies have been done over the years and many of these
are presented in the ANNOTATED BIBLIOGRAPHY in Appendix B. The one which is
most suited for this investigation and which has been widely accepted as pro
viding meaningful results is the work by Fanger [4].
4.1.1 - Fanger's Comfort Equations Fanger's study generalizes the physio
logical basis of comfort and allows comfort for most activity levels to be pre
dicted analytically in terms of the environmental parameters presented in
3.0 - DEFINITIONS AND TERMINOLOGY. The Fanger Comfort Equations are based on
a logically derived heat balance equation for the occupants thermal equili
brium and on observations that during a state of comfort defined by neutral
temperature sensation, a unique relation exists between level of activity,
skin temperature, and evaporative loss from the body [1]. The comfort equation
contains the following grouped variables:
Ic]_t fcl " A function of the type of clothing
M —-,-IJ', v - A function of the type of activity ADu
v, ta, Pa, tmrt - Environment variables
Fanger's Comfort Equation is represented by the following two equations.
-25-
M M (l-n) - 0.35 [43.0 - 0.61 (l-i;) - Pa] -
ADu ADu
M M 0.42 [ (l-»7) - 50.0 ] - 0.0023 (44 - Pa) -
ADu ADu
M 0.0014 (34.0 - ta) -
ADu
3.4 x 10-8 fcl [tcl + 273)4 - (tnrt + 273)4]
+ fcl hc (fccl - ta) (14)
M M tcl = 35.7 - 0.32 (1-IJ) - 0.18 I c l [ (l-q) -
ADu ADu
M M 0.35 [43.0 - 0.61 (1-r;) - PJ - 0.42 [ (1-r;) - 50]
ADu ADu
M M - 0.0023 (44.0 - Pa) -0.0014 (34.0 - ta)]
ADu ADu (15>
where:
M = metabolic rate
ADu = DuBois body surface area
rj = external mechanical efficiency
Pa = water vapor pressure of ambient air
ta = ambient air temperature
fc1 = the ratio of the surface of the clothed body to the surface
area of the nude body
tc^ = clothing surface temperature
tj -j- = mean radiant temperature
-26-
h c = convective heat transfer coefficient, generally a function
of air velocity
IC;L = thermal resistance of clothing
The ASHRAE [1] comfort envelope is shown in Figure 11 but
applies only for sedentary (1 Met =58.2 W/m2) and slightly active (<1.2 MET),
normally clothed persons at low relative air velocities when the MRT equals
the air temperature. Figures 12, 13 and 14 depict comfort lines or curves
20
15
S
0
-5
-10
• 7 / |
:
-
/
i %
/
1 /'
/
-
f?
-
~
-
20 25 30 OPERATIVE TEMPERATURE. "C
Figure 11. Acceptable Ranges of Operative Temperature and Humidity
for Persons Clothed in Typical Summer and Winter Clothing
at Light to Sedentary Activity [1].
through various combinations of variables in order to create comfort for
constant values of some of the other variables. The quantitative influence of
clothing, activity and environmental parameters given by Fanger's Comfort
Equation and shown in these figures or comfort charts has been confirmed by
studies of the individual parameters [8-13].
-27-
SEDENTAHY I mel
AIR TEMPERATURE = MEAN RADIANT TEMPERATURE. 'C AIR TEMPERATURE = MEAN RADIANT TEMPERATURE. "C
MEDIUM ACTIVITY 2 mel
LIGHT CLOTHING
I c | = 0.5clo
MEDIUM ACTIVITY 2 mel
MEDIUM CLOTHING
I e i= I .Oclo
10 20
1—I—|—I—p
25
AIR TEMPERATURE = MEAN RADIANT TEMPERATURE. *C AIR TEMPERATURE = MEAN RADIANT TEMPERATURE, "C
• -
3 0 -
" -
2 5 -
O
uf
Ul a. I
Si g is-CO
fe 1 0 -
-•
5 -
.;
HIGH ACTIVITY 3 met
LIGHT CLOTHING
I c I = 0 . 5 c l o
J^T° ^Ao y*
J^>^ i - T [ i
T O lu»
3° ^S* \f ^ Ysk s>
Jta 13_
X>^ • * " * ^
1 1 1 1
r&S JS S ••tir S^ /
jr$^\y^ S rA'S' '
^^ -g^^ - - ^ o 2 i ^ - - ^ '
1 1 1 1 1 1 1 1 1 I 1
HIGH ACTIVITY 3 met
MEDIUM CLOTHING
I c | S | 0 C l O
10 IS 2 0 25 3 0
AIR TEMPERATURE = MEAN RADIANT TEMPERATURE. *C
- I — P I — | — P
2 1
AIR TEMPERATURE - MEAN RADIANT TEMPERATURE. "C
FIGURE 12 COMBINED INFLUENCE OF HUMIDITY AND AMBIENT TEMPERATURE [ 1 ]
- 2 8 -
SEDENTARY 1 met
» c l '
•
• • . .
0.5 do
ft/1
....
^
IING
•
V ^
^ o
/ S /
\
v l A V I *
....
•
/ / /
V \ ....
/ /
10 IS 2 0 2 5 3 0 35 AIR TEMPERATURE. "C
10 15 20 25 30 35 AIR TEMPERATURE. *C
V N
\
....
^ s
\
- A •
< V& V
V
. . I I . . . I
/ / /
/ /
* / / /
MEDIUM ACTIVITY 2 mel
, MEDIUM CLOTHING
\ I d ' I O d o
i ! \ 1 .... ....
• • . •
5 IO IS 20 25 30 35 AIR TEMPERATURE. -C
I I I I—I I I I 10 IS 20 25 3 0 35
AIR TEMPERATURE, *C
FIGURE 13 COMBINED INFLUENCE OF MEAN RADIANT TEMPERATURE AND AIR TEMPERATURE [ 1 ]
- 2 9 -
1.4-
1.2-
>. IO-*-u §0B-
> | 0 6 -
LIGHT CLOTHING IcfOSeto IC|*I.I
1 1 1 1
'§^
1 1 1 I
*"/
7 J
f (J i i i i
I i ? 1 / J / 1 / 7
7 y / / , / , i i i i
i i \ \ i \ 1 / */
of h i /
J / i i i i i
10 13 20 23 AIR TEMPERATURE o MEAN RADIANT TEMPERATURE.'C
I I 1 I | I I I I | I IO 13 20 23 90
AIR TEMPERATURE " MEAHRADIANTTEMPERATURE, *C
Figure 14. Combined Influence of Air Velocity and Ambient Temperature [1].
In Figure 12 the comfort lines are curves through different combinations of
ambient temperature and humidity that provide thermal comfort. The six charts
apply to six different combinations of activity and clothing, where the air
temperature equals mean radiant temperature. In Figure 13, the comfort lines
are curves through different combinations of mean radiant temperature and air
temperature that provide thermal comfort. The six charts apply to six
-30-
different combinations of activity and clothing at 50 percent relative
humidity. In Figure 14, the comfort lines corresponding to five different
activity levels are curves through different combinations of relative air
velocity and ambient temperature which provide optimal thermal comfort. The
two charts apply for persons wearing 0.5 and 1.0 clo with the relative
humidity at 50 percent.
For practical application of these charts, first estimate the activity
level and clothing quantity, taking room use into account. The combination of
the four environmental parameters (ta, MRT, RH and velocity) that provide
thermal comfort can then be found from Figs. 12-14.
4.1.2 - Changes in Air Temperature with Changes in MRT for Equal Comfort Some
comfort condition examples for radiant heating applications have been worked
out in order to illustrate how the ambient air temperature should be changed
for different mean radiant temperatures at various activity and clothing lev
els. These are illustrated in Table 1. In this table, the marked values are
when the mean radiant temperature is equal to the air temperature. The strong
effect of activity and clothing level can be illustrated by looking at the
temperatures when they are equal. This value ranges from 80.6 F for sedentary
activity and light clothing to 61.2 F for medium activity and clothing. For
sedentary activity and medium clothing these temperatures are 76.5 F which is
a common situation in convective heating systems. Notice that the air temper
ature should be lowered to 72 F for comfort if the MRT = 80.6 F.
For the other situation with medium activity and light clothing the temperatures
are equal at 69.8 F. In this situation, if the MRT = 77 F the room air tem
perature needs to be at 63.5 F for comfort. This illustrates the large
changes in room air temperature which are needed in order to provide comfort
for changes in MRT, activity level and clothing level.
-31-
TABLE 1 Comparison of Room Air Temperature and MRT for Comfort Conditions at
Different Activity and Clothing Levels with 30% Relative Humidity and
0.2 m/s Relative Velocity
Sedentary Activity
Light Clothing
MRT, F
68
71.6
75.2
78.8
• 80.6
82.4
86.0
89.6
93.2
96.8
ta F
87.8
85.1
83.3
81.5 •
80.6
79.7
75.6
72.9
71.6
68.4
Sedentary Activity
Medium Clothing
MRT, F
60.8
62.6
64.4
66.2
68
69.8
71.6
73.4 •
76.5
77
80.6
84.2
86
ta. F
82.4
81.5
80.6
79.9
79
78.1
77
75.2 •
76.5
74.7
72
69.3
68.4
Medium Activity
Light Clothing
MRT, F
60.8
62.6
64.4
66.2
68
• 69.8
71.6
73.4
75.2
77
78.8
ta. F
75
74.5
72.3
71.4
70.2
• 69.8
67.8
66.0
64.4
63.5
62.8
Medium Activity
Medium Clothing
MRT, F
50
51.8
53.6
55.4
57.7
59.0 •
61.2
62.6
64.4
68.0
71.6
75.2
ta. F
67.6
66.2
65.5
63.9
63.0
62.6 •
61.2
59.9
57.6
56.3
53.1
50.7
4.1.3. Asymmetric Radiation and Comfort. Radiant temperature asymmetry
was defined in 3.0-DEFINITIONS AND TERMINOLOGY and recent work has suggested
some values for acceptable radiation asymmetry for comfort. A figure pre
sented by Fanger [14] is shown as Figure 15 delineating radiant temperature
asymmetry for heated ceilings and walls and for cooled walls and ceilings. If
10 percent dissatisfaction is acceptable, then the lowest acceptable radiant
-32-
temperature asymmetry is approximately 12 to 13 F. The ISO standard [15]
suggests 9 F.
•a ai
•H >H to
• r l 4-1 CO
as CO
• H
« n (U
o <u
100
60
40
20
10 -
5 4 3
Warm Wall
5 10 15 20 25 30 35
Radiant Temperature Asymmetry, °C
Figure 15 Percentage of People Expressing Discomfort Due to Asymmetric
Radiation [14]
Additional work done by Olesen and Nielsen [7] resulted in the values
shown in Figure 16 for vertical radiation cooling. This allows about a 25 F
(15K)
-33-
CD
nf CO CO
cu bO ca •U c CU
o
cu P4
60 40
20
10
6
4
2
1
-
• m
-
-
« --
-
:
•
m
0 2 4 6 8 10 12 14 16 18 20 22 24
Radiant Temperature Asymmetry, K
Figure 16 Percent Dissatisfied in Spot Cooling Due to Vertical Radiation [7]
radiation asymmetry for 10 percent of the occupants being dissatisfied.
Generally, about a 9 to 15 F radiant temperature asymmetry appears to be
acceptable. The true value of this quantity in actual applications is diffi
cult to establish since it depends on the type of radiant unit, its location
relative to the occupant, the location of furnishings in the room, the amount
of reradiation and the angle factor between the occupant and the radiant unit.
For these reasons, it has not been calculated in this study. Additional
research is required in this area.
4.2 - Description of Common Types of Radiant Systems
A survey of manufacturers and designers was conducted in order to identify
the commonly used types of radiant systems. From the results of this survey
the following descriptions were prepared. In Table 2, an applications matrix
was developed from the manufacturers information and the previously given
descriptions of the various types of radiant heating and cooling systems. Ten
-34-
types of radiant systems are compared and some of their general characteris
tics and typical applications are indicated. It should be kept in mind how
ever that there are other applications and characteristics which are not
identified here. It is not possible to list all conceivable applications and
operating traits for all radiant systems.
4.2.1 - Hydronic Floor Panels
This type of radiant heating system (see Figure 3) consists of pipes
imbedded in concrete floors with heated water being circulated through the
pipes. For comfort applications, the floor surface temperature is generally
limited to a maximum value of 85°F. Typically, plastic pipe is used, although
metal pipes or copper tubing can also be used. Hydronic floor systems provide
a uniform source of heating and require no mechanical air circulation unless
forced ventilation is required. They can be zoned for various types of load
situations and create a minimal amount of noise during operation.
Hydronic floor panel systems are well suited for applications where a
large change in load does not occur in a short time span. They are commonly
applied in residences, factories, warehouses, garages, and light commercial
structures. Their transient responses are slow due to the high thermal mass
of the concrete floor, although new lower mass floors have helped to alleviate
this problem.
4.2.2 - Electric Floor Panels
This type of radiant heating system (see Figure 3) consists of electrical
heating elements imbedded in concrete floors. For comfort applications, the
floor surface temperature is limited to a maximum value of 85°F. Electrically
heated floor systems provide a uniform source of heating and require no auxil
iary systems unless forced ventilation is required for the space. They can be
-35-
zoned for various types of load situations and create no noise during oper
ation. They are easily controlled with shielded thermostats, and portions can
be sequenced for more efficient energy usage.
Electric floor panel systems are well suited for applications where a
large change in load does not occur in a short time period. They are commonly
applied in residences, factories, warehouses, garages, and light commercial
structures. Their transient response is slow due to the high thermal mass of
the concrete floor, although new lower mass floors have helped to alleviate
this problem.
4.2.3 - Air floors
In air floor types of radiant heating systems, the heated air from a
furnace is circulated through passageways in the floor (wood construction or
concrete with imbedded tile) . For comfort applications where people are
present, the floor surface temperature is limited to a maximum value of 85°F.
Air heated floor systems provide a uniform source of heating and do require
mechanical circulation of heated air. This source of air can also be used for
providing ventilation if required. These systems can be zoned for various
types of load situations and normally create only a minimal amount of noise
during operation.
Air heated floor systems are well suited for applications where a large
change in load does not occur during a short time span. They are normally
controlled with shielded thermostats. They are commonly applied in
residences, warehouses, and light commercial structures.
4.2.4 - Hydronic Wall Panels
Hydronic wall panels can be modular metal panels with tubing connected to
the backside, or tubing connected to the wall surface and covered with
-36-
plaster. Hydronic wall panels are used in place of ceiling panels when the
panel location in the ceiling would interfere with lighting fixtures or some
type of required suspensions from the ceiling. These units have the same
characteristics and features as the hydronic ceiling panels. Generally more
heated panel area would be required than with ceiling panels. The surface
temperature would have to be limited if it can be contacted by people.
These units are applied in hospitals, office buildings, industrial plants
and sports facilities.
4.2.5 - Electric Wall Panels
Electric wall panels are available in various sizes for different types of
applications and are composed of heaters located between a wall surface
material and insulation on the back of the panel. Electric cable can also be
attached to the wall and covered with plaster to accomplish the same type of
heating. Electric wall panels would be used in lieu of ceiling panels when
the panel location in the ceiling would interfere with lighting fixtures or
required suspensions from the ceiling, or if the ceiling is too high for
practical application. These units have the same characteristics and features
as the electric ceiling panels. Generally more heated panel area will be
required than with ceiling panels due the wall location. The surface
temperature would also have to be limited if it can be contacted by people.
These units are applied in hospitals, office building, industrial plants
and sports facilities.
4.2.6 - Hydronic Ceiling Panels (Metal or Plaster1)
Hydronic ceiling panels (see Figure 2) can be made up of modular metal
panels laid in the ceiling, where tubes on the backside of the ceiling panels,
carry circulated water for heating or cooling or tubing attached to the ceil-
-37-
ing substructure covered with architectural plaster. These systems can pro
vide both heating and cooling and can use any source of energy since water is
the circulated heat transfer fluid. For heating, the panel surface tempera-
ture is generally in the range of 120°F to 180°F depending on ceiling height.
Hydronic ceiling panels provide a uniform source of heating or cooling and do
radiate to the floor and walls to maintain them at a comfortable temperature.
When cooling is done, the entering water temperature can be no lower than the
room air dew point, and the latent load in the space must be removed with a
separate system.
These units can be zoned for various types of load situations arid can be
used in conjunction with other types of heating/cooling systems. The metal
panel systems will respond quickly to load changes in the space.
These types of systems are generally applied in hospitals, office
buildings, schools, air terminals, industrial plants, and sports facilities.
4.2.7 - Electric Ceiling Panels
Electric ceiling panels (see Figure 2) are available in various sizes for
different applications and are made up of various types of heaters located
between a ceiling surface material and insulation on the back of the panel.
These are used for heating only applications and generally operate with sur
face temperatures between 120°F and 180°F. Electric ceiling panels provide a
uniform source of heating or cooling and radiate to the floor and walls main
taining them at comfortable temperatures. These units can be zoned and con
trolled sequentially in order to accommodate various load situations. They
can also be used in conjunction with other types of heating systems within a
building. Their transient response is generally good due to their low thermal
mass.
Electric radiant celing panels are applied in hospitals, office
-38-
buildings, schools, air terminals, industrial plants and sports facilities.
4.2.8 - Miscellaneous Electric Systems
There are several forms of unusual types of radiant heating systems
available. One form is an electric carpet that is composed of electrical
heating wires. There are also cloth wall hangings or drapes of similar
construction involving electrical heating wires.
These devices are generally used for retrofit types of applications or
where some additional heating is necessary. They are easily installed and do
not require construction changes within the building.
4.2.9 - Gas Fired Radiant Porous Refractory Surface Infrared Units
These radiant heating units (see Figure 5) supply a mixture of air and gas
through a porous refractory material, and combustion occurs at the surface of
the refractory burner. In some instances, a metallic grid is placed at the
surface to enhance performance of the unit. These units are unvented and
place the products of combustion (mainly carbon dioxide and water vapor) into
the space being heated. Units are available with different types of
reflectors or lens, which concentrate and direct the radiant energy into
suitable patterns. Units used for total building heating typically operate
with a surface temperature in the range of 1500 F (815°C) to 2000 F (1094°C)
and are self-contained and use shielded air-source thermostats for control.
They are most commonly used in occupancies where a large room volume is
present with high ceilings. Common applications include factories,
warehouses, assembly areas, aircraft hangars, arenas, and auditoriums. They
are well suited where people or materials are to be heated and it is not
necessary to heat the surrounding air. They are not suited for applications,
where combustible or explosive fumes are present, and where the additional
-39-
moisture due to the combustion products will be detrimental.
4.2.10 - Gas-Fired Radiant Tube Infrared Units
In this type of gas-fired infrared system, (see Figure 4), gas and air are
burned in a combustion chamber and the products of combustion are forced
through a tube providing a radiant energy source. The products of combustion
are then exhausted to the outside and not into the space being heated. Vari
ous types of reflectors located above the tube are used to concentrate and
direct the radiant energy to the lower levels of the structure where it is
needed. They are available as U-shaped or linear tubes for versatility in
designs. The radiant heating tubes are generally about 4 inches in diameter,
and during operation, the tubes generally vary in temperature from 900°F to
500°^ al°ng the length of the tube between the combustion chamber and the
exhaust vent.
These units provide a uniform source of radiant energy at a low intensity
level and are generally used in factories, warehouses, garages, aircraft
hangars, arenas, auditoriums, and other total heating applications in high
volume spaces. The major advantage of these units over other types of gas-
fired infrared units is their larger radiating surface and the fact that the
combustion products are vented from the space.
4.2.11 - Electric Infrared Units
Electric infrared units (see Figure 5) use metal rods, quartz tubes, lamps
or panels for delivering infrared heating to large volume spaces. Some units
use reflectors to concentrate and direct the radiant energy to areas where it
is needed. They do not require combustion air for operation and therefore do
not have to dispose of combustion products. Panel type units have a surface
temperature between 200°F and 1100°F while the metal and quartz tube units
-40-
operate at a surface temperature between 1500°F and 1800°F. Quartz lamps
generally operate at about 4000°F surface temperature.
These units are designed to be used for spot heating applications and
large volume spaces where people, objects, or surfaces are to be heated rather
than air. These include factories, warehouses, assembly areas, aircraft han
gars, arenas, swimming pools, and auditoriums. They are normally controlled
from shielded thermostats in the space.
-41-
m E 2 /miavncNS WORK
Type of
Radiant System
Hydremic Floor
Electric Floor
A1r Floor
Hydremic Wall
Electric W a l l
Hydremic Celling
Electric Celling
Ceramic Infrared
Tube Infrared
Electric Infrared
Surface Tenp.
F
85
85
85
100
100
55-230
120-200
1500-1700
700-1200
1100-4000
Integral With
Construction or
Add-On Possibility .
Integral
Integral
integral
iiitegral
Integral
Add-on
Add-on
Add-on
Add-on
Add-on
Response Tire
slow
slow
medium
msdlum
(redium
good
good
good
good
good
Total or
Spot Keating
total
total
total
total
total
total
total or
spot
total or
spot
total or
spot
total or
spot
Cooling Capacity
No
No
No
Yes
No-
Yes
No
No
No
No
Exhaust Venting System
Required
No
No
No
No
No
No
No
Yes
Yes
No
Condensation to be
Considered
No
No
No
I f Cooling
No
I f Cooling
No
I f not vented
I f not vented
No
Residential
X
X
X
X
X
X
X
Industrial
X
'•
X
X
X
X
Applications
Warehouse
X
X
X
X
X
Garage
X
X
X
X
X
X
Comrerclal Office
X
X
X
X
Sports Facility
X
X
X
X
X
School
X
X
X
X
Hospital
X
X
5.0 - CALCULATION OF DESIGN HEATING LOAD
This investigation is directed at providing a design procedure for
radiant heating and cooling systems. The major concern is whether or
not the ASHRAE standard heating load design procedure can be used for radiant
systems. The procedure developed for this project is based on the best avail
able information for radiant and convective exchange, but has not been vali
dated with experimental data. This procedure will be presented first, fol
lowed by a procedure developed specifically for this project, and a discussion
on the differences between these techniques and some other techniques found in
the literature.
5.1 - Standard ASHRAE Design Heat Loss Procedure
In Chapter 25 of Reference 1, what is commonly referred to as the ASHRAE
standard procedure is presented. From that source, the following general
procedure has been reproduced. GENERAL PROCEDURE
To calculate a design heating load, prepare the following information about building design and weather data at design conditions.
1. Select outdoor design weather conditions: temperature, wind direction and wind speed. Winter climatic data can be found in Chapter 24.
2. Select the indoor air temperature to be maintained in each space during coldest weather.
3. Estimate temperatures in adjacent unheated spaces. 4. Select or compute heat transfer coefficients for outside walls and
glass; for inside walls, nonbasement floors and ceilings, if these are next to unheated spaces; and the roof if it is next to heated spaces.
5. Determine net area of outside wall, glass and roof next to heated spaces, as well as any cold walls, floors or ceilings next to unheated spaces. These determinations can be made from building plans or from the actual building, using inside dimensions.
6. Compute heat transmission losses for each kind of wall, glass, floor, ceiling and roof in the building by multiplying the heat transfer coefficient in each case by the area of the surface and the temperature difference between indoor and outdoor air or adjacent unheated space.
7. Compute heat losses from basement or grade-level slab floors using the methods in this chapter.
8. Select unit values and compute the energy associated with infiltration of cold air around outside doors, windows and other openings. These unit values depend on the kind or width of crack, wind speed and the temperature difference between indoor and outdoor air. An alternative method is to use air changes. (See Chapter 22.)
9.' When positive ventilation using outdoor air is provided by an air-heating or air-conditioning unit, the energy required to warm the outdoor air to the space temperature must be provided by the unit. The principle for calculation of this load component is identical to that for infiltration. If mechanical exhaust from the space is provided in an amount equal to the outdoor air drawn in by the unit, the unit must also provide for natural infiltration losses. If no mechanical exhaust is used and the outdoor air supply equals or exceeds the amount of natural infiltration that can occur without ventilation, some reduction in infiltration may occur.
- 4 3 -
10. The sum of the transmission losses or heat transmitted through the confining walk, floor, ceiling, glass and other surfaces, plus the energy associated with cold air entering by infiltration or required to replace mechanical exhaust, represents the total heating load.
11. In buildings with a sizeable and reasonably steady internal heat release from sources other than the heating system, compute and deduct this heat release under design conditions from the total heat losses computed above.
12. Consider using pick-up loads that may be required in intermittently heated buildings or in buildings using night thermostat setback. Pick-up loads frequently require an increase in heating equipment capacity to bring the temperature of structure, air and material contents to the specified temperature.
In addition, a table which is given in [1] that summarizes the typical load
calculations is reproduced below.
TABLE 3 - Summary of ASHRAE Standard Design Heat Loss Calculations rX)
Healing Load Equation Reference, Table, Description
Roofs, ceilings, walls, glass q = U - A- TD
-•-Chapter 23, Tables 3 and 4
~[~ ^ ITemperature difference between inside and outside design dry bulbs. Chapter 24. For tem-*1peratures in unhealed spaces, see Eq. (1); for attic temperatures, s
-Area calculated from plans , see Eq. (2)
Walls below grade q = U' A • TD
-See Table 3
-*- Use Fig. 4 to assist in determining TD
Floors Above grade
On grade
Below grade
q = U-A-TD
q = F2 • P . TD
q = {/• A • TD
- For crawl space temperatures, see Eq. (4)
-See Table 5
Perimeter of Slab
-Use Fig. 4 to assist in determining TD -•-See Table 4
Infiltration and ventilation air Sensible Latent
J—-qs = 1200K* Af
q, =2808K. LW
_JVolume of outdoor air entering building. See Chapter 22 for estimating methods for in-Ifiltration
-Humidity ratio difference
-44-
The components of the ASHRAE standard design heat loss (HLD) are transmission
losses and infiltration losses. The transmission losses are:
S UjAi (75 - toa) (16)
and the infiltration loss is
1.1 CFMI (75-toa) (17)
or
ACHxV
1.1 x ( ) x (75 - toa) (18)
60
where ACH is the number of infiltration air changes per hour and V is the volume
of the space.
5.1.1 - Design Inside Air Temperature. It is common practice to select the
inside design dry bulb temperature at 75 F in most localities in the United
States. Generally, this is done without accounting for the comfort con
straints previously described. This temperature is used in both the transmis
sion loss and infiltration loss calculations and a choice of this value will
affect the design loads for the space in proportion to the temperature differ
ence between the inside and outside at design conditions. As indicated in the
section 3.0 -DEFINITIONS AND TERMINOLOGY, the value used for the inside design
temperature in this analysis has been set at 75 F.
5.1.2 - Room Air Temperature Gradients. For rooms which are eight to ten
feet high, a small temperature gradient in room air may exist as discussed,
but is usually not incorporated into the design heat loss calculation. For
higher ceiling/roof rooms or spaces, this gradient can affect the design heat
-45-
loss due to higher air temperatures at the ceiling which can cause higher
transmission and infiltration losses. For the ASHRAE standard design heat
loss calculations (HLD) made later, it is assumed that there is no air temper
ature gradient. However, this gradient is incorporated into another design
heat loss (HLGG) to illustrate what effect it has on the results. Typical
values of this air temperature gradient are 0.5 to 2.0 F per foot.
5.1.3 - Wall. Ceiling. Floor Convection Coefficients. The U-factors indicated
in the transmission loss component include convection on the inside walls,
floors, and ceiling that contain a contribution from radiation as well as con
vection. These values have been standardized over the years and are commonly
used in all design heating and cooling load calculations. When radiant sys
tems are considered, it is important to realize that these standard coeffi
cients may no longer apply due to higher surface temperatures. For the stan
dard ASHRAE design heat loss procedure (HLD) the convection coefficients in
Reference [1] were used and these are given in Table 4. Also shown as Table 5
are common emissivity values for building materials.
Table 4. Surface Conductances (Btu/hr ft^ F) and
Resistances for Air [1]. Surface Emittanee
Position of Direction Non- Reflective Reflective Surface of Heat reflective t = 0.20 t = 0.05
Flow € = 0.90
h, R h| R h, R
STILL AIR Horizontal . . . . . Upward 1.63 0.61 0.91 1.10 0.76 1.32 Sloping—45 deg Upward 1.60 0.62 0.88 1.14 0.73 1.37 Vertical Horizontal 1.46 0.68 0.74 1.35 0.59 1.70 Sloping—45 deg Downward 1.32 0.76 0.60 1.67 0.45 2.22 Horizontal . . . . . . Downward 1.08 0.92 0.37 2.70 0.224.55
MOVING AIR h0 R h0 R h0 ~R (Any Position) 15-mphWind Any 6.00 0.17
(for winter) 7.5-mphWmd Any 4.00 0.25 (for summer) .__ a N o surface has both an air space resistance value and a surface resistance
value. No air space value exists for any surface facing an air space of less than 0.5 in.
°For ventilated attics or spaces above ceilings under summer conditions (heat flow down) see Table 4.
Conductances are for surfaces of the stated emittanee facing virtual blaciibody surroundings at the same temperature as the ambient air. Values are based on a surface-air temperature difference of 10 deg F and for surface temperature of 70 F.
dSee Fig. 2 for additional data.
- 4 6 -
Table 5. Reflectivity and Emittance Values of Various Surfaces [1].
Effective Emittance E of Air Space
Surface Reflectivity. Average One Both in Percent Emittance t surface surfaces
emit- emittance v, iancest
the other <U|0
Aluminum foil, bright 92 to 97 0.05 0.05 0.03 Aluminum sheet 80to95 0.12 0.12 0.06 Aluminum coated paper,
polished 75to84 0.20 0.20 0.11 Steel,galvanized,bright... 70to80 6.25 0.24 0.15 Aluminum paint 30 to 70 0.50 0.47 0.35 Building materials: wood,
paper, masonry, '. nonmetallic paints 5 to 15 0.90 0.82 0.82
Regular glass 5 to 15 0.84 0.77 0.72
The U-factors for the transmission losses also contain outside convection
coefficients for the walls and floors "and ceilings if appropriate.
5.2 - Development of Design Heat Loss Procedure for Radiant Systems
It is necessary with radiant types of systems to be able to estimate the
design heat loss value so that units can be sized and located. It is also
important for this study, to be able to compare this design load with the ASH-
RAE standard procedure (HID) described above, and to investigate the effect of
changing specific parameters in the design process for radiant units. These
include effect of higher room surface temperatures, consequences of higher
mean radiant temperatures and lower air temperatures, and changes in the
infiltration heat loss term.
In Figure 17, a schematic of the room configuration used for the calcula-;
tion of the radiant design heat loss values is shown. There are six surfaces
specified, four walls, a floor, and a ceiling. For the floor, an emissisivity
(e), convection coefficient (hc) and a U-factor (U) are specified. The ceil
ing is composed of several portions: heating or cooling panels and ceiling.
-47-
I •p-oo l
Loss-w3
PANEL Heat Input
£
h
GLASS
Wall 1 e, h ,U
c
Loss-wl
Wall 2
£,h ,U c
Loss-w2
'infiltration
Floor
e,h ,U c
Loss - F
FIGURE 17. SCHEMATIC OF ROOM CONFIGURATION USED FOR THE CALCULATION OF RADIANT DESIGN HEAT LOSS VALUES
For all of the equally sized panels an emissivity (e) and convection coeffi
cient (hc) must be specified. There is no U-factor specified for the heat
ing/cooling panels since this would vary considerably from unit to unit and it
can be taken into account in the design process. The remainder of ceiling
(See Figure 17) has an emissivity (e). convection coefficient (hc), and a
U-factor (U) specified.
The four walls can be individually described by giving an emissivity (e),
convection coefficient (hc) and a U-factor (U) for each wall. This allows the
walls to be outside or inside walls by using the actual U or a small U (0.001
Btu/hr ft^ F) value. Also, the contribution of glass in the outside walls can
be varied by specifying appropriate values of the wall U-factor.
By specifying the size of the room (length and width), the room height,
the number of ceiling panels and their coordinate locations the geometry of
room is defined. From this information the angle factors between all of the
room surfaces can then be calculated.
As indicated in Figure 17 there is a contribution to the total design heat
loss by the infiltration term. For this analysis an air change per hour (ACH)
was specified as input information and then with the room volume known an air
volume could be calculated.
It is important to understand that all of the surfaces shown in Figure 17
are coupled thermally through their radiant exchange and their convective
exchange with the room air. In addition, the outside walls will transfer heat
to the surroundings as will the floor and the ceiling. The heating panels
will be supplying heat by radiation and convection to the other surfaces and
the room air. At the same time, the infiltration air will be affecting the
overall heat balance of the room air.
The following sections contain a description of the system of equations
solved using a computer program and how these equations were formulated. To
-49-
achieve this aim, it was found that the following system of equations needed
to be solved:
i) Heat balance on the room surfaces (six surfaces).
ii) Heat balance on the complete room,
iii) The comfort equations (two equations).
iv) The definition of mean radiant temperature.
This results in ten equations to be solved, where nine of the equations are
coupled and eight of the nine are non-linear. This required solving a system
of nine non-linear equations simultaneously. The solution was done using an
algorithm based on Newton's Method. Once this system of equations was solved
and all of the temperatures known, the design heat losses and other parameters
were evaluated.
5.2.1 - Heat Balance on Room Surfaces TA1. Each room surface area A^ as
illustrated in Figure 18 is in radiant exchange with all the other surfaces
and is
Surface i
Figure 18. Surface Heat Exchange Model
in convective exchange with the air in the room. The sum of these two heat
flows, qr and qcv will, under steady state conditions, be equal to the
conductive heat flow through the surface as shown below.
1r + <Icv + qcd = ° (19>
-50-
wherej
qr = net radiant heat transfer from Aj
qcv — convection between air and surface A^
qC(j = conduction through surface A^
5.2.1.1 - qr. Radiant Exchange Rate. For emittances of surfaces at or
above 0.9, surface reflections can be ignored from surfaces and the radiant
exchange can be expressed as:
4 4
q r i = £i<7 Ti - S ei a Ti FA _A (20)
where,
qr j_ = net radiant heat transferred from surface i per unit area, A^ and
per unit time.
T^ «=» absolute temperature of surface A^
e^ = emittance of surface i.
a = Stefan-Boltzman Constant
FA^-Aj =• angle factor from surface i to surface j.
The angle factors were calculated from algorithms available in References
5, 6, and 16.
5.2.1.2 - qcv Convective Heat Transfer. This term is evaluated from the
following equation,
qcv.i = h c > i (Ti " Ta> <21>
where,
qcv £ <= convective heat transfer from surface i to air per unit area Aj
hc ^ = the appropriate convection heat transfer coefficient
Ta = air temperature
T^ = surface A^ temperature
The hc ^ heat transfer coefficients selected were different for the non-
radiant heating and radiant heating cases. The reason for this is that
-51-
in the non-radiant heating systems, higher air velocities and lower surface
temperatures are expected. For the non-radiant heating calculation the fol
lowing coefficients were used.
Ceiling/floor, upward heat flow [1, 16]
hc =0.712 Btu/hr ft2 F
Ceiling/floor, downward heat flow [1, 16]
hc = 0.162 Btu/hr ft2 F
Walls [17]
At hc = 2.03 (—)
0-22, w/m2 C H
where,
At = average surface to average air temperature difference, C
H = height of room, m
(5.68 w/m2 C = 1 Btu/hr ft2 F)
For the radiant heating systems the following convection coefficients were
used.
a) Heated Ceiling
Heated Ceiling Panels [16,18],
(At)0-25 h„ = 0.041 , Btu/hr ft2 F
De0-25
where,
De =• equivalent diameter (4 times the area divided by perimeter)
Unheated Ceiling Portion [16, 18],
hc - 0.162 Btu/hr ft2 F
Walls [16, 18],
(At)0-32
hc t= 0.29 H0.05
Floor [16, 18],
-52-
hc - 0.712 Btu/hr ft2 F
b) Heated Floor
Heated Floor Portion [16,18]
AtO-31 hc - 0.39
D, 0.08
where,
equivalent diameter (4 times the area divided by perimeter)
Unheated Floor/Ceiling Area [16,18],
hc = 0.712 Btu/hr ft2 oF, upward heat flow
hc = 0.162 Btu/hr ft2oF, downward heat flow
Walls [16,18] 0.29(At)0-32
hc -H0.05
5.2.1.3 - qCc[. Conductive Heat Transfer. Under steady state conditions, the heat conduction per unit area Aj_ is given by
qcd.i - ^ (Ti - TQ) (22) where,
C^ = overall wall conductance from inside surface to outside air
*1
Kl + — +
al
... +
where,
x^ = thickness of each homogeneous section of the wall
k^ = thermal conductivity of the material
a^ =• conductance of each air space in the wall
hQ = coefficient of heat transfer by convection and radiation at
the outside surface of the wall
Tj_ = inside surface temperature of surface i.
TQ = outside ambient design temperature
For calculation in the program the following was used to calculate Cj.
-53-
I l l (23)
Ci Ui ht
where,
Uj = The overall heat transfer coefficient from the inside air
to the outside air using standard or typical ASHRAE
values [1]
h^ •= Convective heat transfer coefficient from inside air to
inside surface i. This was the typical design value for
this coefficient as given in Table 2. These are standard
or typical values used by designers and includes
convection and radiation heat transfer.
Eq. 23 was used in order to eliminate the standard dual convection coeffi
cient which includes both radiation and convection terms. It was necessary to
use only the true convection coefficient since the procedure in the calcula
tion method accounted for the radiation.
5.2.2 - Heat Balance on the Complete Room. It is necessary from the first law
of thermodynamics to maintain a heat balance on the air within the room (see
Figure 17). This is given by the following equation.
where,
and.
Total Heat Gain - Total Heat Loss -= 0 (24)
Total Heat Gain - Q i n p u t + Qpeople + Qlights
Total Heat Loss - QTransmission Loss + ^Infiltration Loss
Qinput "* Heat input by supply air in the convective heating case
or by panels in the panel heating case.
Qpeople = Internal sensible heat gain from people in space (this
was set equal to zero for the design heating case).
-54-
Qliehts = Internal heat gain due to lights (this was set equal
to zero for the design heating case),
^transmission loss = ^um °^ t n e conduction losses through each
of the six surfaces,
^infiltration loss = Heat loss due to infiltration air.
5.2.3 - Comfort Equations. The objective of the heating or cooling system is
to provide thermal comfort for people in the room illustrated in Figure 17.
In order to do this a set of comfort criteria needed to be selected. For this
study, the Fanger Comfort Criteria [4] were chosen and were previously dis
cussed in Section 4.1.1. Fanger considers the simultaneous influence of six
operating variables for comfort. These are,
• Activity level (internal heat production in units of MET)
• Thermal Resistance of Clothing (in units of clo)
• Air Temperature ( F)
• Mean Radiant Temperature ( F)
• Humidity Level (in terms of relative humidity)
• Relative Air Velocity (m/s)
The comfort equations can be expressed as
M tcl - f( , t,, Pa, Ta) (25)
ADu
t .,- = f(F p i for i = 1,6, tt for i= 1 to 6) (26)
Mi 0 = f(Ta, tcl, tnrt, Pa, , r,) (27)
ADu where,
Apu = Dubois surface area of a person, m^
M = metabolic rate of person, MET
Pa = partial pressure of water vapor in the room air (a
function of air temperature and relative humidity)
-55-
Ta — room air temperature, C
Fp i «= angle factor from the center of a seated person at center
of room to surface i
tj_ = surface i temperature
r\ = mechanical efficiency of person
tc2 = temperature of clothing surface
t m r t = mean radiant temperature at center of room
The functional relationship indicated by Eq. 25 is that given as Eq. 15.
The functional relationship indicated by Eq. 26 is that given as Eq. 4. And
finally, the functional relationship indicated by Eq. 27 is that given by Eq.
14. For details of the development of these equations one should see Fanger
5.2.4 - Other Parameters Evaluated. The following other significant par
ameters were evaluated from the computer results.
• ASHRAE Standard Heat Loss (HID) - Eqn. 10
• Actual Design Heat Loss (HLA) - Eqn. 11
• Conduction Design Heat Loss (HLC) - Eqn. 12
• Conduction Design Heat Loss with Room Air Temperature Gradient
(HLCG) - Eqn. 13
• Actual Heat Input
In the case of convective heating systems, the actual
heat input is the same as the conduction design heat
loss with room air temperature gradient, HLCG. But, in
the case of panel heating, the actual heat input term
assumes there is no heat loss from the reverse sides of
the panels. HLCG assumes the same conductive resistance
at the back of the heated panels as in the rest of the
-56-
surface and Is therefore a greater value than the actual
heat input term. The design procedure for each type
of system will then take into account the losses from
the backs of the heating panels. This is necessary since
each manufacturer will have different types of insulating
schemes.
• Percentage Difference 1 - The difference in percent between
HLA and HLD.
• Percentage Difference 2 - The difference in percent between
HLC and HLD.
• Percentage Difference 3 - The difference in percent between
HLCG and HLD.
• Percentage Difference 4 - The difference in percent between
Actual Heat Input and HLD.
• Operative Temperature - Eqn. 7
• Effective Radiant Flux - Eqn. 8
• Average Unheated Surface Temperature (AUST) - Eqn. 9
• Parameter 1 -
Qinput 0 , Btu/hr ft2 °F (28)
Panel Area (T - Ta) P
where,
Qinput = Actual Heat Input
Tp = panel surface temperature
Ta = room air temperature
• Parameter 3 - Dimensionless
-57-
Qinput
Panel Area [T4 - (AUST)4] a P
where,
a = Stefan Boltzman Constant
• Percentage Radiation -
QRP
QRP + QCVP
(29)
(30)
-58-
where,
QRP ~* radiant heat output by panels
QCVp = convective heat output by panels
5.3 - Comparison of Calculated Design Radiant Loads With the Standard ASHRAE
Design Load Calculation
Many cases have been run for both forced air and radiant systems in order
to determine the effect of various parameters and variables on the design heat
loss. A base configuration was selected and this was used to make initial
calculations and then changes in the parameters were made in order to test
their effect on the value of the design heat loss. The configuration was the
following.
Outside Design Temperature = 3 F
Room Dimensions: Length = 30 ft.
Width = 30 ft.
Height = 9 ft.
U-Factors (Btu/hr ft2 F):
Wall 1 - half wall with U = 0.1 and half glass with
U = 0.58. Glass distributed uniformly
over the wall,
Walls 2,3,4 - U - 0.1
Floor - U = 0.07
Ceiling - U - 0.07
Emissivities:
Panels: 0.9
Walls: 0.9
Floor: 0.9
Ceiling: 0.9
Convection Coefficients (Btu/hr ft2 °F):
-59-
See Section 5.2.1.2
Comfort Variables:
Metabolic Rate: - 75 k cal/hr m2
Clothing Resistance =0.75 clo (fci —1.1)
Relative Air Velocity =0.15 m/s
Relative Humidity = 30%
Infiltration Rate =0.5 air changes per hour
For convection heating, the supply air flow rate was set at 0.75
CFM per sq. ft. of floor area.
For convection heating, the air temperature gradient was set at
0.75 F per foot with a reference height of 5 feet from the
floor.
For radiant heating, there was no supply air, and the room air
temperature gradient was set at 0 F per foot.
For these design calculations, the number of people was set
at zero, and no lighting load was considered. For
cooling cases, this would not be the case.
5.4 -Test Case Calculations. In order to be able to evaluate the performance
of the computational scheme, the forced air heating case was taken as a test
case. This allowed the design heat loss values to be calculated and compared
with the standard ASHRAE procedure. In these calculations the convection
coefficient on the walls, floors, and ceiling were not changed during the
operation of the system. The values given in Section 5.2.1.2 were used and
remained constant (except for the walls where they were a function of the At).
The standard forced air heating cases are given in table 6 for various
heights of the room. As the room height increases, the ASHRAE Design Heat
Loss (HLD) increases and correspondingly so does the supply air temperature.
This is due to increased infiltration as well as the increase in wall and
-60-
TABLE 6. FORCED AIR HEATING - STANDARD CASE CALCULATIONS
ROOM HEIGHT, FT
ASHRAE DESIGN HEAT LOSS, BTU/HR
ACTUAL DESIGN HEAT LOSS, BTU/HR
PERCENTAGE DIFFERENCE 1
CONDUCTION DESIGN HEAT LOSS 1, BTU/HR
PERCENTAGE DIFFERENCE 2
CONDUCTION DESIGN HEAT LOSS 2, BTU/HR
PERCENTAGE DIFFERENCE 3
ACTUAL HEAT INPUT, BTU/HR
PERCENTAGE DIFFERENCE 4
FLOOR TEMPERATURE, DEG. F
ROOM AIR TEMPERATURE, DEG. F
MEAN RADIANT TEMPERATURE, DEG. F
OPERATIVE TEMPERATURE, DEG. F
EFFECTIVE RADIANT FIELD, BTU/HR.
A.U.S.T, DEG. F
SUPPLY AIR TEMPERATURE, DEG. F
SQ FT
8.0
24796.8
25642.2
3.4
22796.8
-8.1
22942.6
-7.5
22942.6
-7.5
61.5
77.5
62.6
69.2
-10.7
62.4
101.1
9.0
26762.4
27792.8
3.9
24645.9
-7.9
24864.6
-7.1
24864.6
-7.1
61.4
77.8
62.2
69.1
-11.3
62.3
103.4
10.0
28728.0
29895.6
4.1
26443.6
-8.0
26747.4
-6.9
26747.4
-6.9
61.2
77.9
61.9
69.1
-11.5
62.1
105.4
12.0
32659.2
34124.6
4.5
30039.1
-8.0
30549.4
. -6.5
30549.4
-6.5
60.9
78.2
61.5
69.0
-12.1
61.7
109.7
15.0
38556.0
40482.8
5.0
35399.4
-8.2
36310.7
-5.8
36310.7
-5.8
60.3
78.6
61.0
68.8
-12.7
61.1
116.0
20.0
48384.0
51256.8
5.9
44389.9
-8.3
46212.4
-4.5
46212.4
-4.5
59.7
79.3
60.0
68.6
-13.9
60.4
126.8
25.0
58212.0
62055.1
6.6
53297.4
-8.4
56334.9
-3.2
56334.9
-3.2
59.1
79.8
59.3
68.4
-14.7
59.8
137.7
glass areas as the wall height is raised causing larger heat losses. This
shows up also in a reduced value of AUST with increasing height. As the room
height increases, the infiltration air leaving the room at the ceiling level
is at a higher temperature due to an air temperature gradient. The ASHRAE
design heat loss, HLD, overestimates the calculated heat loss HLC or HLCG by
about 7% for an 8 feet high room and by about 3% for a 25 feet high room even
with a temperature gradient. It is also important to notice that the room air
temperature for comfort is about 77 F for the 8 feet high and almost 80 F for
the 25 feet high room. This is due to the mean radiant temperature dropping
because of more glass surface in the higher room and therefore a higher air
temperature being required to satisfy the comfort equations. These higher air
temperatures are consistent with the results presented in Table 1 for comfort
conditions.
Tables 7, 8 and 9 show similar results as Table 6 except that the air
temperature gradient was changed to 0.5, 1.0 and 1.5 F per foot respectively.
Similar results are exhibited except that the ASHRAE design heat loss, HLD,
underestimates the heat loss by about 2% for the 25 feet high room with a
temperature gradient of 1.5 F per foot.
Tables 10 and 11 give the results for the forced air heating system stan
dard case with different infiltration rates and for a 15 ft. and 25 ft. high
room respectively. The results in Table 12 are for the same conditions as in
Table 11 except that the tJ-factors were increased to what might be expected in
industrial situations. Comparison of these results show that the ASHRAE
design heat loss calculation can underestimate the size of the heating load
for high (greater than 2) infiltration air changes. This underestimation can
be up to 16% at 4 air changes per hour. It should be noted in Tables 10, 11,
and 12 that the supply air temperatures are not appropriate. The air flow
rate was set at 0.75 CFM/ft , and for higher heat losses as found here, this
-62-
TABLE 7. FORCED AIR HEATING - WITH GRADIENT = 0.5°F/FT.
ROOM HEIGHT, FT
ASHRAE DESIGN HEAT LOSS, BTU/HR
ACTUAL DESIGN HEAT LOSS, BTU/HR
PERCENTAGE DIFFERENCE 1
CONDUCTION DESIGN HEAT LOSS 1 , BTU/HR
PERCENTAGE DIFFERENCE 2
CONDUCTION DESIGN HEAT LOSS 2, BTU/HR
PERCENTAGE DIFFERENCE 3
ACTUAL HEAT INPUT, BTU/HR
PERCENTAGE DIFFERENCE 4
FLOOR TEMPERATURE, DEG. F
ROOM AIR TEMPERATURE, DEG. F
MEAN RADIANT TEMPERATURE, DEG. F
OPERATIVE TEMPERATURE, DEG. F
EFFECTIVE RADIANT FIELD, BTU/HR.
A.U.S.T, DEG. F
SUPPLY AIR TEMPERATURE, DEG. F
SQ FT
8.0
24796.8
25642.2
3.4
22796.8
-8 .1
22894.0
- 7 . 7
22894.0
- 7 . 7
61.5
77.5
62.6
69.2
-10 .7
62.4
101.0
9.0
26762.4
27792.8
3.9
24645.9
- 7 . 9
24791.7
- 7 . 4
24791.7
- 7 . 4
61.4
77.8
62.2
69.1
-11 .3
62.3
103.3
10.0
28728.0
29895.6
4.1
26443.6
- 8 . 0 r
26646.1
-7 .2
26646.1
- 7 . 2
61.2
77.9
61.9
69.1
-11.5
62.1
105.3
12.0
32659.2
34124.6
4.5
30039.1
- 8 . 0
30379.3
-7 .0
30379.3
-7 .0
60.9
78.2
61.5
69.0
-12 .1
61.7
109.5
15.0
38556.0
40482.8
5.0
35399.4
- 8 . 2
36006.9
-6 .6
36006.9
-6 .6
60.3
78.6
61.0
68.8
-12 .7
61.1
115.6
20.0
48384.0
51256.8
5.9
44389.9
- 8 . 3
45604.9
-5 .7
45604.9
-5 .7
59.7
79.3
60.0
68.6
-13 .9
60.4
126.2
25.0
58212.0
62055.1
6.6
53297.4
- 8 . 4
55322.4
-5 .0
55322.4
- 5 . 0
59.1
79.8
59.3
68 .4
-14 .7
59,8
136.7
TABLE 8. FORCED AIR HEATING - WITH GRADIENT = 1.0°F/FT.
ROOM HEIGHT, FT
ASHRAE DESIGN HEAT LOSS, BTU/HR
ACTUAL DESIGN HEAT LOSS, BTU/HR
PERCENTAGE DIFFERENCE 1
CONDUCTION DESIGN HEAT LOSS 1, BTU/HR
PERCENTAGE DIFFERENCE 2
CONDUCTION DESIGN HEAT LOSS 2, BTU/HR
PERCENTAGE DIFFERENCE 3
ACTUAL HEAT INPUT, BTU/HR
PERCENTAGE DIFFERENCE 4
FLOOR TEMPERATURE, DEG. F
ROOM AIR TEMPERATURE, DEG. F
MEAN RADIANT TEMPERATURE, DEG. F
OPERATIVE TEMPERATURE, DEG. F
EFFECTIVE RADIANT FIELD, BTU/HR.
A.U.S.T, DEG. F
SUPPLY AIR TEMPERATURE, DEG. F
SQ FT
8.0
24796.8
25642.2
3.4
22796.8
-8.1
22991.2
-7.3
22991.2
-7.3
61.5
77.5
62.6
69.2
-10.7
62.4
101.1
9.0
26762.4
27792.8
3.9
24645.9
-7.9
24937.5
-6.8
24937.5
-6.8
61.4
77.8
62.2
69.1
-11.3
62.3
103.4
10.0
28728.0
29895.6
4.1
26443.6
-8.0
26848.6
-6.5
26848.6
-6.5
61.2
77.9
61.9
69.1
-11.5
62.1
105.5
12.0
32659.2
34124.6
4.5
30039.1
-8.0
30719.5
-5.9
30719.5
-5.9
60.9
78.2
61.5
69.0
-12.1
61.7
109.8
15.0
38556.0
40482.8
5.0
35399.4
-8.2
36614.4
-5.0
36614.4
-5.0
60.3
78.6
61.0
68.8
-12.7
61.1
116.3
20.0
48384.0
51256.8
5.9
44389.9
-8.3
46819.9
-3.2
46819.9
-3.2
59.7
79.3
60.0
68.6
-13.9
60.4
127.4
25.0
58212.0
62055-1
6.6
53297.4
-8.4
57347.4
-1.5
57347.4
-1.5
59.1
79.8
59.3
68.4
-14.7
59.8
138.8
TABLE 9. FORCED AIR HEATING - WITH GRADIENT = 1.5°F/FT.
ROOM HEIGHT, FT
ASHRAE DESIGN HEAT LOSS, BTU/HR
ACTUAL DESIGN HEAT LOSS, BTU/HR
PERCENTAGE DIFFERENCE 1
CONDUCTION DESIGN HEAT LOSS 1, BTU/HR
PERCENTAGE DIFFERENCE 2
CONDUCTION DESIGN HEAT LOSS 2, BTU/HR
PERCENTAGE DIFFERENCE 3
ACTUAL HEAT INPUT, BTU/HR
PERCENTAGE DIFFERENCE 4
FLOOR TEMPERATURE, DEG. F
ROOM AIR TEMPERATURE, DEG. F
MEAN RADIANT TEMPERATURE, DEG. F
OPERATIVE TEMPERATURE, DEG. F
EFFECTIVE RADIANT FIELD, BTU/HR.
A.U.S.T, DEG. F
SUPPLY AIR TEMPERATURE, DEG. F
SQ FT
8.0
24796.8
25642.2
3.4
22796.8
-8.1
23088.4
-6.9
23088.4
-6.9
61.5
77.5
62.6
69.2
-10.7
62.4
101.2
9.0
26762.4
27792.8
3.9
24645.9
-7.9
25083.3
-6.3
25083.3
-6.3
61.4
77.8
62.2
69.1
-11.3
62.3
103.6
10.0
28728.0
29895.6
4.1
26443.6
-8.0
27051.1
-5.8
27051.1
-5.8
6-1.2
77.9
61.9
69.1
-11.5
62.1
105.8
12.0
32659.2
34124.6
4.5
30039.1
-8.0
3!1059.7
-4.9
31059.7
-4.9
60.9
78.2
61.5
69.0
-12.1
61.7
110.2
15.0
38556.0
40482.8
5.0
35399.4
-8.2
37221.9
-3.5
37221.9
-3.5
60.3
78.6
61.0
68.8
-12.7
61.1
116.9
20.0
48384.0
51256.8
5.9
44389.9
-8.3
48034.9
-0.7
48034.9
-0.7
59.7
79.3
60.0
68.6
-13.9
60.4
128.7
25.0
58212.0
62055.1
6.6
53297.4
-8.4
59372.4
2.0
59372.4
2.0
59.1
79.8
59.3
68.4
-14.7
59.8
140.8
TABLE 10. FORCED AIR HEATING.- STANDARD CASE-15 FT
AIR CHANGES PER HOUR, AC/H
. ASHRAE DESIGN HEAT LOSS, BTU/HR
ACTUAL DESIGN HEAT LOSS, BTU/HR
PERCENTAGE DIFFERENCE 1
CONDUCTION DESIGN HEAT LOSS 1, BTU/HR
PERCENTAGE DIFFERENCE 2
CONDUCTION DESIGN HEAT LOSS 2, BTU/HR
PERCENTAGE DIFFERENCE 3
i« ACTUAL HEAT INPUT, BTU/HR
1 PERCENTAGE DIFFERENCE 4
FLOOR TEMPERATURE, DEG. F
ROOM AIR TEMPERATURE, DEG. F
MEAN RADIANT TEMPERATURE, DEG. F
OPERATIVE TEMPERATURE, DEG. F
EFFECTIVE RADIANT FIELD, BTU/HR. SQ FT
A.U.S.T, DEG. F
SUPPLY AIR TEMPERATURE, DEG. F
HEIGHT AND VARIABLE INFILTRATION RATES
1.0
47304.0
49667.9
5.0
44584.6
-5.7
46407.1
-1.9
46407.1
-1.9
60.3
78.6
61.0
68.8
-12.7
61.1
126.3
2.0
64800.0
68038.3
5.0
62954.9
-2.8
66599.9
2.8
66599.9
*'2.8
60.3
78.6
61.0
68.8
-12.7
61.1
147.1
3.0
82296.0
86408.6
5.0
81325.2
-1.2
86792.7
5.5
86792.7
5.5
60.3
78.6
61.0
68.8
-12.7
61.1
167.9
4.0
99792.0
104778.9
5.0
99695.6
-0.1
106985.6
7.2
106985.6
7.2
60.3
78.6
61.0
68.8
-12.7
61.1
188.7
•#£e gXrv-^gj &*':••:~? FB-Hras p»;-S:;«<
TABLE 1 1 . FORCED AIR HEATING - STANDARD CASE WITH 25 FT. HEIGHT AND VARIABLE INFILTRATION RATES
ON
I
INFILTRATION, AC/H
ASHRAE DESIGN HEAT LOSS, BTU/HR
ACTUAL DESIGN HEAT LOSS, BTU/HR
PERCENTAGE DIFFERENCE 1
CONDUCTION DESIGN HEAT LOSS 1, BTU/HR
PERCENTAGE DIFFERENCE 2
CONDUCTION DESIGN HEAT LOSS 2, BTU/HR
PERCENTAGE DIFFERENCE 3
ACTUAL HEAT INPUT, BTU/HR
PERCENTAGE DIFFERENCE 4
FLOOR TEMPERATURE, DEG. F
ROOM AIR TEMPERATURE, DEG. F
MEAN RADIANT TEMPERATURE, DEG. F
OPERATIVE TEMPERATURE, DEG. F
EFFECTIVE RADIANT FIELD, BTU/HR. SQ FT
A.U.S.T, DEG. F
SUPPLY AIR TEMPERATURE, DEG. F
1.0
72792.0
77597.6
6.6
68840.0
-5.4
74915.0
2.9
74915.0
2.9
59.1
79.8
59.3
68.4
-14.7
59.8
156.8
2.0
101952.0
108682.7
6.6
99925.1
-2.0
112075.1
9.9
112075.1
9.9
59.1
79.8
59.3
68.4
-14.7
59.8
195.1
3.0
131112.0
139767.8
6.6
131010.2
-0.1
149235.2
13.8
149235.2
13.8
59.1
79.8
59.3
68.4
-14.7
59.8
233.3
4.0
160272.0
170852.9
6.6
162095.3
1.1
186395.3
16.3
186395.3
16.3
59.1
79.8
59.3
68.4
-14.7
59.8
271.5
TABLE 12. FORCED AIR HEATING - STANDARD CASE WITH U ,lc .. „„ „„,,,„„ = 0.25 Btu/hr ft2°F,
? walls,floors, ceiling Uglass = 1 ,° Btu/hr ft- F» 25 ft- HEIGHT AND VARIABLE INFILTRATION
INFILTRATION, AC/H
ASHRAE DESIGN HEAT LOSS, BTU/HR
ACTUAL DESIGN HEAT LOSS, BTU/HR
PERCENTAGE DIFFERENCE 1
CONDUCTION DESIGN HEAT LOSS 1, BTU/HR
PERCENTAGE DIFFERENCE 2
CONDUCTION DESIGN HEAT LOSS 2, BTU/HR
PERCENTAGE DIFFERENCE 3
ACTUAL HEAT INPUT, BTU/HR
PERCENTAGE DIFFERENCE 4
FLOOR TEMPERATURE, DEG. F
ROOM AIR TEMPERATURE, DEG. F
MEAN RADIANT TEMPERATURE, DEG. F
OPERATIVE TEMPERATURE, DEG. F
EFFECTIVE RADIANT FIELD, BTU/HR.
A.U.S.T, DEG. F
SUPPLY AIR TEMPERATURE, DEG. F
SQ FT
1.0
135810.0
161338.2
18.8
119554.6
-12.0
125629.6
-7.5
125629.6
-7.5
42.5
88.5
45.9
65.1
-30.3
47.7
217.8
2.0
164970.0
195979.4
18.8
154195.8
-6.5.
166345.8
0.8
166345.8
0.8
42.5
88.5
45.9
65.1
-30.3
47.7
259.7
3.0
194130.0
230620.7
18,8
188837.0
-2.7
207062.0
6.7
207062.0
6.7
42.5
88.5
45.9
65.1
-30.3
47.7
301.6
4.0
223290.0
265261.9
18.8
223478.2
0.1
247778.2
11.0
247778.2
11.0
42.5
88.5
45.9
65.1
-30.3
47.7
343.4
would have to be raised to approximately 3 CFM/ftr to yield reasonable supply
air temperatures. This calculation does not affect the design load calcula
tions .
The convective calculations appear to be reasonable and correct and do not
show any unusual results. They indicate that the program is calculating
values that are expected and show that the ASHRAE standard design procedure
tends to slightly overestimate design losses even with an air temperature
gradient present except for high (above 2) air changes per hour of infiltra
tion.
5.5 - Radiant Panel Heating Systems Calculations
5.5.1 - Single Panel Radiant Heating Cases. The same base case was taken
as in the forced air system except a single radiant heating panel was used to
supply heat to the room and there was no heated supply air. In this proce
dure the panel temperature was assumed as input information and a trial and
error procedure was used to determine the required area for the heat loss from
the space. In this calculation, the emissivity of the panel heater was set
at 0.9 and its convection coefficient was as previously specified.
In Table 13, the results for panel surface temperatures from 120 to 180 F
are shown for the base case room. As expected, the area required for heating
with panels reduced as the panel temperature increased. For 120 F,
approximately 49% of the ceiling area was covered with radiant panels while
for a 180 F panel temperature approximately 20% of the ceiling was covered
with radiant panels. The areas calculated here were compared with the
required area from two manufacturers of hydronic panels and showed quite close
agreement. At 120 F one manufacturer's procedure indicated 453 sq. ft. and
the other manufacturer's procedure indicated 415 sq. ft. The calculation here
indicated 439 sq. ft. At 180 F the two numbers were 216 and 185 sq. ft. and
the calculated area was 176 sq. ft. This information appears to verify the
-69-
TABLE 13. BASE CASE FOR RADIANT PANEL HEATING:
PANEL TEMPERATURE, DEC F
PANEL AREA REQUIRED , SQ FT
ASHRAE DESIGN HEAT LOSS, BTU/HR
ACTUAL DESIGN HEAT LOSS, BTU/HR
PERCENTAGE DIFFERENCE 1
CONDUCTION DESIGN HEAT LOSS 1, BTU/HR
PERCENTAGE DIFFERENCE 2
CONDUCTION DESIGN HEAT LOSS 2, BTU/HR
PERCENTAGE DIFFERENCE 3
ACTUAL HEAT INPUT, BTU/HR
PERCENTAGE DIFFERENCE 4
PERCENTAGE RADIATION
PERCENT CEILING COVERED BY PANELS
HEAT OUTPUT PER UNIT PANEL AREA,
FLOOR TEMPERATURE, DEG. F
ROOM AIR TEMPERATURE, DEG. F
MEAN RADIANT TEMPERATURE, DEG. F
OPERATIVE TEMPERATURE, DEG. F
EFFECTIVE RADIANT FIELD, BTU/HR.
A.U.S.T, DEG. F
PARAMETER 1, BTU/HR.SQ FT.F
PARAMETER 3, DIMENSIONLESS
BTU/HR.SQ FT
SQ FT
120.0
438.8
26762.4
23664.8
-11.6
25662.7
-4.1
25662.7
-4.1
23655.1
-11.6
95.3
48.8
53.9
74.4
66.7
77.0
72.4
7.6
68.7
1.0091
0.0168
PANEL HEATING - CEILING HEIGHT = 9 FT.
130.0
359.7
26762.4
23667.7
-11.6
25654.4
-4.1
25654.4
-4.1
24005.8
-10.3
95.3
40.0
66.7
74.7
66.7
77.0
72.4
7.5
68.6
1.0523
0.0142
140.0
302.2
26762.4
23671.1
-11.6
25650.1
-4.2
25650.1
-4.2
24263.4
-9.3
95.3
33.6
80.3
74.8
66.7
77.0
72.4
7.5
68.5
1.0937
0.0124
150.0
258.7
26762.4
23674.1
-11.5
25647.6
. -4.2
25647.6
-4.2
24459.6
-8.6
95.3
28.7
94.6
74.8
66.7
77.0
72.4
7.5
68.. 5
1.1338
0.0109
160.0
224.7
26762.4
23676.3
-11.5
25646.1
-4.2
25646.1
-4.2
24613.7
-8.0
95.4
25.0
109.6
74.9
66.7
77.0
72.4
7.5
68.4
1.1730
0.0098
170.0
197.9
26762.4
23678.0
-11.5
25645.2
-4.2
25645.2
-4.2
24735.2
-7.6
95.4
22.0
125.0
74.9
66.7
77.0
72.4
7.5
68.4
1.2111
0.0089
180.0
175.6
26762.4
23679.1
-11.5
25644.4
-4.2
25644.4
-4.2
24836.8
-7.2
95.4
19.5
141.5
74.9
66.7
77.0
72.4
7.5
68.4
1.2497
0.0081
calculation procedure since reasonable agreement is found with rated heating
panels.
The design heat loss calculated here (HLC) is about 4% below the ASHRAE
standard design heat loss calculation (HLD). This is attributed to the higher
wall, floor, and ceiling temperatures experienced in the radiant system than
in the forced air systems. For the radiant case the AUST remained at about 68
to 69 F and in the forced air system it ranged between 60 to 62 F, causing
additional heat loss through the surfaces. The room air temperature for com
fort conditions in the radiant cases (about 67 F) is 10 F less than in the
forced air case which reduces infiltration loss. However, this reduction in
infiltration loss does not overcome the increased loss due to higher surface
temperatures. Comparison with the forced air case shows about 3% more loss in
the radiant situation. This result is not significant in light of the many
assumptions made in both cases.
It should be noted in Table 13 that higher floor temperatures are present
in the radiant case than in the forced air case. This is significant since it
illustrates that the radiant systems heat surfaces which in turn heat the
occupants and the air while forced air systems heat the air which then heats
the occupants and the surfaces. Also, keep in mind that comfort conditions
were satisfied at the center location for a seated person and that due to
radiant temperature asymmetry discomfort could be experienced at the higher
panel temperatures. Normally, the higher panel temperatures would be used in
rooms with higher ceilings.
The values for floor temperature, room air temperature, mean radiant tem
perature, operative temperature, effective radient field and AUST remain rela
tively constant as the panel temperature increases. Parameter 1 and Parameter
3 were calculated just to observe their behavior in the radiant types of sys
tems. Parameter 1 is a "pseudo" overall heat transfer coefficient (defined by
-71-
475
i
I
175 120 130 140 150 160
panel temperature Cdeg. f)
170 180
FIGURE 19. REQUIRED HEATING PANEL AREA AS A FUNCTION OF PANEL TEMPERATURE
p^ m^ w^ m r-
TABLE 14. EFFECTS OF PANEL EMISSIVITY WITH PANEL
PANEL EMISSIVITY
PANEL AREA REQUIRED ,' SQ FT
ASHRAE DESIGN HEAT LOSS, BTU/HR
ACTUAL DESIGN HEAT LOSS, BTU/HR
PERCENTAGE DIFFERENCE 1
CONDUCTION DESIGN HEAT LOSS 1, BTU/HR
PERCENTAGE D1FFERENCE 2
CONDUCTION DESIGN HEAT LOSS 2, BTU/HR
PERCENTAGE DIFFERENCE 3
ACTUAL HEAT INPUT, BTU/HR
PERCENTAGE DIFFERENCE 4
PERCENTAGE RADIATION
PERCENT CEILING COVERED BY PANELS
HEAT OUTPUT PER UNIT PANEL AREA, BTU/HR.SQ FT
FLOOR TEMPERATURE, DEG. F
ROOM AIR TEMPERATURE, DEG. F
MEAN RADIANT TEMPERATURE, DEG. F
OPERATIVE TEMPERATURE, DEG. F
EFFECTIVE RADIANT FIELD, BTU/HR. SQ FT
A.U.S.T, DEG. F
PARAMETER 1, BTU/HR.SQ FT.F
PARAMETER 3, DIMENSIONLESS
0.88
316.8
26762.4
23628.0
-11.7
25654.1
-4.1
25654.1
-4.1
24203.7
-9.6
95.1
35.2
76.4
74.3
66.6
77.1
72.5
7.7
68.3
1.0392
0.0117
= 140°F
0.90
301.9
26762.4
23684.5
-11.5
25728.6
-3.9
25728.6
-3.9
24342.1
-9.0
95.3
33.5
80.6
74.5
66.7
77.0
72.4
7.5
68.5
1.0990
0.0124
0.92
288.3
26762.4
23735.9
-11.3
25796.5
-3.6
25796.5
-3.6
24468.7
-8.6
95.6
32.0
84.9
74.7
66.9
76.8
72.4
7.?
68.7
1.1590
0.0131
0.94
275.8
26762.4
23782.9
-11.1
25858.4
, -3.4
25858.4
-3.4
24585.0
-8.1
95.8
30.6
89.1
74.9
67.0
76.6
72.4
7.0
68.8
1.2194
0.0138
Eqn. 28) and Parameter 3 is a dimensionless factor related to the radiant
exchange process (defined by Eqn. 29). There does not appear to be any sig
nificant trends to either of these parameters.
Figure 19, shows the expected nonlinearity of the required panel area as a
function of panel temperature for the radiant base case given in Table 13.
Table 14 shows a comparison for the different assumptions concerning the
radiant panel emissivity. The emissivity was varied between 0.88 and 0.94 for
a panel temperature of 140 F. This resulted in significant changes in
required radiant panel area (a 13% drop in area as emissivity changed from
0.88 to 0.94). This variation in emissivity also affected the heat output per
unit area and parameters 1 and 3. Some of the other quantities showed only
slight changes as the emissivity was varied. Manufacturers indicate that a
panel emissivity of 0.9 is typical over the life of the radiant panel.
In Table 15, the emissivities of the walls, floor and ceiling were varied
between 0.8 and 0.95 for a situation when the panel temperature was at 140 F.
This caused the required panel area to increase by only 3%. The only other
variable to change significantly with this change in surface emissivity was
the floor temperature which went from 72 to 76 F as emissivity went from 0.8
to 0.95. For the remainder of the calculations, a value of 0.9 for the sur
face emissivities has been used and the calculations do not appear to be sen
sitive to changes in the surface emissivity.
Tables 16, 17 and 18 show the effects obtained when the convection
coefficient for the radiant panel is changed by a factor of 2, 5 and 10
respectively. This calculation was carried out since there is a great deal of
uncertainty concerning the value of the convection coefficient from surfaces
when there is a large delta T such as exists in the radiant panel case. Min
(18) has made this point in his work and indicated that it is a difficult
parameter to evaluate because of geometrical considerations.
-74-
15. EFFECTS OF WALL, FLOOR AND CEILING
WALL EMISSIVITY
PANEL AREA REQUIRED , SQ FT
ASHRAE DESIGN HEAT LOSS, BTU/HR
ACTUAL DESIGN HEAT LOSS, BTU/HR
PERCENTAGE DIFFERENCE 1
CONDUCTION DESIGN HEAT LOSS 1, BTU/HR
PERCENTAGE DIFFERENCE 2
CONDUCTION DESIGN HEAT LOSS 2, BTU/HR
PERCENTAGE DIFFERENCE 3
ACTUAL HEAT INPUT, BTU/HR
PERCENTAGE DIFFERENCE 4
PERCENTAGE RADIATION
PERCENT CEILING COVERED BY PANELS
HEAT OUTPUT PER UNIT PANEL AREA, BTU/HR.SQ FT
FLOOR TEMPERATURE, DEG. F
ROOM AIR TEMPERATURE, DEG. F
MEAN RADIANT TEMPERATURE, DEG. F
OPERATIVE TEMPERATURE, DEG. F
EFFECTIVE RADIANT FIELD, BTU/HR. SQ FT
A.U.S.T, DEG. F
PARAMETER 1, BTU/HR.SQ FT.F
PARAMETER 3, DIMENSIONLESS
VITY WITH PANEL TEMPERATURE = 140°F
0.80
295.8
26762.4
23700.0
-11.4
26303.1
-1.7
26303.1
-1.7
25015.2
-6.5
95.6
32.9
84.6
71.9
66.8
76.9
72.4
7.4
69.2
1.1531
0.0131
0.85
299.3
26762.4
23671.8
-11.5
25984.5
-2.9
25984.5
-2.9
24645.2
-7.9
95.4
33.3
82.3
73.4
66.7
77.0
72.4
7.5
68.9
1.1218
0.0127
0.90
302.2
26762.4
23671.1
-11.6
25650.1
-4.2
25650.1
-4.2
24263.4
-9.3
95.3
33.6
80.3
74.8
66.7
77.0
72.4
7.5
68.5
1.0937
0.0124
0.95
304.8
26762.4
23678.1
-11.5
25317.7
-5.4
25317.7
-5.4
23882.0
-10.8
95.2
33.9
78.4
76.0
66.7
77.0
72.4
7.5
68.1
1.0677
0.0120
TABLE 16. EFFECTS OF CHANGING PANEL CONVECTION
PANEL TEMPERATURE, DEC F
PANEL AREA REQUIRED , SQ FT
ASHRAE DESIGN HEAT LOSS, BTU/HR
ACTUAL DESIGN HEAT LOSS, BTU/HR
PERCENTAGE DIFFERENCE 1
CONDUCTION DESIGN HEAT LOSS 1, BTU/HR
PERCENTAGE DIFFERENCE 2
CONDUCTION DESIGN HEAT LOSS 2, BTU/HR
PERCENTAGE DIFFERENCE 3
ACTUAL HEAT INPUT, BTU/HR
PERCENTAGE DIFFERENCE 4
PERCENTAGE RADIATION
PERCENT CEILING COVERED BY PANELS
HEAT OUTPUT PER UNI.T PANEL AREA,
FLOOR TEMPERATURE, DEG. F
ROOM AIR TEMPERATURE, DEG. F
MEAN RADIANT TEMPERATURE, DEG. F
OPERATIVE TEMPERATURE, DEG. F
EFFECTIVE RADIANT FIELD, BTU/HR.
A.U.S.T, DEG. F
PARAMETER 1, BTU/HR.SQ FT.F
PARAMETER 3, DIMENSIONLESS
BTU/HR.SQ FT
SQ FT
120.0
418.7
26762.4
23827.2
-11.0
25637.2
-4.2
25637.2
-4.2
23719.9
-11.4
91.2
46.5
56.7
74.1
67.1
76.5
72.3
6.8
68.5
1.0691
0.0177
CIENT BY A FACTOR OF TWO
130.0
343.1
26762.4
23833.8
-10.9
25626.9
-4.2
25626.9
-4.2
24053.3
-10.1
91.1
38.1
70.1
74.3
67.1
76.4
72.3
6.8
68.4
1.1133
0.0150
140.0
288.2
26762.4
23839.8
-10.9
25620.5
-4.3
25620.5
-4.3
24297.1
-9.2
91.1
32.0
84.3
74.4
67.1
76.4
72.3
6.8
68.3
1.1554
0.0130
150.0
246.8
26762.4
23841.5
-10.9
25618.8
fc4.3
25618.8
-4.3
24484.9
-8.5
91.2
27.4
99.2
74.4
67.1
76.4
72.3
6.8
68.2
1.1961
0.0115
160.0
214.4
26762.4
23841.7
-10.9
25618.5
-4.3
25618.5
-4.3
24632.8
-8.0
91.2
23.8
114.9
74.5
67.1
76.4
72.3
6.8
68.2
1.2360
0.0103
170.0
188.9
26762.4
23841.0
-10.9
25618.9
-4.3
25618.9
-4.3
24749.8
-7.5
91.2
21.0
131.0
74.5
67.1
76.4
72.3
6.8
68.2
1.2747
0.0093
180.0
167.7
26762.4
23839.6
-10.9
25619.5
-4.3
25619.5
-4.3
24847.8
-7.2
91.3
18.6
148.2
74.5
67.1
76.4
72.3
6.8
68.2
1.3140
0.0085
TABLE 17. EFFECTS OF CHANGING PANEL
PANEL TEMPERATURE, DEC F
PANEL AREA REQUIRED , SQ FT
ASHRAE DESIGN HEAT LOSS, BTU/HR
ACTUAL DESIGN HEAT LOSS, BTU/HR
PERCENTAGE DIFFERENCE 1
CONDUCTION DESIGN HEAT LOSS 1, BTU/HR
PERCENTAGE DIFFERENCE 2
CONDUCTION DESIGN HEAT LOSS 2, BTU/HR
PERCENTAGE DIFFERENCE 3
ACTUAL HEAT INPUT, BTU/HR
PERCENTAGE DIFFERENCE 4
PERCENTAGE RADIATION
PERCENT CEILING COVERED BY PANELS
HEAT OUTPUT PER UNIT PANEL AREA, BTU/HR.SQ FT
FLOOR TEMPERATURE, DEG. F
ROOM AIR TEMPERATURE, DEG. F
MEAN RADIANT TEMPERATURE, DEG. F
OPERATIVE TEMPERATURE, DEG. F
EFFECTIVE RADIANT FIELD, BTU/HR. SQ FT
A.U.S.T, DEG. F
PARAMETER 1, BTU/HR.SQ FT.F
PARAMETER 3, DIMENSIONLESS
COEFFICIENTS BY A FACTOR OF FIVE
120.0
369.9
26762.4
24233.1
-9.5
25563.7
-4.5
25563.7
-4.5
23866.9
-10.8
81.1
41.1
64.5
73.3
68.2
75.1
72.0
5.0
67.8
1.2433
0.0204
130.0
302.9
26762.4
24236.0
-9.4
25561.6
-4.5
25561.6
-4.5
24170.0
-9.7
80.9
33.7
79.8
73.4
68.2
75.1
72.0
• 5.0
67.8
1.2893
0.0172
10.0
254.4
26762.4
24237.1
-9.4
25561.6
-4.5
25561.6
-4.5
24391.8
-8.9
80.9
28.3
95.9
73.5
68.2
75.1
72.0
5.0
67.7
1.3337
0.0149
150.0
217.8
26762.4
24236.6
-9.4
25562.5
-4.5
25562.5
-4.5
24560.1
-8.2
80.9
24.2
112.7
73.5
68.2
75.1
72.0
5.0
67.7
1.3768
0.0132
160.0
189.8
26762.4
24235.0
-9.4
25563.5
-4.5
25563.5
-4.5
24689.6
-7.7
80.9
21.1
130.1
73.6
68.2
75.1
72.0
5.0
67.7
1.4185
0.0118
170.0
166.9
26762.4
24232.3
-9.5
25564.7
-4.5
25564.7
-4.5
24795.9
-7.3
81.0
18.5
148.6
73.6
68.2
75.1
72.0
5.0
67.7
1.4606
0.0107
180.0
148.2
26762.4
24228.7
-9.5
25565.9
-4.5
25565.9
-4.5
24882.9
-7.0
81.0
16.5
167.9
73.6
68.2
75.1
72.0
5.0
67.7
1.5024
0.0097
TABLE 18. EFFECTS OF CHANGING PANEL CONVECTION
PANEL TEMPERATURE, DEC F
PANEL AREA REQUIRED , SQ FT
ASHRAE DESIGN HEAT LOSS, BTU/HR
ACTUAL DESIGN HEAT LOSS, BTU/HR
PERCENTAGE DIFFERENCE 1
CONDUCTION DESIGN HEAT LOSS 1, BTU/HR
PERCENTAGE DIFFERENCE 2
CONDUCTION DESIGN HEAT LOSS 2, BTU/HR
PERCENTAGE DIFFERENCE 3
ACTUAL HEAT INPUT, BTU/HR
PERCENTAGE DIFFERENCE 4
PERCENTAGE RADIATION
PERCENT CEILING COVERED BY PANELS
HEAT OUTPUT PER UNIT PANEL AREA,
FLOOR TEMPERATURE, DEC F
ROOM AIR TEMPERATURE, DEC F
MEAN RADIANT TEMPERATURE, DEC F
OPERATIVE TEMPERATURE, DEC F
EFFECTIVE RADIANT FIELD, BTU/HR.
A.U.S.T, DEC F
PARAMETER 1, BTU/HR.SQ FT.F
PARAMETER 3, DIMENSIONLESS
BTU/HR.SQ FT
SQ FT
120.0
313.6
26762.4
24671.3
-7.8
25513.7
-4.7
25513.7
-4.7
24071.3
-10.1
69.1
34.8
76.8
72.4
69.4
73.5
71.7
3.0
67.3
1.5134
0.0246
CIENTS BY A FACTOR OF TEN
130.0
256.0
26762.4
24678.6
-7.8
25514.2
-4.7
25514.2
-4.7
24335.3
-9.1
68.8
28.4
95.0
72.4
69.4
73.5
71.7
3.0
67.2
1.5657
0.0208
140.0
214.7
26762.4
24682.8
-7.8
25515.2
-4.7
25515.2
-4.7
24525.9
-8.4
68.6
23.9
114.2
72.5
69.4
73.5
71.7
3.0
67.2
1.6161
0.0180
150.0
184.1
26762.4
24684.4
-7.8
25516.3
' -4.7
25516.3
-4.7
24667.2
-7.8
68.5
20,5
134.0
72.5
69.4
73.5
71.7
3.0
67.2
1.6642
0.0159
160.0
159.9
26762.4
24683.9
-7.8
25517.4
-4.7
25517.4
-4.7
24779.7
-7.4
68.4
17.8
154.9
72.5
69.4
73.5
71.7
3.0
67.2
1.7122
0.0142
170.0
140.6
26762.4
24681.6
-7.8
25518.5
-4.6
25518.5
-4.6
24869.6
-7.1
68.5
15.6
176.8
72.5
69.4
73.5
71.7
3.0
67.2
1.7596
0.0128
180.0
124.9
26762.4
24677.9
-7.8
25519.6
-4.6
25519,6
-4.6
24943.0
-6.8
68.6
13.9
199.6
72.5
69.4
73.5
71.7
3.0
67.2
1.8065
0.0117
By increasing the panel convection coefficient by 100% (Table 16), the
area required for heating the space is reduced by 4.5% for a 150 F panel tem
perature. By increasing the convection coefficient by 500% (Table 17) the
area is reduced by 15.8%. Therefore, it is not a significant variation if the
convection coefficient is known within a factor of two. It is interesting to
note that the percent difference in design heat losses remains about the same
(-4%) for all of these cases. Some of the other parameters change slightly
with significant changes in panel convection coefficient. The percent radia
tion delivered by the panels does change as the convection coefficient is
increased and this variation is illustrated in Figure 20.
5.5.2 - Effect Due to Infiltration for Radiant Panel Systems. Different
values of infiltration rates (0.5 to 4.0 air changes per hour) were assumed
for the base case configuration and these results are given in Tables 19, 20
and 21 for 130, 150, and 170 F panel surface temperatures respectively.
As seen in these tables the ASHRAE standard design heat loss (HID)
overpredicts the calculated design heat loss (HLC) by up to 16% for an
infiltration rate of 4 air changes per hour. If this were compared to a
forced air system with an air temperature gradient, approximately one percent
more loss would be added to this number (See Section 5.4) so that there might
be a difference of approximately 17%. The percent difference in the design
loads as a function of infiltration is shown in Figure 21.
As the infiltration rate increases, the floor temperature, mean radiant
temperature, operative temperature, effective radiant flux and AUST increase
significantly, while the room air temperature and parameters 1 and 3 decrease
significantly. These changes need to be considered in the design process for
radiant panel systems.
5.5.3 - Effect of Glass Distribution. Different combinations and quantities
-79-
100
I 00 o i
C 0
•• I +>
. o ••-I
•n a
ex. QI
o +J c QI U L a a.
90 -
80 -
70 -
60 4 6
Convection multiplier
FIGURE 20. EFFECTS ON PERCENT RADIATION DELIVERED BY THE PANEL AS THE CONVECTION MULTIPLIER IS CHANGED
TABLE 19. EFFECTS DUE TO CHANGING AIR INFILTRATION RATES FOR A PANEL TEMPERATURE OF 130°F
INFILTRATION AIR, AC/H
PANEL AREA REQUIRED , SQ FT
ASHRAE DESIGN HEAT LOSS, BTU/HR
ACTUAL DESIGN HEAT LOSS, BTU/HR
PERCENTAGE DIFFERENCE 1
CONDUCTION DESIGN HEAT LOSS 1, BTU/HR
PERCENTAGE DIFFERENCE 2
CONDUCTION DESIGN HEAT LOSS 2, BTU/HR
PERCENTAGE DIFFERENCE 3
ACTUAL HEAT INPUT, BTU/HR
PERCENTAGE DIFFERENCE 4.
PERCENTAGE RADIATION
PERCENT CEILING COVERED BY PANELS
HEAT OUTPUT PER UNIT PANEL AREA, BTU/HR.SQ FT
FLOOR TEMPERATURE, DEG. F
ROOM AIR TEMPERATURE, DEG. F
MEAN RADIANT TEMPERATURE, DEG. F
OPERATIVE TEMPERATURE, DEG. F
EFFECTIVE RADIANT FIELD, BTU/HR. SQ FT
A.U.S.T, DEG. F
PARAMETER 1, BTU/HR.SQ FT.F
PARAMETER 3, DIMENSIONLESS
0.50
358.7
26762.1
23696.6
-11.5
25815.7
-3.5
25815.7
-3.5
24167.9
-9.7
95.4
39.9
67.4
74.1
66.8
76.9
72.4
7.4
68.6
1.0638
0.0144
0.75
394.4
29386.8
25636.3
-12.8
28174.2
-4.1
28174.2
-4.1
26364.7
-10.3
95.2
43.8
66.9
74.9
65.8
78.1
72.7
9.0
69.1
1.0400
0.0142
1.00
429.6
32011.2
27497.4
-14.1
30481.1
-4.8 t
30481.1 •
-4.8
28501.6
-11.0
95.1
47.7
66.3
75.6
64.8
79.3
72.9
10.6
69.7
1.0167
0.0139
1.50
498.1
37260.0
31060.9
-16.6
34889.7
-6.4
34889.7
-6.4
32575.4
-12.6
94.9
55.3
65.4
76.9
63.0
81.6
73.4
13.6
70.8
0.9751
0.0136
2.00
564.0
42508.8
34419.8
-19.0
39056.3
-8.1
39056.3
-8.1
36417.1
-14.3
94.6
62.7
64.6
78.1
61.3
83.7
73.8
16.4
71.8
0.9385
0.0133
3.00
687.6
53006.4
40625.3
-23.4
46706.8
-11.9
46706.8
-11.9
43453.4
-18.0
94.2
76.4
63.2
80.1
58.2
87.4
74.5
21.4
73.7
0.8788
0.0128
4.00
799.6
63504.0
46304.3
-27.1
53552.2
-15.7
53552.2
-15,7
49742.5
-21.7
93.8
88.8
62.2
81.5
55.5
90.5
75.0
25.7
75.4
0.8340
0.0125
TABLE 20. EFFECTS DUE TO CHANGING AIR INFILTRATION RATES FOR A PANEL TEMPERATURE OF 150°F
INFILTRATION AIR, AC/H
PANEL AREA REQUIRED , SQ FT
ASHRAE DESIGN HEAT LOSS, BTU/HR
ACTUAL DESIGN HEAT LOSS, BTU/HR
PERCENTAGE DIFFERENCE 1
CONDUCTION DESIGN HEAT LOSS 1, BTU/HR
PERCENTAGE DIFFERENCE 2
CONDUCTION DESIGN HEAT LOSS 2, BTU/HR
PERCENTAGE DIFFERENCE 3
ACTUAL HEAT INPUT, BTU/HR
PERCENTAGE DIFFERENCE 4
PERCENTAGE RADIATION
PERCENT CEILING COVERED BY PANELS
HEAT OUTPUT PER UNIT PANEL AREA,
FLOOR TEMPERATURE, DEG. F
ROOM AIR TEMPERATURE, DEG. F
MEAN RADIANT TEMPERATURE, DEG. F
OPERATIVE TEMPERATURE, DEG. F
EFFECTIVE RADIANT FIELD, BTU/HR.
A.U.S.T, DEG. F
PARAMETER 1, BTU/HR. SQ FT..F
PARAMETER 3; DIMENSIONLESS
BTU/HR.SQ FT
SQ FT
0.50
259.2
26762.4
23697.8
-11.5
25813.3
-3.5
25813.3
-3.5
24621.4
-8.0
95.4
28.8
95.0
74.2
66.8
76.9
72.4
7.4
68.5
1.1425
0.0110
0.75
284.4
29386.8
25644.9
-12.7
28172.3
-4.1
28172.3
-4.1
26865.6
-8.6
95.3
31.6
94.5
74.9
65.8
78.1
72.7
8.9
69.0
1.1236
0.0109
1.00
308.4
32011.2
27505.0
-14.1
30483.2
-4.8
30483.2
-4.8
29058.7
-9.2
95.2
34.3
94.2
75.7
64.9
79.3
72.9
10.5
69.5
1.1053
0.0108
1.50
356.5
37260.0
31072.9
-16.6
34894.7
-6.3
34894.7
-6.3
33233.2
-10.8
95.0
39.6
93.2
77.1
63.0
81.5
73.4
13.5
70.5
1.0708
0.0105
2.00
402.7
42508.8
34435.0
-19.0
39064.9
-8.1
39064.9
-8.1
37173.2
-12.6
94.8
44.7
92.3
78.4
61.3
83.6
73.8
16.3
71.5
1.0398
0.0103
3.00
489.1
53006.4
40640.1
-23.3
46715.0
-11.9
46715.0
-11.9
44388.4
-16.3
94.5
54.3
90.8
80.7
58.2
87.4
74.5
21.4
73.2
0.9875
0.0100
4.00
567.7
63504.0
46297.4
-27.1
53550.0
-15.7
53550.0
-15.7
50826.4
-20.0
94.2
63.1
89.5
82.6
55.5
90.5
75.1
25.7
74.6
0.9464
0.0097
TABLE 21. EFFECTS DUE TO CHANGING AIR INFILTRATION RATES FOR A PANEL TEMPERATURE AT 170°F
INFILTRATION AIR, AC/H
PANEL AREA REQUIRED , SQ FT
ASHRAE DESIGN HEAT LOSS, BTU/HR
ACTUAL DESIGN HEAT LOSS, BTU/HR
PERCENTAGE DIFFERENCE 1
CONDUCTION DESIGN HEAT LOSS 1, BTU/HR
PERCENTAGE DIFFERENCE 2
CONDUCTION DESIGN HEAT LOSS 2, BTU/HR
PERCENTAGE DIFFERENCE 3
ACTUAL HEAT INPUT, BTU/HR
PERCENTAGE DIFFERENCE 4
PERCENTAGE RADIATION
PERCENT CEILING COVERED BY PANELS
HEAT OUTPUT PER UNIT PANEL AREA, BTU/HR.SQ FT
FLOOR TEMPERATURE, DEG. F
ROOM AIR TEMPERATURE, DEG. F
MEAN RADIANT TEMPERATURE, DEG. F
OPERATIVE TEMPERATURE, DEG. F
EFFECTIVE RADIANT FIELD, BTU/HR. SQ FT
A.U.S.T, DEG. F
PARAMETER 1, BTU/HR.SQ FT.F
PARAMETER 3, DIMENSIONLESS
0.50
198.0
26762.4
23696.5
-11.5
25813.7
-3.5
25813.7
-3.5
24902.2
-7.0
95.4
22.0
125.7
74.2
66.8
76.9
72.4
7.4
68.5
1.2189
0.0090
0.75
217.0
29386.8
25647.3
-12.7
28173.3
-4.1
28173.3
-4.1
27175.0
-7.5
95.3
24.1
125.2
75.0
65.8
78.1
72.7
8.9
68.9
1.2031
0.0089
1.00
235.8
32011.2
27507.8
-14.1
30485.6
-4t8
30485.6
-4.8
29395.0
-8.2
95.3
26.2
124.6
75.7
64.9
79.3
72.9
10.5
69.4
1.1868
0.0088
1.50
272.0
37260.0
31079.0
-16.6
34899.1
, -6.3
•34899.1
-6.3
33629.3
-9.7
95.1
30.2
123.7
77.2
63.1
81.5
73.4
13.5
70.4
1.1572
0.0086
2.00
305.9
42508.8
34445.2
-19.0
39072.5
-8.1
39072.5
-8.1
37632.1
-11.5
95.0
34.0
123.0
78.5
61.3
83.6
73.8
16.3
71.4
1.1312
0.0085
3.00
370.5
53006.4
40658.5
-23.3
46729.8
-11.8
46729.8
-11.8
44961.2
-15.2
94.7
41.2
121.4
80.9
58.2
87.3
74.5
21.4
73.0
1.0850
0.0082
4.00
429.0
63504.0
46319.8
-27.1
53570.1
-15.6
53570.1
-15.6
51501.8
-18.9
94.5
47.7
120.1
82.9
55.5
,90.5
75.0
25.7
74.3
1.0478
0.0080
I 00 I
0}
01 U c. 0) L at
07 a +» c ai u L <J Q.
-12 -
- IS -
-18
i n f i l t r a t i o n Cach)
FIGURE 2 1 . EFFECT OF A I R I N F I L T R A T I O N I N PANEL HEATING ON PERCENT DIFFERENCE I N DESIGN LOAD CALCULATIONS
wm rx-3 r—5 '• -•* r ^
TABLE 22. EFFECTS OF CHANGES IN GLASS DISTRIBUTION - PANEL TEMPERATURE = 140°F
GLASS DISTRIBUTION CASE NUMBER
PANEL AREA REQUIRED , SQ FT
ASHRAE DESIGN HEAT LOSS, BTU/HR
ACTUAL DESIGN HEAT LOSS, BTU/HR
PERCENTAGE DIFFERENCE 1
CONDUCTION DESIGN HEAT LOSS 1, BTU/HR
PERCENTAGE DIFFERENCE 2
CONDUCTION DESIGN HEAT LOSS 2, BTU/HR
PERCENTAGE DIFFERENCE 3
ACTUAL HEAT INPUT, BTU/HR
PERCENTAGE DIFFERENCE 4
PERCENTAGE RADIATION
PERCENT CEILING COVERED BY PANELS
HEAT OUTPUT PER UNIT PANEL AREA,
FLOOR TEMPERATURE, DEG. F
ROOM AIR TEMPERATURE, DEG. F
MEAN RADIANT TEMPERATURE, DEG. F
OPERATIVE TEMPERATURE, DEG. F
EFFECTIVE RADIANT FIELD, BTU/HR.
A.U.S.T, DEG. F
PARAMETER 1, BTU/HR.SQ FT.F
PARAMETER 3,.DIMENSIONLESS
BTU/HR.SQ FT
SQ FT
1
249.0
22096.8
19744.9
-10.6
21110.2
-4.5
21110.2
-4.5
19954.8
-9.7
95.4
27.7
80.1
73.8
67.3
76.2
72.3
6.4
69.4
1.1013
0.0126
2
302.2
26762.4
23671.1
-11.6
25650.1
-4.2
25650.1
-4.2
24263.4
-9.3
95.3
33.6
80.3
74.8
66.7
77.0
72.4
7.5
68.5
1.0937
0.0124
3
359.3
31428.0
27407.1
-12.8
30469.1
-3.1
30469.1
-3.T
28839.3
-8.2
95.2
39.9
80.3
75.8
65.8
78.1
72.7
9.0
67.8
1.0803
0.0121
4
410.9
36093.6
31189.3
-13.6
34945.1
-3.2
34945.1
. -3.2
33107.3
-8.3
95.2
45.7
80.6
76.6
65.2
78.8
72.8
10.0
66.8
1.0761
0.0119
5
466.5
40759.2
34757.8
-14.7
39796.1
-2.4
39796.1
-2.4
37737.7
-7.4
95.2
51.8
80.9
77.4
64.4
79.9
73.0
11.3
66.0
1.0687
0.0117
Case number T: No glass In any wall Case number 2: One w a l l , ha l f glass Case number 3: One w a l l , a l l glass Case number 4: One w a l l , a l l glass-second w a l l , ha l f glass Case number 5: Two w a l l s , a l l glass
of glass have been considered in the radiant base case which was previously
described. The results from these calculations are given in Table 22 for a
panel temperature of 140 F. The radiant base case is shown as Case 2 in Table
22 and a room with no glass is given as Case 1. Case 3 is one wall which is
all glass, Case 4 is one wall with all glass and half of another wall with
glass, and Case 5 is the room with two walls all glass. As anticipated, as
the quantity of glass increases the panel area increases. Also, as can be
seen in Table 22, the difference between HID and HLC becomes smaller as the
quantity of glass in the room increases with only a -2% difference showing up
in Case 5. Since the panel area increases in order to make up for increased
heat losses as the quantity of glass is increased, the floor temperature also
rises. This in turn causes the room air temperature for comfort to be reduced
from 67 to 64 F. In Figure 22, the required panel area is plotted for each
case shown in Table 22. Likewise, in Figure 23 the floor temperature is
plotted. It is interesting to note that an 87% increase in panel area in the
room results in a only a 3.6 F increase in floor temperature.
5.5.4 - Changes in Wall. Floor and Ceiling U-Factors. The U-Factors in the
radiant base case were changed and various calculations were made to determine
this effect on the design heat loss. The wall U-factors were changed from 0.1
to 0.2 Btu/hr ft2 F and the floor and ceiling values from 0.07 to 0.1 Btu/hr
ft2 F. The U-factor for the glass was changed from 0.58 to 1.0 Btu/hr ft2 F.
All of these were changed at one time so that an initial and a new case were
compared at three panel temperatures of 130, 150 and 170 F. These results are
given in Table 23. The variation is very much as expected in that with
increased U-factors there is an increased heat loss and greater panel area
required. This results in an increased floor temperature and MRT and a
decreased air temperature and AUST. In each case, the difference between HID
and HLC has been reduced by about one-half so that the ASHRAE standard design
-86-
m^i r* K e ^ ss-s-.i-'a ey
I oo
I
a en
UJ
z
4
CASE NUMBER
FIGURE 22. PANEL AREA REQUIRED AS A FUNCTION OF THE QUANTITY OF GLASS
I 00 oo I
O 111 Q
UJ
a:
UJ Q.
2
cc o o
80.0 -r
79.0 -
78.0 -
77.0 -
76.0 -
75.0 -
74.0 -.
73.0 -
72.0 -
71.0 -
70.0 -• 1 2
T " 3
T " 4
CASE NUMBER
FIGURE 23. FLOOR TEMPERATURE AS A FUNCTION OF THE QUANTITY OF GLASS
§'••'• •>•:••: s t< '.V! » - . - • - *
TABLE 23. EFFECTS OF INCREASED U-FACTORS ON RADIANT HEATING PANEL PERFORMANCE
i 00
I
U-FACTOR CASE OLD NEW OLD NEW OLD NEW
PANEL TEMPERATURE, DEC F 590.0 590.0 610.0 610.0 630.0 630.0
PANEL AREA REQUIRED , SQ FT 359.7 546.3 258.7 397.3 197.9 305.3
ASHRAE DESIGN HEAT LOSS, BTU/HR 26762.4 41536.8 26762.4 41536.8 26762.4 41536.8
ACTUAL DESIGN HEAT LOSS, BTU/HR 23667.7 35605.8 23674.1 35622.0 23678.0 35611.4
PERCENTAGE DIFFERENCE 1 -11.6 -14.3 -11.5 -14.2 -11.5 -14.3
CONDUCTION DESIGN HEAT LOSS 1, BTU/HR 25654.4 40542.8 25647.6 40403.9 25645.2 40374.7
PERCENTAGE DIFFERENCE 2 -4.1 -2.4 -4.2 -2.7 -4.2 -2.8
CONDUCTION DESIGN HEAT LOSS 2, BTU/HR 25654.4 40542.8 25647.6 40403.9 25645.2 40374.7
PERCENTAGE DIFFERENCE 3 -4.1 -2.4 -4.2 " -2.7 -4.2 -2.8
ACTUAL HEAT INPUT, BTU/HR 24005.8 37107.6 24459.6 37890.3 24735.2 38435.0
PERCENTAGE DIFFERENCE 4 -10.3 -10.7 -8.6 -8.8 -7.6 -7.5
PERCENTAGE RADIATION 95.3 95.2 95.3 95.2 95.4 95.3
PERCENT CEILING COVERED BY PANELS 40.0 60.7 28.7 44.1 22.0 33.9
HEAT OUTPUT PER UNIT PANEL AREA, BTU/HR.SQ FT 66.7 67.9 94.6 95.4 125.0 125.9
FLOOR TEMPERATURE, DEG. F 74.7 75.6 74.8 76.4 74.9 76.7
ROOM AIR TEMPERATURE, DEG. F 66.7 64.7 66.7 64.7 66.7 64.7
MEAN RADIANT TEMPERATURE, DEG. F 77.0 79.5 77.0 79.4 77.0 79.5
OPERATIVE TEMPERATURE, DEG. F 72.4 72.9 72.4 72.9 72.4 72.9
EFFECTIVE RADIANT FIELD, BTU/HR. SQFT 7.5 10.8 7.5 10.7 7.5 10.8
A.U.S.T, DEG. F 68.6 66.3 68.5 65.9 68.4 65.7
PARAMETER 1, BTU/HR.SQ FT.F 1.0523 1.0390 1.1338 1.1176 1.2111 1.1947
PARAMETER 3, DIMENSIONLESS 0.0142 0.0136 0.0109 0.0105 0.0089 0.0086
SURFACE INITIAL U NEW U
WALL 1 0.34 0.60
WALL 2 0.1 0.2
WALL 3 0.1 0.2
WALL 4 0.1 0.2
FLOOR 0.07 0.10
CEILING 0.07 0.10
TABLE 24. EFFECTS OF CHANGES IN ROOM LENGTH AND WIDTH WITH A 140°F PANEL TEMPERATURE
ROOM LENGTH*WIDTH, FT*FT
PANEL AREA REQUIRED , SQ FT
ASHRAE DESIGN HEAT LOSS, BTU/HR
ACTUAL DESIGN HEAT LOSS, BTU/HR
PERCENTAGE DIFFERENCE 1
CONDUCTION DESIGN HEAT LOSS 1, BTU/HR
PERCENTAGE DIFFERENCE 2
CONDUCTION DESIGN HEAT LOSS 2, BTU/HR
PERCENTAGE DIFFERENCE 3
ACTUAL HEAT INPUT, BTU/HR
PERCENTAGE DIFFERENCE 4
PERCENTAGE RADIATION
PERCENT CEILING COVERED BY PANELS
HEAT OUTPUT PER UNIT PANEL AREA, BTU/HR.SQ
FLOOR TEMPERATURE, DEG. F
ROOM AIR TEMPERATURE, DEG. F
MEAN RADIANT TEMPERATURE, DEG. F
OPERATIVE TEMPERATURE, DEG. F
EFFECTIVE RADIANT FIELD, BTU/HR. SQ FT
A.U.S.T, DEG. F
PARAMETER 1, BTU/HR.SQ FT.F
PARAMETER 3, DI MENS I ONLESS
20*20
162.6
14659.2
12901.9
-12.0
14121.6
-3.7
14121.6
-3.7
13384.0
-8.7
94.9
40.7
82.3
73.7
66.4
77.4
72.5
8.1
67.9
1.1191
0.0126
30*30
302.2
26762.4
23671..1
-11.6
25650.1
-4.2
25650.1
-4.2
24263.4
-9.3
95.3
33.6
80.3
74.8
66.7
77.0
72.4
7.5
68.5
1.0937
0.0124
40*40
468.9
42048.0
36820.0
-12.4
39711.4
-5.6
39711.4
-5.6
37571.6
-10.6
95.6
29.3
80.1
74.0
66.0
77.8
72.6
8.6
68.0
1.0822
0.0122
40*20
300.9
26726.4
23579.7
-11.8
25826.0
' -3.4
25826.0
-3.4
24454.1
-8.5
95.2
37.6
81.3
74.6
66.5
77.2
72.5
7.8
68.0
1.1047
0.0124
30*15
197.2
17658.0
15571.1
-11.8
17178.8
-2.7
17178.8
-2.7
16284.6
-7.8
94.9
43.8
82.6
74.3
66.5
77.2
72.5
7.9
67.8
1.1217
0.0126
procedure more closely predicts the required heat input.
5.5.5 - Effect of Changes in Room Length and Width. In Table 24, are the
results for five different size rooms: 20 ft. x 20 ft., 30 ft. x 30 ft., 40
ft. x 40 ft., 40 ft. x 20 ft., and 30 ft. x 15 ft. all 9 ft. high. The impor
tant thing to notice about these results is that as the room size increases
the ASHRAE standard design load procedure tends to increasingly over predict
the required heater size. This tendency is not great (3.7% for 20 x 20 and
5.6% for 40 x 40), but it is an important trend.
Also illustrated here is the fact that a square room or building will tend
to be over sized if the ASHRAE standard design load is used. In Table 25,
results for four square buildings are tabulated. Again, as the building
becomes larger the ASHRAE standard design procedure (HLD) tends to oversize
(6% for a 10,000 sq. ft. building ) the radiant heating system.
5.5.6 - Changes in Room Height. The radiant base case room was modified to
have a ceiling height between eight and twenty five feet. The results from
these calculations are presented in Table 26. There are two important trends
to observe from these results. First, as the height is increased, more panel
area is required to counteract the increased room heat loss and because of
changing room geometry, more of the walls intercept the radiant energy thus
increasing the AUST. This in turn causes the second trend to occur in that
the difference between HLD and HLC decreases because of more heat conduction
through the walls. This decrease in the difference between HLD and HLC as
room height is increased is illustrated by the plot shown in Figure 24.
5.5.7 - Changes in Outside Design Temperature. In order to see what effect
outside design temperature had on the design load calculation; five other
outside design temperatures were used and these results are given in Table 27.
There is a slight tendency for the percent difference between HLD and HLC to
increase with milder climates (-3.6% at -5 F to -5.3% at 15 F) . This does not
-91-
TABLE 25. EFFECTS OF CHANGING ROOM SIZE WITH A 140°F PANEL TEMPERATURE
ROOM LENGTH*WIDTH, FT*FT
PANEL AREA REQUIRED , SQ FT
ASHRAE DESIGN HEAT LOSS, BTU/HR
ACTUAL DESIGN HEAT LOSS, BTU/HR
PERCENTAGE DIFFERENCE 1
CONDUCTION DESIGN HEAT LOSS 1, BTU/HR
PERCENTAGE DIFFERENCE 2
CONDUCTION DESIGN HEAT LOSS 2, BTU/HR
PERCENTAGE DIFFERENCE 3
ACTUAL HEAT INPUT, BTU/HR
PERCENTAGE DIFFERENCE 4
PERCENTAGE RADIATION
PERCENT CEILING COVERED BY PANELS
HEAT OUTPUT PER UNIT PANEL AREA, BTU/HR
FLOOR TEMPERATURE, DEG. F
ROOM AIR TEMPERATURE, DEG. F
MEAN RADIANT TEMPERATURE, DEG. F
OPERATIVE TEMPERATURE, DEG. F
EFFECTIVE RADIANT FIELD, BTU/HR. SQ FT
A.U.S.T, DEG. F
PARAMETER 1, BTU/HR.SQ FT.F
PARAMETER 3, DIMENSIONLESS
20*20
162.6
14659.2
12901.9
-12.0
14121.6
-3.7
14121.6
-3.7
13384.0
-8.7
94.9
40.7
82.3
73.7
66.4
77.4
72.5
8.1
67.9
1.1191
0.0126
30*30
302.2
26762.4
23671.1
-11.6
25650.1
-4.2
25650.1
-4.2
24263.4
-9.3
95.3
33.6
80.3
74.8
66.7
77.0
72.4
7.5
68.5
1.0937
0.0124
40*40
468.9
42048.0
36820.0
-12.4
39711.4
-5.6
39711.4
-5.6
37571.6
-10.6
95.6
29.3
80.1
74.0
66.0
77.8
72.6
8.6
68.0
1.0822
0.0122
100*100
2285.8
200952.0
176195.8
-12.2
188495.6
-6.0
188495.6
" -6.0
177908.1
-11.3
96.4
22.9
77.8
73.7
66.2
77.6
72.5
8.3
69.0
1.0566
0.0120
TABLE 26. EFFECTS OF CHANGING ROOM HEIGHT WITH A
ROOM HEIGHT, FT
PANEL AREA REQUIRED , SQ FT
ASHRAE DESIGN HEAT LOSS, BTU/HR
ACTUAL DESIGN HEAT LOSS, BTU/HR
PERCENTAGE DIFFERENCE 1
CONDUCTION DESIGN HEAT LOSS 1, BTU/HR
PERCENTAGE DIFFERENCE 2
CONDUCTION DESIGN HEAT LOSS 2, BTU/HR
PERCENTAGE DIFFERENCE 3
ACTUAL HEAT INPUT, BTU/HR
PERCENTAGE DIFFERENCE 4
PERCENTAGE RADIATION
PERCENT CEILING COVERED BY PANELS
HEAT OUTPUT PER UNIT PANEL AREA,
FLOOR TEMPERATURE, DEC. F
ROOM AIR TEMPERATURE, DEG. F
MEAN RADIANT TEMPERATURE, DEG. F
OPERATIVE TEMPERATURE, DEG. F
EFFECTIVE RADIANT FIELD, BTU/HR.
A.U.S.T, DEG. F
PARAMETER 1, BTU/HR.SQ FT.F
PARAMETER 3, DIMENSIONLESS
BTU/HR.SQ FT
SQ FT
8.0
274.2
24796.8
21765.7
-12.2
23528.2
-5.1
23528.2
-5.1
22280.8
-10.1
95.3
30.5
81.2
73.5
66.2
77.6
72.6
8.3
67.7
1.0995
0.0123
°F PANEL TEMPERATURE
9.0
301.9
26762.4
23684.5
-11.5
25728.6
-3.9
25728.6
-3.9
24342.1
-9.0
95.3
33.5
80.6
74.5
66.7
77.0
72.4
7.5
68.5
1.0990
0.0124
10.0
323.6
28728.0
25332.6
-11.8
27642.7
-3.8
27642.7
-3.8
26159.7
-8.9
95.3
36.0
80.8
74.6
66.5
77.2
72.5
7.9
68.6
1.0983
0.0124
12.0
367.8
32659.2
28675.0
-12.2
31529.6
, -3.5
31529.6
-3.5
29847.7
-8.6
95.3
40.9
81.1
74.7
66.2
77.6
72.6
8.3
68.7
1.0983
0.0124
15.0
432.0
38556.0
33584.5
-12.9
37271.5
-3.3
37271.5
-3.3
35307.3
-8.4
95.3
48.0
81.7
74.4
65.7
78.2
72.7
9.1
68.8
1.0989
0.0124
20.0
556.1
48384.0
42498.0
-12.2
47707.2
-1.4
47707.2
-1.4
45149.8
-6.7
95.3
61.8
81.2
74.8
66.2
77.6
72.5
8.3
70.1
1.0994
0.0126
25.0
677.8
58212.0
51235.5
-12.0
57951.5
-0.4
57951.5
-0.4
54824.8
-5.8
95.3
75.3
80.9
74.6
66.4
77.4
72.5
8.1
70.8
1.0971
0.0127
I
L H s +J CD
W
0
63000
58000 -
53000
48000 -
43000 -
HJ 38000 o a x
33000 h
28000
23000
A = HLD - ASHRAE Design Heat Loss
0 = HLC - Conduction Design Heat Loss
10 15 20 25
Room height (ft)
FIGURE 24. EFFECT OF ROOM HEIGHT CHANGE ON DESIGN HEAT LOSS FOR RADIANT PANELS
TABLE 27. EFFECTS DUE TO OUTSIDE DESIGN TEMPERATURE CHANGES WITH A 130°F PANEL TEMPERATURE
OUTSIDE DESIGN TEMPERATURE, DEC F
PANEL AREA REQUIRED , SQ FT
ASHRAE DESIGN HEAT LOSS, BTU/HR
ACTUAL DESIGN HEAT LOSS, BTU/HR
PERCENTAGE DIFFERENCE 1
CONDUCTION DESIGN HEAT LOSS 1, BTU/HR
PERCENTAGE DIFFERENCE 2
CONDUCTION DESIGN HEAT LOSS 2, BTU/HR
PERCENTAGE DIFFERENCE 3
^ ACTUAL HEAT INPUT, BTU/HR
I PERCENTAGE DIFFERENCE 4
PERCENTAGE RADIATION
PERCENT CEILING COVERED BY PANELS
HEAT OUTPUT PER UNIT PANEL AREA, BTU/HR.SQ FT
FLOOR TEMPERATURE, DEG. F
ROOM AIR TEMPERATURE, DEG. F
MEAN RADIANT TEMPERATURE, DEG. F
OPERATIVE TEMPERATURE, DEG. F
EFFECTIVE RADIANT FIELD, BTU/HR. SQ FT
A.U.S.T, DEG. F
PARAMETER 1, BTU/HR.SQ FT.F
PARAMETER 3, DIMENSIONLESS
-5.0
399.0
29736.0
26437.8
-11.1
28675.7
-3.6
28675.7
-3.6
26634.3
-10.4
95.3
44.3
66.8
75.0
66.1
77.7
72.6
8.5
68.4
1.0437
0.0141
0.0
374.5
27877.5
24706.4
-11.4
26787.3
-3.9
26787.3
-3.9
24996.0
-10.3
95.3
41.6
66.7
74.8
66.5
77.3
72.5
7.9
68.5
1.0491
0.0142
3.0
359.7
26762.4
23667.7
-11.6
25654.4
-4.1
25654.4
-4.1
24005.8
-10.3
95.3
40.0
66.7
74.7
66.7
77.0
72.4
7.5
68.6
1.0523
0.0142
5.0
349.7
26019.0
22975.2
-11.7
24899.4
-4.3
24899.4
-4.3
23343.1
-10.3
95.3
38.9
66.7
74.6
66.8
76.8
72.4
7.3
68.7
1.054.6
0.0143
10.0
324.8
24160.5
21243.8
-12.1
23012.1
-4.8
23012.1
-4.8
21675.5
-10.3
95.3
36.1
66.7
74.4
67.2
76.4
72.3
6.8
68.8
1.0604
0.0144
15.0
299.5
22302.0
19512.3
-12.5
21125.4
-5.3
21125.4
-5.3
19992.8
-10.4
95.4
33.3
66.7
74.1
67.5
76.0
72.2
6.2
69.0
1.0663
0.0145
appear to be a significant trend, and is apparently due to reduced
infiltration and conduction losses at the higher outside temperatures.
5.5.8 - Changes in Number of Panels. All of the calculations presented to
this point have been for a single panel located in the center of the space or
room at ceiling height. The radiant base case was used again except that 2,4,
and 6 equal area panels were used for supplying the radiant heat to the room.
This was done to determine what effect panel distribution has on the design
heating load. The results from these calculations are given in Table 28. The
panel area required and difference in design heating loads did not change sig
nificantly. It is interesting to note that the floor temperature did drop by 1
F in going from one to six panels and the AUST increased by a slight (less
than 1 F) amount. This is due to the walls intercepting more of the radiant
energy in the 6 panel case than in the single panel case due to changing angle
factors.
5.5.9 - Perimeter Panel System. Several cases were run with a perimeter (nar
row panel running parallel to the outside wall) radiant panel system and these
results are given in Table 29. These cases are for a 15' x 15' x 8' room with
three inside walls and one outside wall with half glass. The U-factor for the
floor and ceiling were the same as the radiant base case. The radiant panel
was 36"wide and ran parallel to the outside wall with the window. There is no
apparent difference between HID and HLC as far as the design loads are con
cerned. This is apparently due to the proximity of the radiant heating sur
face to the cold surface or wall resulting in higher convective losses. This
is only about 4% different than the results shown in Table 13 for the single
panel radiant base case. This is not a significant trend considering all of
the unknown variables that can enter into consideration in the actual case.
The other variables in Tables 13 and 29 are quite similar so that this special
type of application of panels does not alter the conclusions from the single
-96-
„sg g p : " ^ p ; ' ^ p.-"-v>a f-
CHANGES AS A RESULT OF THE NUMBER OF RADIANT HE
NUMBER OF PANELS
PANEL AREA REQUIRED , SQ FT
ASHRAE DESIGN HEAT LOSS, BTU/HR
ACTUAL DESIGN HEAT LOSS, BTU/HR
PERCENTAGE DIFFERENCE 1
CONDUCTION DESIGN HEAT LOSS 1, BTU/HR
PERCENTAGE DIFFERENCE 2
CONDUCTION DESIGN HEAT LOSS 2, BTU/HR
PERCENTAGE DIFFERENCE 3
ACTUAL HEAT INPUT, BTU/HR
PERCENTAGE DIFFERENCE 4
PERCENTAGE RADIATION
PERCENT CEILING COVERED BY PANELS
. HEAT OUTPUT PER UNIT PANEL AREA, BtU/HR.SQ FT
FLOOR TEMPERATURE, DEG. F
ROOM AIR TEMPERATURE, DEG. F
MEAN RADIANT TEMPERATURE, DEG. F
OPERATIVE TEMPERATURE, DEG. F
EFFECTIVE RADIANT FIELD, BTU/HR. SQ FT
A.U.S.T, DEG. F
PARAMETER 1, BTU/HR.SQ FT.F
PARAMETER 3, DIMENSIONLESS
NG PANELS FOR A 140°F PANEL TEMPERATURE
1
302.2
26762.4
23671.1
-11.6
25650.1
-4.2
25650.1
-4.2
24263.4
-9.3
95.3
33.6
80.3
74.8
66.7
77.0
72.4
7.5
68.5
1.0937
0.0124
2
301.7
26762.4
23665.6
-11.6
25649.0
-4.2
25649.0
-4.2
24263.6
-9.3
95.3
33.5
80.4
74.2
66.7
77.0
72.4
7.6
68.7
1.0954
0.0124
4
301.8
26762.4
23681.8
-11.5
25840.6
-3.4
25840.6
-3.4
24450.4
-8.6
95.4
33.5
81.0
73.6
66.7
77.0
72.4
7.5
68.8
1.1068
0.0126
6
301.8
26762.4
23682.3
-11.5
25850.1
-3.4
25850.1
-3.4
24460.0
-8.6
95.4
33.5
81.1
73.4
66.7
77.0
72.4
7.5
68.8
1.1075
0.0126
TABLE 29. RESULTS FOR A PERIMETER RADIANT PANEL HEATING SYSTEM - 15' x 15' x 8' ROOM WITH THREE INSIDE WALLS AND ONE OUTSIDE WALL WITH HALF GLASS - 36" WIDE PANEL - FLOOR AND CEILING AT RADIANT BASE CASE CONDITIONS
PERIMETER HEATER TEMPERATURE, DEC F
PANEL AREA REQUIRED , SQ FT
ASHRAE DESIGN HEAT LOSS, BTU/HR
ACTUAL DESIGN HEAT LOSS, BTU/HR
PERCENTAGE DIFFERENCE 1
CONDUCTION DESIGN HEAT LOSS 1, BTU/HR
PERCENTAGE DIFFERENCE 2
CONDUCTION DESIGN HEAT LOSS 2, BTU/HR
PERCENTAGE DIFFERENCE 3
ACTUAL HEAT INPUT, BTU/HR
PERCENTAGE DIFFERENCE 4
PERCENTAGE RADIATION
PERCENT CEILING COVERED BY PANELS
HEAT OUTPUT PER UNIT PANEL AREA, BTU/HR.SQ FT
FLOOR TEMPERATURE, DEG. F
ROOM AIR TEMPERATURE, DEG. F
MEAN RADIANT TEMPERATURE, DEG. F
OPERATIVE TEMPERATURE, DEG. F
EFFECTIVE RADIANT FIELD, BTU/HR. SQ FT
A.U.S.T, DEG. F
PARAMETER 1, BTU/HR.SQ FT.F
PARAMETER 3, DIMENSIONLESS
175.0
45.0
6397.9
5770.2
-9.8
6381.1
r0.3
6381.1
-0.3
6171.7
-3.5
94.8
20.0
137.2
71.2
67.9
75.4
72.1
5.4
69.5
1.2808
0.0089
180.0
42.5
6397.9
5769.3
-9.8
6390.0
-0.1
6390.0
-0.1
6192.3
-3.2
94.8
18.9
145.7
71.3
67.9
75.4
72.1
5.5
69.5
t.2993
0.0085
185.0
40.2
6397.9
5768.5
-9.8
6397.3
0.0
6397.3
0.0
6210.1
-2.9
94.8
17.9
' 154.4
71.4
67.9
75.4
72.1
5.5
69.5
1.3176
0.0081
panel calculations.
5.6 Comparison of Forced Air and Radiant Ceiling Panels.
Several of the situations from forced air and radiant ceiling panel sys
tems were compared and these are shown in Figures 25 and 26. Figure 25 shows
radiant ceiling panels at 120 F (rad-120) and at 180 F (rad-180) compared with
forced air heating systems with an air temperature gradient of 0.75 F/ft
(con-0.75) and a gradient of 1.5 F/ft (con-1.50). From this figure it appears
that the increased infiltration heat loss, including an air temperature gra
dient in the forced air cases, is not enough to overcome the effect of an
increased AUST in the radiant case (Bar-Conduction 2 in Figure 25). Keep in
mind that these results are fo'r a nine feet high room and one-half air change
per hour.
In Figure 26, two different room heights are compared for forced air and
radiant heating systems. The con-8 case is for an 8 feet high ceiling
forced air system and rad-8 is for an 8 feet high ceiling radiant system. The
con-25 and rad-25 are for the same variables except that the ceiling is 25 feet
high. These results show the same trends as discussed above except that panel
heating system design loads become equivalent to the forced air design loads
and the ASHRAE standard design heating loads as long as room air temperature
gradients are considered. This is due to more of the radiant energy falling
on the walls as the height of the building is increased.
5.7 Heated Floor Cases.
The radiant heated floor type of system has also been simulated. The base
case which was used for this was the same as that previously described except
that it has a room height of 8 feet and the outside design temperature was
selected to be 10 F. The room height of 8 feet was chosen for this case since
the radiant floor type system is commonly applied to residential structures.
The 10 F outside temperature was selected since the floor temperature is lini
o o I
OH I \
fc O O o
en V)
3 X
C=CONVECTIVE , R=RADIANT
R(120 R(180 F) C(0.75) C(1.50)
1771 ASHRAE f V ^ I CONDUCTION 1 &?Z\ CONDUCTION 2
FIGURE 25. COMPARISON OF THE FORCED AIR AND RADIANT SYSTEMS AT SELECTED SETS OF CONDITIONS
fK^Vi «s*.-v;-.a F* v.q pv i
o I
80.0 C=CONVECTIVE , R=RADIANT
70.0 -
£* 60.0 4 X \
m o o 2 40.0 4
50.0 -
(A Q 30.0 -
I 20.0 -
10.0 -
0.0 C(8 f t ) R(8 f t ) C(25 f t ) R(25 f t )
1771 ASHRAE I V ^ I CONDUCTION 1 U77X CONDUCTION 2
FIGURE 26. COMPARISON OF THE FORCED AIR AND RADIANT SYSTEMS FOR TWO ROOM HEIGHTS
ited to 85 F for comfort requirements (See Annotated Bibliograpy-Appendix B)
and with a 3 F outside temperature not.many floor temperature variations were
available. For this case, the floor temperature was varied between 81 F and
85 F and the required floor area for heating was calculated assuming a uniform
and constant floor temperature. These results are presented in Table 30.
From the data in Table 30 it can be seen that the percent difference
between HID and HLC is constant at about -7%. The actual heat input is lower
than HLD because most of the floor is covered with radiant heating surface and
no loss from the floor to the surroundings is considered for the heated
area. In the desipn process, this heat loss would be taken into account. The
room air temperature remains constant at about 70 F and the MRT was approxi
mately 73 F.
Next, the outside design temperature was varied between 5 F and 20 F to
indicate its effect on the design heat loss and these results for an 84 F
floor temperature are given in Table 31. The trend from these calculations is
that as the climate becomes milder HLD and HLC begin to diverge. However,
this is only 1.5% for an outside temperature change from 5 F to 20 F.
In Table 32, the U-factor for the floor was changed between 0.07 and 0.15
Btu/hr ft^ F while the floor temperature was maintained at 84 F. It is seen
that there is a slight increase (-6.9 to -7.9) in the deviation from the ASH-
RAE standard design procedure. The other variables in the calculation (except
actual heat input and floor area) are affected very little by this change.
In Table 33, the infiltration rate was varied from 0.5 to 1.25 air changes
per hour for the base configuration room with an 84 F floor temperature. The
percent change in design load only increased an insignificant amount (1/2%).
At 1.25 ACH the floor is 100% active with heating surface. There are reduc
tions in room temperature and an increase in MRT and a resulting increase in
the AUST. This results also in the percent radiation from the heated floor
-102-
TABLE 30. RESULTS FOR HEATED FLOORS - BASE CASE
HEATED FLOOR TEMPERATURE, DEC F
PANEL AREA REQUIRED , SQ FT
ASHRAE DESIGN HEAT LOSS, BTU/HR
ACTUAL DESIGN HEAT LOSS, BTU/HR
PERCENTAGE DIFFERENCE 1
CONDUCTION DESIGN HEAT LOSS 1, BTU/HR
PERCENTAGE DIFFERENCE 2
CONDUCTION DESIGN HEAT LOSS 2, BTU/HR
PERCENTAGE DIFFERENCE 3
ACTUAL HEAT INPUT, BTU/HR
PERCENTAGE DIFFERENCE 4
PERCENTAGE RADIATION
PERCENT FLOOR COVERED BY PANELS
HEAT OUTPUT PER UNIT PANEL AREA, BTU/HR.SQ FT
CEILING TEMPERATURE, DEG, F
ROOM AIR TEMPERATURE, DEG. F
MEAN RADIANT TEMPERATURE, DEG. F
OPERATIVE TEMPERATURE, DEG. F
EFFECTIVE RADIANT FIELD, BTU/HR. SQ FT
A.U.S.T, DEG. F
PARAMETER 1, BTU/HR.SQ FT.F
PARAMETER 3, DIMENSIONLESS
81.0
875.7
22386.0
20536.8
-8.3
20816.5
-6.9
20846.5
-6.9
17326.8
-22.6
63.3
97.3
19.8
69.9
69.6
73.2
71.6
2.6
67.0
1.7252
0.1170
82.0
815.8
22386.0
20564.8
-8.1
20842.5
-6.9
20842.5
-6.9
17561.1
-21.6
62.7
90.6
21.5
69.9
69.7
73.1
71.6
2.5
66.8
1.7378
0.1086
83.0
762.8
22386.0
20591.8
-8.0
20835.3
-6.9
20835.3
-6.9
17767.9
-20.6
62.1
84.8
23.3
70.0
69.8
73.0
71.6
2.3
66.7
1.7504
0.1016
84.0
724.4
22386.0
20612.8
-7.9
20830.5
-6.9
20830.5
-6.9
17917.9
-20.0
61.7
80.5
24.7
70.0
69.9
72.9
71.6
2.2
66.6
1.7606
0.0967
85.0
681.0
22386.0
20638.0
-7.8
20825.4
-7.0
20825.4
-7.0
18087.3
-19.2
61.2
75.7
26.6
70.0
69.9
72.8
71.5
2.1
66.5
1.7734
0.0913
TABLE 31. EFFECTS OF OUTSIDE AIR TEMPERATURE CHANGE FOR A HEATED FLOOR AT 84°F
OUTSIDE DESIGN TEMPERATURE DEG. F
PANEL AREA REQUIRED , SQ FT
ASHRAE DESIGN HEAT LOSS, BTU/HR
ACTUAL DESIGN HEAT LOSS, BTU/HR
PERCENTAGE DIFFERENCE 1
CONDUCTION DESIGN HEAT LOSS 1, BTU/HR
PERCENTAGE DIFFERENCE 2
CONDUCTION DESIGN HEAT LOSS 2, BTU/HR
PERCENTAGE DIFFERENCE 3
ACTUAL HEAT INPUT, BTU/HR
PERCENTAGE DIFFERENCE 4
PERCENTAGE RADIATION
PERCENT FLOOR COVERED BY PANELS
HEAT OUTPUT PER UNIT PANEL AREA, BTU/HR
CEILING TEMPERATURE, DEG. F
ROOM AIR TEMPERATURE, DEG. F
MEAN RADIANT TEMPERATURE, DEG. F
OPERATIVE TEMPERATURE, DEG. F
EFFECTIVE RADIANT FIELD, BTU/HR. SQ FT
A.U.S.T, DEG. F
PARAMETER 1, BTU/HR.SQ FT.F
PARAMETER 3, DIMENSIONLESS
5.0
755.8
24108.0
22286.9
-7.6
22545.9
-6.5
22545.9
-6.5
19258.4
-20.1
61.7
84.0
25.5
69.9
69.7
73.1
71.6
2.5
66.3
1.7713
0.0947
10.0
724.4
22386.0
20612.8
-7.9
20830.5
-6.9
20830.5
-6.9
17917.9
-20.0
61.7
80.5
24.7
70.0
69.9
72.9
71.6
2.2
66.6
1.7606
0.0967
15.0
681.6
20664.0
18943.4
-8.3
19115.1
-7.5
19115.1
-7.5
16598.5
-19.7
61.7
75.7
24.4
70.1
70.0
72.7
71.5
2.0
66.9
1.7523
0.0977
20.0
636.5
18942.0
17273.4
-8.8
17400.9
-8.1
17400.9
-8.1
15260.1
-19.4
61.6
70.7
24.0
70.2
70.2
72.5
71.5
1.7
67.2
1.7441
0.0989
TABLE 32. EFFECTS OF FLOOR U-FACTORS ON A HEATED
FLOOR U-FACTOR, BTU/HR.SQ FT.F
PANEL AREA REQUIRED , SQ FT
ASHRAE DESIGN HEAT. LOSS, BTU/HR
ACTUAL DESIGN HEAT LOSS, BTU/HR
PERCENTAGE DIFFERENCE 1
CONDUCTION DESIGN HEAT LOSS 1, BTU/HR
PERCENTAGE DIFFERENCE 2
CONDUCTION DESIGN HEAT LOSS 2, BTU/HR
PERCENTAGE DIFFERENCE 3
ACTUAL HEAT INPUT, BTU/HR
PERCENTAGE DIFFERENCE 4
PERCENTAGE RADIATION
PERCENT FLOOR COVERED BY PANELS
HEAT OUTPUT PER UNIT PANEL AREA,
CEILING TEMPERATURE, D E C F
ROOM AIR TEMPERATURE, DEG. F
MEAN RADIANT TEMPERATURE, DEG. F
OPERATIVE TEMPERATURE, DEG. F
EFFECTIVE RADIANT FIELD, BTU/HR.
A.U.S.T, DEG. F
PARAMETER 1, BTU/HR.SQ FT.F
PARAMETER 3, DIMENSIQNLESS
BTU/HR.SQ FT
SQ FT
0.07
724.4
22386.0
20612.8
-7.9
20830.5
-6.9
20830.5
-6.9
17917.9
-20.0
61.7
80.5
24.7
70.0
69.9
72.9
71.6
2.2
66.6
1.7606
0.0967
AT 84°F
o.io
734.2
24141.0
22239.9
-7.9
22372.0
-7.3
22372.0
-7.3
18160.7
-24.8
61.8
81.6
24.7
70.0
69.9
72.9
71.6
2.2
66.4
1.7640
0.0961
0.15
747.5
27066.0
24952.9
-7.8
24928.2
-7.9
24928.2
-7.9
18515.5
-31.6
61.9
83.1
24.8
69.9
69.9
72.8
71.5
2.1
66.2
1.7692
0.0952
TABLE 33. EFFECTS DUE TO INFILTRATION FOR A HEATED FLOOR AT 84°F FOR A 30' x 30" x 8' ROOM
o ON I
INFILTRATION RATE, AC/H
PANEL AREA REQUIRED , SQ FT
ASHRAE DESIGN HEAT LOSS, BTU/HR
ACTUAL DESIGN HEAT LOSS, BTU/HR
PERCENTAGE DIFFERENCE 1
CONDUCTION DESIGN HEAT LOSS 1, BTU/HR
PERCENTAGE DIFFERENCE 2
CONDUCTION DESIGN HEAT LOSS 2, BTU/HR
PERCENTAGE DIFFERENCE 3
ACTUAL HEAT INPUT, BTU/HR
PERCENTAGE DIFFERENCE 4
PERCENTAGE RADIATION
PERCENT FLOOR COVERED BY PANELS
HEAT OUTPUT PER UNIT PANEL AREA, BTU/HR.SQ FT
CEILING TEMPERATURE, D E C F ROOM AIR TEMPERATURE, DEG. F
MEAN RADIANT TEMPERATURE, DEG. F
OPERATIVE TEMPERATURE, DEG. F
EFFECTIVE RADIANT FIELD, BTU/HR. SQ FT
A.U.S.T, DEG. F
PARAMETER 1, BTU/HR.SQ FT.F
PARAMETER 3, DIMENSIONLESS
0.50
724.4
22386.0
20612.8
-7.9
20830.5
-6.9
20830.5
-6.9
17917.9
-20.0
61.7
80.5
24.7
70.0
69.9
72.9
71.6
2.2
66.6
1.7606
0.0967
0.75
777.6
24492.0
22302.4
-8.9
22800.8
-6.9
22800.8
-6.9
19661.0
-19.7
59.5
86.4
25.3
70.6
69.2
73.8
71.7
3.4
67.1
1.6959
0.0947
1.00
837.5
26598.0
23955.1
-9.9
24734.4
-7.0
24734.4
-7.0
21337.7
-19.8
57.5
93.1
25.5
71.1
68.5
74.6
71.9
4.4
67.7
1.6378
0.0944
1.25
896.0
28704.0
25570.9
-10.9
26574.8
-7.4
26574.8
-7.4
22984.8
-19.9
55.5
99.6
25.7
71.6
67.9
75.4
72.1
5.5
68.3
1.5842
0.0946
pw-jsaj P:'-''
TABLE 3 4 . HEATED FLOOR CASES ROOM
3 INSIDE WALLS AND 1 OUTSIDE WALL WITH HALF GLASS FOR A 15 ' x 15 ' x 8 '
o i
FLOOR TEMPERATURE, DEC F
PANEL AREA REQUIRED , SQ FT
ASHRAE DESIGN HEAT LOSS, BTU/HR
ACTUAL DESIGN HEAT LOSS, BTU/HR
PERCENTAGE DIFFERENCE 1
CONDUCTION DESIGN HEAT LOSS 1, BTU/HR
PERCENTAGE DIFFERENCE 2
CONDUCTION DESIGN HEAT LOSS 2, BTU/HR
PERCENTAGE DIFFERENCE 3
ACTUAL HEAT INPUT, BTU/HR
PERCENTAGE DIFFERENCE 4
PERCENTAGE RADIATION
PERCENT FLOOR COVERED BY PANELS
HEAT OUTPUT PER UNIT PANEL AREA, BTU/HR.SQ FT
CEILING TEMPERATURE, DEC F ROOM AIR TEMPERATURE, DEG. F
MEAN RADIANT TEMPERATURE, DEG. F
OPERATIVE TEMPERATURE, DEG. F
EFFECTIVE RADIANT FIELD, BTU/HR. SQ FT
A.U.S.T, DEG. F
PARAMETER 1, BTU/HR.SQ FT.F
PARAMETER 3, DIMENSIONLESS
83.0
197.7
5775.9
5318.6
-7 .9
5U33.U
-5 .9 '
5433.4
-5.9
4617.2
-20.1
60.3
87.9
23.4
69.2
69.9
72.9
71.6
2.2
68.0
1.7634
0.1112
84.0
184.9
5775.9
5324.3
-7.8
5434.2 ! - 5 . 9
5434.2
-5 .9
4672.0
-19.1
59.9
82.2
25.3
69.2
69.9
72.8
71,5
2.1
68.0.
1.7819
0.1045
85.0
173.5
5775.9
5329.7
-7.7
5434.8
-5 .9
5434.8
-5 .9
4720.1
-18.3
59.5
77.1
27.2
69.2
70.0
72.8
71.5
2.0
67.9
1.7988
0.0986
TABLE 35. HEATED FLOOR CASES - 2 INSIDE WALLS AND 2 OUTSIDE WALLS, ONE WITH HALF GLASS FOR A 15' x 15' x 8' ROOM
FLOOR TEMPERATURE, DEC F
PANEL AREA REQUIRED , SQ FT
ASHRAE DESIGN HEAT LOSS, BTU/HR
ACTUAL DESIGN HEAT LOSS, BTU/HR
PERCENTAGE DIFFERENCE 1
CONDUCTION DESIGN HEAT LOSS 1, BTU/HR
PERCENTAGE DIFFERENCE 2
CONDUCTION DESIGN HEAT LOSS 2, BTU/HR
PERCENTAGE DIFFERENCE 3
ACTUAL HEAT INPUT, BTU/HR
PERCENTAGE:DIFFERENCE 4
PERCENTAGE RADIATION
PERCENT FLOOR COVERED BY PANELS
HEAT OUTPUT PER UNIT PANEL AREA, BTU/HR.SQ FT
CEILING TEMPERATURE, D E C F
ROOM AIR TEMPERATURE, DEG. F
MEAN RADIANT TEMPERATURE, DEG. F
OPERATIVE TEMPERATURE, DEG. F
EFFECTIVE RADIANT FIELD, BTU/HR. SQ FT
A.U.S.T, DEG. F
PARAMETER 1, BTU/HR.SQ FT.F
PARAMETER 3, DIMENSIONLESS
83.0
217.2
6548.1
6031.3
-7.9
6091.4
-7.0
6091.4
-7.0
5216.5
-20.3
61.5
96.5
24.0
69.0
69.9
72.9
71.6
2.2
67.2
1.8152
0.1088
84.0
203.7
6548.1
6038.3
-7.8
6090.2
'-7.0
6090.2
-7.0
5269.9
-19.5
60.9
90.5
25.9
69.0
69.9
72.8
71.5
2.1
67.1
1.8269
0.1022
85.0
191.7
6548.1
6045.0
-7.7
6089.3
-7.0
6089.3
-7.0
5317.6
-18.8
60.4
85.2
27.7
69.1
70.0
72.7
71.5
2.0
67.1
1.8385
0.0964
portion going from 61.7% to 55.5%.
In Tables 34 and 35, cases for 3 inside walls and 2 inside walls are
presented. In these two situations the room is 15' x 15' x 8' and in each
case one outside wall contains half glass. The other variables are as given
in the base case. There is little effect in either situation on the differ
ence between HLD and HLC.
5.8 - Infrared Heating Cases.
Analysis has also been carried out for infrared modular (square or rectan
gular) heating types of units. For this infrared base case and the base case
for the U-tube types of infrared units discussed in Section 5.9, the follow.-
ing items were changed from the initial base case description. The ceiling
height was set at nine feet and the outside design temperature at 3 F. The
U-factors were changed as follows:
Walls - U = 0.25 Btu/hr ft2 F
Floors - U = 0.25 Btu/hr ft2 F
Ceiling - U - 0.25 Btu/hr ft2 F
Glass - U - 1.0 Btu/hr ft2 F
This change in wall and floor construction was made since these types of
radiant units are most commonly applied to industrial buildings where the
U-factors are commonly higher than what was specified in the original base
case.
Two situations were calculated for the modular and U-tube infrared cases.
The first situation was when there were no reflectors or deflectors on the
units (which is not the normal operating condition) and the second is when
there were reflectors or deflectors on the units and these reflectors are per
fect and that the placement of the units is such that none of the direct
radiation from the infrared units falls on the walls of the structure. This
second situation would be the ideal design and placement case for infrared
-109-
modular and U-tube infrared heaters. Figures 4 and 5 illustrate the first
situation and Figure 27 illustrates the ideal situation with no direct radi
ation falling on the walls. This appeared to be the most reasonable approach
to this type of heater since each manufacturer has a series of different
reflector designs and suggestions or directions for placement of the units.
By looking at these two extremes --no reflectors or deflectors and perfect
reflectors or deflectors the range of performance of generic units can be
identified.
Table 36 summarizes results for three infrared surface temperatures when
there are no reflectors or deflectors. The areas of the heaters which are .
shown in Table 36 are the total of 4 infrared heaters located at the ceiling
(without reflectors) and were compared with several manufacturers and found to
be in good agreement with their published ratings. The percent difference
betwen HLD and HLC was constant at approximately +3%. This increase in design
heating load is apparently due to a lower AUST because of increased U-factors
and also more of the radiant energy being intercepted by the walls.
In Table 37, results are presented for four 1700 F infrared units located
at the ceiling (without reflectors or deflectors) for the base case as the
ceiling height is extended to 25 feet. This indicates that as the heaters are
raised in the room more of the radiant energy is absorbed by the larger wall
area resulting in greater conduction losses. This results in a greater design
heat loss (up to 10% at 25 feet) than what is found from the ASHRAE standard
design heat loss calculation. This shows up as an increased AUST (from 63 F at
9 feet to 66 F at 25 feet). It should be kept in mind however that high tem
perature radiant units are normally mounted at the 12 to 15 feet level in an
industrial building and use reflectors to direct the radiant energy away from
the walls and toward the floor or occupants. This lowering of the units and
use of directive reflectors would nullify this 3 to 10% difference in design
-110-
&:•>:'•.•'/• p w - ' S ? p:---..^ rw:-'
Vent System
FIGURE 27. PLACEMENT OF INFRARED MODULAR UNITS WITH DEFLECTORS AND REFLECTORS TO PREVENT DIRECT WALL RADIATION
TABLE 36. INFRARED MODULAR UNITS - BASE CASE WITH NO REFLECTORS OR DEFLECTORS AND VARIABLE SURFACE TEMPERATURE
INFRA RED HEATER TEMPERATURE, DEC F
PANEL AREA REQUIRED , SQ FT
ASHRAE DESIGN HEAT LOSS, BTU/HR
ACTUAL DESIGN HEAT LOSS, BTU/HR
PERCENTAGE DIFFERENCE 1
CONDUCTION DESIGN HEAT LOSS 1, BTU/HR
PERCENTAGE DIFFERENCE 2
CONDUCTION DESIGN HEAT LOSS 2, BTU/HR
PERCENTAGE DIFFERENCE 3
ACTUAL HEAT INPUT, BTU/HR
PERCENTAGE DIFFERENCE 4
PERCENTAGE RADIATION
PERCENT CEILING COVERED BY PANELS
HEAT OUTPUT PER UNIT PANEL AREA, BTU/HR.SQ FT
FLOOR TEMPERATURE, DEG. F
ROOM AIR TEMPERATURE, DEG. F
MEAN RADIANT TEMPERATURE, DEG. F
OPERATIVE TEMPERATURE, DEG. F
EFFECTIVE RADIANT FIELD, BTU/HR. SQ FT
A.U.S.T, DEG. F
PARAMETER 1, BTU/HR.SQ FT.F
PARAMETER 3, DIMENSIONLESS
1600.0
2.4
64378.8
51343.2
-20.2
66224.4
> 2,'9
66224.4
2.9
66188.0
2.8
99.4
0.3
27832.4
73.7
60.4
84.7
74.0
17.8
62.6
18.0769
0.0006
1700.0
2.0
64378.8
51328.3
-20.3
66228.8
2.9
66228.8
2.9
66198.6
2.8
99.5
0.2
33635.6
73.7
60.4
84.7
74.0
17.8
62.6
.20.5157
0.0006
1800.0
1.6
64378.8
51315.4
-20.3
66232.6
2.9
66232.6
2.9
66207.4
2.8
99.5
0.2
40312.8
73.7
60.4
84.8
74.0
17.9
62.6
23.1746
0.0005
TABLE 37. INFRARED MODULAR UNITS WITH NO 1700 F - EFFECT OF ROOM HEIGHT
ROOM HEIGHT, FT
PANEL AREA REQUIRED , SQ FT
ASHRAE DESIGN HEAT LOSS, BTU/HR
ACTUAL DESIGN HEAT LOSS, BTU/HR
PERCENTAGE DIFFERENCE 1
CONDUCTION DESIGN HEAT LOSS 1, BTU/HR
PERCENTAGE DIFFERENCE 2
CONDUCTION DESIGN HEAT LOSS 2, BTU/HR
PERCENTAGE DIFFERENCE 3
ACTUAL HEAT INPUT, BTU/HR
PERCENTAGE DIFFERENCE 4
PERCENTAGE RADIATION
PERCENT CEILING COVERED BY PANELS
HEAT OUTPUT PER UNIT PANEL AREA, BTU/HR.SQ FT
FLOOR TEMPERATURE, DEG. F
ROOM AIR TEMPERATURE, DEG. F
MEAN RADIANT TEMPERATURE, DEG. F
OPERATIVE TEMPERATURE, DEG. F
EFFECTIVE RADIANT FIELD, BTU/HR. SQ FT
A.U.S.T, DEG. F
PARAMETER 1, BTU/HR.SQ FT.F
PARAMETER 3, DIMENSIONLESS
OR DEFLECTORS AND SURFACE TEMPERATURE AT
9.0
2.0
64378.8
51328.3
-20.3
66228.8
2.9
66228.8
2.9
66198.6
2.8
99.5
0.2
33635.6
73.7
60. 4
84.7
74.0
17.8
62.6
20.5157
0.0006
10.0
2.1
67932.0
53973.1
-20.5
69934.7
2.9
69934.7
2.9
69903.0
2.9
99.5
0.2
33646.9
73.5
60.2
85.0
74.0
18.2
62.6
20.5179
0.0006
12.0
2.3
75038.4
59514.3
-20.7
77682.6
3.5,
77682.6-
3.5
77647.5
3.5
99.5
0.3
33646.8
73.4
60.1
85.1
74.1
18.3
62.8
20.5167
0.0006
15.0
2.6
85698.0
67341.4
-21.4
88684.3
3.5
88684.3
3.5
88644.7
3.4
99.5
0.3
33648.2
72.4
59.6
85.7
74.2
19.2
62.7
20.5109
0.0006
20.0
3.3
103464.0
83916.3
-18.9
111769.5
8.0
111769.5
8.0
111718.1
8.0
99.5
0.4
33646.8
73.7
61.4
83.5
73.8
16.2
65.1
20.5327
0.0006
25.0
4.0
121230.0
99536.9
-17.9
133661.2
10.3
133661.2
10.3
133599.0
10.2
99.5
0.4
33646.2
73.3
62.1
82.7
73.6
15.1
66.3
20.5414
0.0006
TABLE 38. INFRARED MODULAR UNITS WITH NO REFLECTORS OR DEFLECTORS AND SURFACE TEMPERATURE AT 1700 F AND HEIGHT = 15' - EFFECT OF INFILTRATION RATE
AIR CHANGES PER HOUR, AC/H
PANEL AREA REQUIRED , SQ FT
ASHRAE DESIGN HEAT LOSS, BTU/HR
ACTUAL DESIGN HEAT LOSS, BTU/HR
PERCENTAGE DIFFERENCE 1
CONDUCTION DESIGN HEAT LOSS 1, BTU/HR
PERCENTAGE DIFFERENCE 2
CONDUCTION DESIGN HEAT LOSS 2, BTU/HR
PERCENTAGE DIFFERENCE 3
ACTUAL HEAT INPUT, BTU/HR
PERCENTAGE DIFFERENCE 4
PERCENTAGE RADIATION
PERCENT CEILING COVERED BY PANELS
HEAT OUTPUT PER UNIT PANEL AREA, BTU/HR.SQ FT
FLOOR TEMPERATURE, DEG. F
ROOM AIR TEMPERATURE, DEG. F
MEAN RADIANT TEMPERATURE, DEG. F
OPERATIVE TEMPERATURE, DEG. F
EFFECTIVE RADIANT FIELD, BTU/HR. SQ FT
A.U.S.T, DEG. F
PARAMETER 1, BTU/HR.SQ FT.F
PARAMETER 3, DIMENSIONLESS
• 1.0
2.9
94446.0
71677.8
-24.1
95968.8
1.6
95968.8
1.6
95926.2
1.6
99.5
0.3
33647.4
73.7
57.6
88.0
74.6
22.3
63.3
20.4863
0.0006
2.0
3.2
111942.0
79646.7
-28.9
1092p2.3
-2.4
109202.3
-2.4
109154.3
-2.5
99.5
0.4
33646.5
76.0
54.2
92.1
75.4
27.8
64.3
20.4432
0.0005
3.0
3.6
129438.0
86809.3
-32.9
121735.1
-6.0
121735.1
-6.0
121680.9
-6.0
99.5
. 0.4
33644.7
78.5
51.3
96.2
76.4
33.2
. 65.8
20.4057
0.0005
4.0
4.0
146934.0
93065.8
-36.7
133075.9
-9.4
133075.9
-9.4
133015.7
-9.5
99.5
0.4
33644.1
80.9
48.6
100.0
77.4
38.0
67.3
20.3721
0.0005
heat loss as illustrated In the following calculations.
In Table 38, the room was 15' highland the modular infrared units were at
1700 F and they did not have reflectors or deflectors. In this case (Table
38), the infiltration rate was changed between 1 and 4 ACH. As can be seen in
Tables 38 and 37, the percent difference in HLD and HLC goes from +3.5% at 0.5
ACH to -9.4% at 4 ACH. This illustrates how much the infiltration rate
affects the design load calculation. This is due to the exchange of lower
temperature air for radiant systems when compared to forced air systems. With
this increase in the infiltration rate, the floor temperature has an increase
to 80.9 F, the room air temperature for comfort has decreased to 49 F and the
mean radiant temperature has increased to 100 F. In this situation, the 4 ACH
would most likely be beyond any normal situation (except for something such as
spot heating) and does not represent a realistic situation. However, the
importance of the change in the design heat loss load compared to the standard
ASHRAE design load as infiltration is changed is strongly supported.
In Table 39, the convection coefficient at the modular infrared units was
changed by up to a factor of 5 for a 15' high room with 3 ACH and a modular
infrared heater surface at 1700 F without reflectors or deflectors. As seen
in this table, the assumption concerning the convection coefficient off of the
heater surface has negligible effect on the calculations made here. If reflec
tors are used, there might be more of an effect due to more area available for
convection heat transfer, however, it is expected to be negligible also.
In Tables 40, 41, and 42 cases were run where the modular infrared units
had perfect reflectors or deflectors and were positioned such that none of
their direct radiation fell on the walls. This situation is illustrated in
Figure 27 for an individual application.
The situation in Table 40 is identical to the situation reported in Table
36 -- infrared base case -- except that Table 40 uses perfect reflectors and
-115-
TABLE 39. INFRARED MODULAR UNITS WITH NO REFLECTORS OR DEFLECTORS AND SURFACE TEMPERATURE AT 1700 F, 15' HIGH, 3 ACH - EFFECT OF CONVECTION COEFFICIENT MULTIPLIER
CONVECTION MULTIPLIER
PANEL AREA REQUIRED , SQ FT
ASHRAE DESIGN HEAT LOSS, BTU/HR
ACTUAL DESIGN HEAT LOSS, BTU/HR
PERCENTAGE DIFFERENCE 1
CONDUCTION DESIGN HEAT LOSS 1, BTU/HR
PERCENTAGE DIFFERENCE 2
CONDUCTION DESIGN HEAT LOSS 2, BTU/HR
PERCENTAGE DIFFERENCE 3
ACTUAL HEAT INPUT, BTU/HR
PERCENTAGE DIFFERENCE 4
PERCENTAGE RADIATION
PERCENT CEILING COVERED BY PANELS
HEAT OUTPUT PER UNIT PANEL AREA,
FLOOR TEMPERATURE, DEG. F
ROOM AIR TEMPERATURE, DEG. F
MEAN RADIANT TEMPERATURE, DEG. F
OPERATIVE TEMPERATURE, DEG. F
EFFECTIVE RADIANT FIELD, BTU/HR.
A.U.S.T, DEG. F
PARAMETER 1, BTU/HR.SQ FT.F
PARAMETER 3,.DIMENSIONLESS
BTU/HR.SQ FT
SQ FT
1.0
3.6
129438.0
86809.3
-32.9
121735.1
-6.0 '
121735.1
-6.0
121680.9
-6.0
99.5
0.4
33644.7
78.5
51.3
96.2
76.4
33.2
65.8
20.4057
0.0005
2.0
3.6
129438.0
87104.7
-32.7
121714.3
-6.0
121714.3
-6.0
121660.5
-6.0
98.9
0.4
33830.2
78.4
51.5
96.0
76.4
32.9
65.7
20.5200
0.0006
5.0
3.5
129438.0
87964.9
-32.0
121664.4
-6.0
121664.4
-6.0
121611.7
-6.0
97.3
0.4
34381.6
78.0
51.9
95.3
76.2
32.0
65.5
20.8606
0.0006
KSfiiJKSS <P-i-*;<s f ^ v i . - j a , r
INFRARED MODULAR UNITS WITH PERFECT REFLECTORS IN A 9 ' HIGH BASE CASE ROOM - EFFECT OF SURFACE TEMPERATURE
INFRA RED HEATER TEMPERATURE, DEC F
PANEL AREA REQUIRED , SQ FT
ASHRAE DESIGN HEAT LOSS, BTU/HR
ACTUAL DESIGN HEAT LOSS, BTU/HR
PERCENTAGE DIFFERENCE 1
CONDUCTION DESIGN HEAT LOSS 1 , BTU/HR
PERCENTAGE DIFFERENCE 2
CONDUCTION DESIGN HEAT LOSS 2 , BTU/HR
PERCENTAGE DIFFERENCE 3
ACTUAL HEAT INPUT, BTU/HR
PERCENTAGE DIFFERENCE 4
PERCENTAGE RADIATION
PERCENT CEILING COVERED BY PANELS
HEAT OUTPUT PER UNIT PANEL AREA, BTU/HR.SQ FT
FLOOR TEMPERATURE, DEG. F
ROOM AIR TEMPERATURE, DEG. F
MEAN RADIANT TEMPERATURE, DEG. F
OPERATIVE TEMPERATURE, DEG. F
EFFECTIVE RADIANT FIELD, BTU/HR. SQ FT
A . U . S . T , DEG. F
PARAMETER 1, BTU/HR.SQ FT.F
PARAMETER 3, DIMENSIONLESS
1600.0
2.3
64378.8
51189.5
-20 .5
63752.4
- 1 . 0
63752.. 4
- 1 . 0
63716.9
- 1 . 0
99.4
0.3
27819.0
83.3
60.2
84.9
74.0
18.1
61.4
18.0661
0.0006
1700.0
1.9
64378.8
51177.4
-20 .5
63755.6
- 1 . 0
63755.6
- 1 . 0
63726.2
- 1 . 0
99.5
0.2
33621.7
83.3
60.2
84.9
74.0
18.1
61.4
20.5051
0.0006
1800. 0
1 . 6
64378. 8
51166. 8
-20 . 5
63758. 2
- 1 . 0
63758. 2
- 1 . 0
63733. 7
- 1 . 0
99. 5
0. 2
40298. 0
83. 3
60. 2
85. 0
74. 0
18. 1
6 1 . 4
23.164 1
0.00015
TABLE 41. INFRARED MODULAR UNITS WITH PERFECT OF ROOM HEIGHT
ROOM HEIGHT, FT 8.0
PANEL AREA REQUIRED , SQ FT" 1.7
ASHRAE DESIGN HEAT LOSS, BTU/HR 60825.6
ACTUAL DESIGN HEAT LOSS, BTU/HR 47042.2
PERCENTAGE DIFFERENCE 1 -22.7
CONDUCTION DESIGN HEAT LOSS 1, BTU/HR 58558.7
PERCENTAGE DIFFERENCE 2 -3.7
CONDUCTION DESIGN HEAT LOSS 2, BTU/HR 58558.7
PERCENTAGE DIFFERENCE 3 -3.7
ACTUAL HEAT INPUT, BTU/HR . 58532.4
PERCENTAGE DIFFERENCE 4 -3.8
PERCENTAGE RADIATION 99.5
PERCENT CEILING COVERED BY PANELS 0.2
HEAT OUTPUT PER UNIT PANEL AREA, BTU/HR.SQ FT 33626.8
FLOOR TEMPERATURE, DEG. F 79.4
ROOM AIR TEMPERATURE, DEG. F 58.7
MEAN RADIANT TEMPERATURE, DEG. F 86.8
OPERATIVE TEMPERATURE, DEG. F 74.4
EFFECTIVE RADIANT FIELD, BTU/HR. SQ FT 20.6
A.U.S.T, DEG. F 59.7
PARAMETER 1, BTU/HR.SQ FT.F 20.4886
PARAMETER 3, DIMENSIONLESS 0.0006
WITH 0.5 ACH AND 1700 F HEATERS - EFFECT
9.0
1.9
64378.8
51177.4
-20.5
63755.6
-1.0
63755.6
-1.0
63726.2
-1.0
99.5
0.2
33621.7
83.3
60.2
84.9
74.0
18.1
61.4
20.5051
0.0006
10.0
2.0
67932.0
53685.4
-21.0
66870.4
-1.6
66870.4
-1.6
66839.8
-1.6
99.5
0.2
33621.2
84.7
59.9
85.3
74.1
18.7
61.1
20.5005
0.0006
12.0
2.2
75038.4
58866.2
-21.6
73314.1
-2.3
73314.1
-2.3
73280.7
-2.3
99.5
0.2
33628.2
87.8
59.5
85.8
74.2
19.3
60.8
20.4975
0.0006
15.0
2.4
85698.0
65878.9
-23.1
82005.6
-4.3
82005.6
-4.3
81969.2
-4.4
99.5
0.3
33624.5
91.5
58.3
87.2
74.4
21.2
59.8
20.4811
0.0006
20.0
3.0
103464.0
80311.5
-22.4
100056.2
-3.3
100056.2
-3.3
100011.7
-3.3
99.5
0.3
33614.7
101.2
58.9
86.5
74.3
20.3
60.5
20.4818
0.0006
25.0
3.4
121230.0
92743.0
-23.5
115527.7
-4.7
115527.7
-4.7
115477.5
-4.7
99.5
0.4
33606.8
108.5
58.1
87.5
74.5
21.6
59.8
20.4668
0.0006
so I
psa r~"? r-^
TABLE 42, INFRARED MODULAR UNITS WITH PERFECT REFLE( EFFECT OF INFILTRATION
INFILTRATION RATE, AC/H
PANEL AREA REQUIRED , SO. FT
ASHRAE DESIGN HEAT LOSS, BTU/HR
ACTUAL DESIGN HEAT LOSS, BTU/HR
PERCENTAGE DIFFERENCE 1
CONDUCTION DESIGN HEAT LOSS 1, BTU/HR
PERCENTAGE DIFFERENCE 2
CONDUCTION DESIGN HEAT LOSS 2, BTU/HR
PERCENTAGE DIFFERENCE 3
ACTUAL HEAT INPUT, BTU/HR
PERCENTAGE DIFFERENCE 4
PERCENTAGE RADIATION
PERCENT CEILING COVERED BY PANELS
HEAT OUTPUT PER UNIT PANEL AREA, BTU/HR.SQ FT
FLOOR TEMPERATURE, DEG. F
ROOM AIR TEMPERATURE, DEG. F
MEAN RADIANT TEMPERATURE, DEG. F
OPERATIVE TEMPERATURE, DEG. F
EFFECTIVE RADIANT FIELD, BTU/HR. SQ FT
A.U.S.T, DEG. F
PARAMETER 1, BTU/HR.SQ FT.F
PARAMETER 3, DIMENSIONLESS
IN 15' HIGH ROOM WITH HEATERS AT 1700 F -
1.0
2.6
94446.0
69909.4
-26.0
88556.3
-6.2
88556.3
-6.2
88517.1
• -6.3
99.5
0.3
33622.2
94.1
56.3
89.6
74.9
24.5
60.2
20.4541
0.0005
2.0
3.0
111942.0
77235.9
-31.0
100781.1
-10.0
100781.1
-10.0
100736.7
-10.0
99.5
0.3
33617.8
99.2
52.7
94.3
75.9
30.6
61.2
20.4064
0.0005
3.0 4.0
3.3 3.6
129438.0 146934.0
82915.1 88056.9
-35.9 -40.1
112335.0 122314.8
-13.2 -16.8
112335.0 122314.8
-13.2 -16.8
112283.9 122257.9
-13.3 -16.8
99.5 99.5
0.4 0.4
33612.4 33607.2
104.4 108.8
49.1 46.1
99.3 103.4
77.2 78.3
37.1 42.5
63.1 64.6
20.3592 20.3196
0.0005 0,0005
unit placement. It is now seen that the percent difference between HLD and
HLC has gone from +3% to -1% indicating that proper placement of the infrared
heaters can account for 4% in design load at these conditions. Also note in
Tables 36 and 40 that the floor temperature has risen 10 F, the room air tem
perature for comfort and MRT have not changed, and the AUST has dropped about
1 F.
The situation in Table 41 is identical to the situation reported in Table
37 except that Table 41 uses perfect reflectors and placement of the heating
units. In these two tables, the effect of room height is considered. It is
seen that the percent difference between HLD and HLC has changed from +10%
with no reflectors to -5% with ideal reflectors and placement indicating that
proper placement of the infrared heaters can account for 15% reduction in the
design heat loss value. Also note in Tables 37 and 41 that the floor temper
ature can get to too high of a value in the ideal situation (108 F)' but in the
actual situation this will not be realized since equipment, furniture and
people will absorb this radiant energy and intercept it before it reaches the
floor. Also observe that in Table 41 the air temperature for comfort is
lower by up to 40 F, the MRT is increased by up to 5 F, and the AUST is
reduced by up to 7 F.
The configuration and conditions in Table 42 are identical to those con
sidered in Table 38 except that Table 42 considers ideal reflectors and unit
heater placement. In these two tables the effect due to air infiltration is
considered. It is seen that the percent difference between HLD and HLC has
changed from up to -9.4% with no reflectors to up to -16.8% with ideal
reflectors. This indicates that with proper reflector design and unit place
ment of infrared heaters that up to 7% of the design heating load can be saved
when considering height of the room. Again, the floor temperature has
increased theoretically by up to 25 F by use of reflectors and proper unit
-120-
placement. The room air temperature for comfort has been reduced by 2 F, the
MRT increased by 3 F and the AUST reduced by 2 F.
5.9 - U-Tube Infrared Cases. ,
The U-tube or straight tube type of configuration for vented gas-fired
infrared units were also analyzed. These units have different orifice sizes
for the same size and length of tube so that the units will operate at differ
ent average surface temperatures. The same base case parameters were used as
described in Section 5.8 for the infrared modular units except for this case
two U-tubes were used in the space instead of the four modular units. Again,
two situations were considered; no reflectors and then with ideal reflectors
and placement. The results from these calculations are shown in Tables 43
thru 47. Tables 43 and 44 give results for the U-tube heaters that do not have
reflectors or deflectors. In Table 43, the average surface temperature was
varied between 700 and 900 F. The area calculated for the average surface
temperature of 700 and 750 F agreed reasonably well with those presented by a
manufacturer of these types units. The same conclusions can be drawn from
these results as from the infrared modular results discussed in Section 5.8.
In fact, there is very little change in the results and the trends are simi
lar.
The results given in Table 44 are for a 750 F average surface temperature,
however, the ceiling was extended in steps up to 20 feet. It is interesting
to note here that the line source of radiation causes the opposite trend in
the difference between HLD and HLC than was observed for the modular infrared
units. The behavior of the U-tube units appears to be similar to that of the
ceiling panel heating types of units.
Tables 45-47 present the results for the U-tube infrared units which have
ideal reflectors and unit placement. Table 45 is the same configuration and
conditions as in Table 43 except that ideal reflectors and unit placement is
-121-
U-TUBE INFRARED UNITS - BASE CASE WITH NO TEMPERATURE
U-TUBE IR-HEATER TEMPERATURE, DEC F
PANEL AREA REQUIRED , SO. FT
ASHRAE DESIGN HEAT LOSS, BTU/HR
ACTUAL DESIGN HEAT LOSS, BTU/HR
PERCENTAGE DIFFERENCE 1
CONDUCTION DESIGN HEAT LOSS 1, BTU/HR
PERCENTAGE DIFFERENCE 2
CONDUCTION DESIGN HEAT LOSS 2, BTU/HR
PERCENTAGE DIFFERENCE 3
ACTUAL HEAT INPUT, BTU/HR
PERCENTAGE DIFFERENCE 4
PERCENTAGE RADIATION
PERCENT CEILING COVERED BY PANELS
HEAT OUTPUT PER UNIT PANEL AREA, BTU/HR.SQ FT
FLOOR TEMPERATURE, DEG. F
ROOM AIR TEMPERATURE, DEG. F
MEAN RADIANT TEMPERATURE, DEG. F
OPERATIVE TEMPERATURE, DEG. F
EFFECTIVE RADIANT FIELD, BTU/HR. SQ FT
A.U.S.T, DEG. F
PARAMETER 1, BTU/HR.SQ FT.F
PARAMETER 3, DIMENSIONLESS
OR DEFLECTORS - CHANGE IN TUBE SURFACE
900.0
12.7
64378.8
51692.1
-19.7
66658.9
3.5
66658.9
3.5
66464.7
3.2
98.5
1.4
5234.8
73.3
60.8
84.2
73.9
17.2
62.1
6.2386
0.0011
850.0
14.8
64378.8
51721.6
-19.7
66649.9
3.5
66649.9
3.5
66423.9
3.2
98.4
1.6
4494.6
73.3
60.8
84.2
73.9
17.1
62.1
5.6962
0.0012
800.0
17.3
64378.8
51754.5
-19.6
66640.3
3.5
66640.3
3.5
66375.5
3.1
98.2
1.9
3834.3
73.2
60.9
84.2
73.9
17.1
62.1 : 5.1883
0.0012
750.0
20.4
64378.8
51791.1
-19.6
66630.0
3.5
66630.0
3.5
66317.6
3.0
98.1
2.3
3247.3
73.2
60.9
84.1
73.9
17.0
62.1
4.7132
0.0013
700.0
24.3
64378.8
51831.9
-19.5
66619.3
3.5
66619.3
3.5
66247.8
2.9
97.9
2.7
2728.5
73.1
61.0
84.1
73.9 ,
16.9
62.1
4.2702
0.0014
W~m f S ; p'"i W~m Wr-'-'i »r*3 f-'-* f :
TABLE 44. U-TUBE INFRARED UNITS - 750 F SURFACE OF CHANGE IN ROOM HEIGHT
ROOM HEIGHT, FT
PANEL AREA REQUIRED , SQ FT .
ASHRAE DESIGN HEAT LOSS, BTU/HR
ACTUAL DESIGN HEAT LOSS, BTU/HR
PERCENTAGE DIFFERENCE 1
CONDUCTION DESIGN HEAT LOSS 1, BTU/HR
PERCENTAGE DIFFERENCE 2
CONDUCTION DESIGN HEAT LOSS 2, BTU/HR
PERCENTAGE DIFFERENCE 3
ACTUAL HEAT INPUT, BTU/HR
PERCENTAGE DIFFERENCE 4
PERCENTAGE RADIATION
PERCENT CEILING COVERED BY PANELS
HEAT OUTPUT PER UNIT PANEL AREA, BTU/HR.SQ FT
FLOOR TEMPERATURE, DEG. F
ROOM AIR TEMPERATURE, DEG. F
MEAN RADIANT TEMPERATURE, DEG. F
OPERATIVE TEMPERATURE, DEG. F
EFFECTIVE RADIANT FIELD, BTU/HR. SQ FT
A.U.S.T, DEG. F
PARAMETER 1, BTU/HR.SQ FT.F
PARAMETER 3, DIMENSIONLESS
AND NO REFLECTORS OR DEFLECTORS - EFFECT
9.0
20.4
64378.8
51791.1
-19.6
66630.0
3.5
66630.0
3.5
66317.6
3.0
98.1
2.3
3247.3
73.2
60.9
84.1
73.9
17.0
62.1
4.7132
0.0013
10.0
21.4
67932.0
54135.7
-20.3
69924.7
•2.9!
69924.7"
2.9
69600.3
2.5
98.1
2.4
3248.2
72.6
60.4
84.8
74.0
17.9
61.7
4.7107
0.0013
12.0
23.4
75038.4
58689.3
-21.8
76463.0
1.9
76463.0
1.9
76115.4
1.4
98.1
2.6
3249.7
71.5
59.3
86.0
74.2
19.6
61.1
4.7056
0.0013
15.0
26.3
85698.0
65300.3
-23.8
86012.3
0.4
86012.3
0.4
85632.0
-0.1
98.1
2.9
3251.5
69.7
57.9
87.7
74.6
21.9
60.1
4.6984
0.0013
20.0
30.9
103464.0
75864.0
-26.7
101112.3
-2.3
101112.3
-2.3
100683.5
-2.7
98.1
3.4
3254.3
66.5
55.8
90.2
75.0
25.2
58.6
4.6884
0.0013
TABLE 45. U-TUBE INFRARED UNITS - BASE CASE WITH IDEAL SURFACE TEMPERATURE
U- TUBE HEATER TEMPERATURE, DEG. F
PANEL AREA REQUIRED , SQ FT
ASHRAE DESIGN HEAT LOSS, BTU/HR
ACTUAL DESIGN HEAT LOSS, BTU/HR
PERCENTAGE DIFFERENCE 1
CONDUCTION DESIGN HEAT LOSS 1, BTU/HR
PERCENTAGE DIFFERENCE 2
CONDUCTION DESIGN HEAT LOSS 2, BTU/HR
PERCENTAGE DIFFERENCE 3
ACTUAL HEAT INPUT, BTU/HR
PERCENTAGE DIFFERENCE 4
PERCENTAGE RADIATION
PERCENT CEILING COVERED BY PANELS
HEAT OUTPUT PER UNIT PANEL AREA, BTU/HR.SQ FT
FLOOR TEMPERATURE, DEG. F
ROOM AIR TEMPERATURE, DEG. F
MEAN RADIANT TEMPERATURE, DEG. F
OPERATIVE TEMPERATURE, DEG. F
EFFECTIVE RADIANT FIELD, BTU/HR. SQ FT
A.U.S.T, DEG. F
PARAMETER 1, BTU/HR.SQ FT.F
PARAMETER 3, DIMENS'I'ONLESS
REFLECTORS AND PLACEMENT - CHANGE IN TUBE
700.0
23.3
64378.8
51414.1
-20.1
63618.7
-1.2
63618.7
-1.2
63256.8
-1.7
97.9
2.6
2714.1
83.3
60.5
84.6
74.0
17.7
61.4
4.2447
0.0014
750.0
19.6
64378.8
51396.1
-20.2
63639.0
-1..1
63639.0
-1.1
63334.9
-1.6
98.1
2.2
3233.3
83.3
60.5
84.6
74.0
17.7
61.4
4.6898
0.0013
800.0
16.6
64378.8
51377.9
-20.2
63655.9
-1.1
63655.9
-1.1
63398.2
-1.5
98.2
1.8
3820.1
83.3
60.5
84.7
74.0
17.7
61.4
5.1662
0.0012
850.0
14.2
64378.8
51360.1
-20.2
63670.0
-1.1
63670.0
-1.1
63450.2
-1.4
98.4
1.6
4480.8
83.3
60.4
84.7
74.0
17.8
61.4
5.6757
0.0012
900.0
12.2
64378.8
51343.0
-20.2
63682.0
-1.1
63682.0
-1.1
63493.2
-1.4
98.5
1.4
5220.7
83.3
60.4
84.7
74.0
17.8
61.4
6.2189
0.0011
U-TUBE INFRARED UNITS - 750 F SURFACE EFFECT OF CHANGE IN ROOM HEIGHT
ROOM HEIGHT, FT
PANEL AREA REQUIRED , SQ FT
ASHRAE DESIGN HEAT LOSS, BTU/HR
ACTUAL DESIGN HEAT LOSS, BTU/HR
PERCENTAGE DIFFERENCE 1
CONDUCTION DESIGN HEAT LOSS 1, BTU/HR
PERCENTAGE DIFFERENCE 2
CONDUCTION DESIGN HEAT LOSS 2, BTU/HR
PERCENTAGE DIFFERENCE 3
ACTUAL HEAT INPUT, BTU/HR
PERCENTAGE DIFFERENCE 4
PERCENTAGE RADIATION
PERCENT CEILING COVERED BY PANELS
HEAT OUTPUT PER UNIT PANEL AREA, BTU/HR.SQ FT
FLOOR TEMPERATURE, DEG. F
ROOM AIR TEMPERATURE, DEG. F
MEAN RADIANT TEMPERATURE, DEG. F
OPERATIVE TEMPERATURE, DEG. F
EFFECTIVE RADIANT FIELD, BTU/HR. SQ FT
A.U.S.T, DEG. F
PARAMETER 1, BTU/HR.SQ FT.F
PARAMETER 3, DIMENSIONLESS
AND WITH IDEAL REFLECTORS AND PLACEMENT -
9.0
19.6
64378.8
51396.1
-20.2
63639.0
-1.1
63639.0
-1.1
63334.9
-1.6
98.1
2.2
3233.3
83.3
60.5
84.6
74.0
17.7
61.4
4'. 6898
0.0013
10.0
20.6
67932.0
53903.5
-20.7
66742.1
-1.8
66742.1 '
-1.8
66424.8
-2.2
98.1
2.3
3231.7
84.7
60.1
85.1
74.0
18.3
61.1
4.6852
0.0013
12.0
22.6
75038.4
59074.6
-21.3
73156.1
-2.5
73156.1
-2.5
72810.8
-3.0
98.1
2.5
3228.7
87.9
59.7
85.6
74.1
19.0
60.8
4.6777
0.0013
15.0
25.2
85698.0
66060.5
-22.9
81802.6
-4.5
81802.6
-4.5
81425.3
-5.0
98.1
2.8
3225.0
91.7
58.5
87.0
74.4
20.9
59.7
4.6643
0.0013
20.0
30.9
103464.0
80365.5
-22.3
99714.2
-3.6
99714.2
-3.6
99252.0
-4.1
98.1
3.4
3214.3
101.7
58.9
86.5
74.3
20.2
60.4
4.6517
0.0013
TABLE 47. U-TUBE INFRARED UNITS AT 750°F WITH IDEAL EFFECT OF CHANGES IN AIR INFILTRATION
INFILTRATION RATE, AC/H
PANEL AREA REQUIRED , SQ FT
ASHRAE DESIGN HEAT LOSS, BTU/HR
ACTUAL DESIGN HEAT LOSS, BTU/HR
PERCENTAGE DIFFERENCE 1
CONDUCTION DESIGN HEAT LOSS 1, BTU/HR
PERCENTAGE DIFFERENCE 2
CONDUCTION DESIGN HEAT LOSS 2, BTU/HR
PERCENTAGE DIFFERENCE 3
ACTUAL HEAT INPUT, BTU/HR
PERCENTAGE DIFFERENCE 4
PERCENTAGE RADIATION
PERCENT CEILING COVERED BY PANELS
HEAT OUTPUT PER UNIT PANEL AREA, BTU/HR.SQ FT
FLOOR TEMPERATURE, DEG. F
ROOM AIR TEMPERATURE, DEG. F
MEAN RADIANT TEMPERATURE, DEG. F
OPERATIVE TEMPERATURE, DEG. F
EFFECTIVE RADIANT FIELD, BTU/HR. SQ FT
A.U.S.T, DEG. F
PARAMETER 1, BTU/HR.SQ FT.F
PARAMETER 3, DIMENSItONLESS
AND PLACEMENT IN A 15' HIGH ROOM -
1.0
27.3
94446.0
70097.6
-25.8
88345.6
-6.5
88345.6
-6.5
87940.0
-6.9
98.1
3.0
3222.2
94.4
56.4
89.4
74.9
24.2
60.1
4.6466
0.0013
2.0
31.1
111942.0
77439.9
-30.8
100498.5
-10.2
100498.5
-10.2
100038.2
-10.6
98.0
3.5
3217.3
99.6
52.8
94.1
75.9
30.4
61.1
4.6152
0.0013
3.0
34.7
129438.0
83180.1
-35.7
111983.4
-13.5
111983.4
-13.5
111454.4
-13.9
98.0
3.9
3211.9
104.8
49.3
99.1
77.2
36.8
62.9
4.5842
0.0013
4.0
37.8
146934.0
88205.4
-40.0
121984.8
-17.0
121984.8
-17.0
121393.1
-17.4
98.0
4.2
3207.3
109.4
46.2
103.3
78.3
42.4
64.5
4.5577
0.0013
used in Table 45. This change results in a decrease in the percent difference
between HLD and HLC of up to 5% for this set of conditions. Therefore, the
use of ideal reflectors and unit placement can result in the savings of up to
5% in the design heat loss. Also showing up in this calculation is an
increase of about 10 F in the floor temperature when all of the infrared
radiant energy from the heater is reflected directly to the floors with none
impinging on the walls.
In Table 46, the same configuration and conditions as in Table 44 are
considered except that in Table 46 ideal reflectors and unit placement are
considered. This change results in a decrease in the percent difference bet
ween HLD and HLC of up to 4 1/2%' for 9' high ceilings. Again, ideal reflec
tors and unit placement can reduce the installed heating capacity up to 5%.
The trend in the floor temperature has been changed in Table 46 because more
area of heating surface has been installed and none of this heat is inter
cepted by the walls. This causes the floor temperature to approach 100 F for
the 20' high room.
In Table 47, the results for U-tube infrared units "at-750 F with ideal
reflectors and placement in a 15' high room are given for infiltration rates
changing from 1 to 4 ACH. This shows that the percent difference between HLD
and HLC can be up to 17% by use of ideal reflectors. However, at the same
time the floor temperature becomes very high (up to 109 F in the theoretical
undisturbed case) and the air temperature for comfort has been reduced to 46
F. These are rather extreme situations and most likely would not be encoun
tered in a total heat situation for infrared heaters.
5.10 - Summary of Design Heating Calculations
The results in the changes in percent difference between HLD and HLCG
given in Tables 6 thru 47 have been summarized in Table 48. HLD is the ASHRAE
standard heat loss calculation procedure and HLC is the design heat loss for a
-127-
TABLE 48 SUMMARY OF CALCULATED RESULTS
Fixed Conditions Variable being changed and its range
Range of the difference in percent between HLCG and HLD Percentage Difference 3
(%)
Forced Air Heating
Base case with Cl 0.5oF temp, gradient
per foot
Base case with C2 0.75oF temp, gradient
per foot
Base case with C3 loF temp, gradient
per foot
Base case with C4 1.5oF temp, gradient
per foot
15' High base C5 case with 0.75oF/ft
temp. gradient
25' High base C6 case with 0.75oF/ft
temp. gradient
25' High base case with 0.75oF/ft.
C7 temp. gradient and U-factors used in infrared cases
8 to 25 ft.
8 to 25 ft.
8 to 25 ft.
8 to 25 ft.
1.0 to 4.0 ACH
1.0 to 4.0 ACH
•7.7 to -5.0
-7.5 to -3.2
-7.3 to -1.5
-6.2 to +1.2
•1.9 to +7.2
+2.9 to +16.3
1.0 to 4.0 ACH -7.5 to +11.0
PANEL HEATING
PI Base case Panels from 120 to 180oF -4.1 to -4.2
Base case with P2 panels at 140oF -panel from
0.88 to 0.94 -4.1 to 3.4
-128-
TABLE 48 (CONTINUED)
Base case with P3 panels at 140oF
Base case with P4 panel hc doubled
Base case with P5 panel hc multiplied
by five
Base case with P6 panel hc multiplied
by ten
Base case wi th P7 pane l s a t 130<>F
Base case with P8 panels at 150oF
Base case with P9 panels at 170oF
P10 Base case with panels at 140oF
ewalls f r o m 0.8 to 0.95
Panels from 120 to 180oF
Panels from 120 to I8O0F
Panels from 120 to I8O0F
0.5 to 4.0 ACH
0.5 to 4.0 ACH
0.5 to 4.0 ACH
Glass quantity from 0 to 50% of total outside wall area
•1.7 to -5.4
-4.2 to -4.3
-4.5 to -4.5
-4.7 to -4.6
-3.5 to -15.7
-3.5 to -15.7
-3.5 to -15.6
-4.5 to -2.4
Pll Base case Panels from 130 to 170oF U increase from 30 to 100%
-2.4 to -4.2
P12 Base case with panels at 140oF
P13 Base case with panels at 140oF
P14 Base case with panels at 130oF
P15 Base case with panels at 140oF
Perimeter panel, P16 15' x 15' x 8'
room, 36" panel, 3 inside walls, 1 outside wall with glass
Room size change from 15' x 30' to 100' x 100'
Room height changed from 8' to 25'
Outside design temperature changed from -5 to 15oF
Number of panels changed from 1 to 6
Panel from 175 to 185oF
-2.7 to -6.0
-5.1 to -0.4
-3.6 to -5.3
-4.2 to -3.4
-0.3 to 0.0
-129-
HEATED FLOOR
TABLE 48 (CONTINUED)
Fl Base case-30'x30'x8' Floor temperature from 81 to 85oF
-6.9 to -7.0
F2 Base case with floor at 84oF
Outside design temperature from 5 to 20oF
-6.5 to -8.1
F3 Base case with floor at 84oF
Floor U-factor changed from 0.07 to 0.15
-6.9 to -7.9
F4 Base case with floor at 84oF
0.5 to 1.25 ACH •6.9 to -7.4
Base case with 15' x F5 15' x 8', 3 inside
walls and one outside wall with half glass
Base case with 15' x F6 15' x 8', 2 inside
walls and 2 outside walls with halfglass
Floor temperature from 83 to 85oF
Floor temperature from 83 to 85oF
-5.9 to -5.9
-7.0 to -7.0
INFRARED MODULAR UNITS
Base case - Infrared, II No reflectors/
deflectors
Surface temperature from 1600 to 1800oF
+2.9 to +2.9
Base case - Infrared, 12 1700oF surface
temperature, No reflectors/ deflectors
Room height changed from 9' to 25'
+2.9 to +10.3
13 Base case - Infrared, 1700oF surface temperature, 15' height, No reflectors /deflectors
14 Base case - Infrared, 1700oF surface temperature, 15' height, 3 ACH, No reflectors/deflectors
1.0 to 4.0 ACH
hc from heater changed up to a factor of 5
1.6 to - 9.4
-6.0
-130-
TABLE 48 (CONTINUED)
ik:
m
15 Base case - Infrared, 9' height, with reflectors/deflectors
Surface temperature from 1600 to I8OO0F
-1.0 to -1.0
16 Base case - Infrared, 15' height, 1700oF, with reflectors/ deflectors
17 Base case - Infrared, 1700oF, 1/2 ACH with reflectors/ deflectors
1.0 to 4.0 ACH
Room height 8' to 25'
•6.2 to -16.8
•3.7 to -4.7
INFRARED - U-TUBES
Ul Base case - Infrared No reflectors/ deflectors
Tube surface changed from 900oF to 700oF
+3.5 to +3.5
U2 Base case - Infrared, 750oF tube temp., No reflectors/ deflectors
Room height changed from 9' to 20'
+3.5 to -2.3
U3 Base case - Infrared with reflectors/
Tube surface changed from 700oF to 900oF
•1.2 to -1.1
U4 Base case - Infrared, 750°F tube temp, with reflectors/ deflectors
Room height changed from 9' to 20'
•1.1 to -3.3
tLi
U5 Base case - Infrared 750oF tube surface, 15' height with reflectors/ deflectors
1.0 to 4.0 ACH -6.5 to -17.0
-131-
space considering the actual conduction through the -walls and the infiltration
load based on the air change method. Table 48 describes the basic cases from
the previous tables, gives the variable being changed and its range of change,
and the percent difference in the two design heat loss calculations. Each of
the types of heating systems are identified such as: Forced Air Heating - CI
- C7, Panel Heating - PI - P16, Heated Floor - Fl - F6, Infrared Modular Units
- 1 1 - 1 7 , and Infrared U-Tubes - Ul - U5.
This summary in Table 48 shows that the ASHRAE design heat loss calcula
tion can oversize a system up to about 17% but the most common value is 4 to
7% oversizing for all of the variables and conditions considered here. For
some situations (C6 and C7) the ASHRAE standard procedure can undersize the
system by up to 15%.
A listing of the computer program which was used to perform all of these
calculations is given in APPENDIX-C. A list of the input variables is also
given there.
-132-
6.0 - DESIGN PROCEDURES
In the previous section, 5.0 - CALCULATION OF DESIGN HEATING LOADS, it was
shown in Table 48 that the use of the ASHRAE Design Heating Load Procedure
(HLD) [1] would result typically in a slightly oversized heating system.
Examination of Table 48 shows for panel heating systems (P-l - P16) that for
variations in panel temperature, room size, and room height that this oversiz-
ing is about 3-6% for 0.5 ACH. However, for larger infiltration rates this
oversizing can be up to 15% for 4.0 ACH. In a similar way for the heated
floor situation the oversizing is about 7% at 0.5 ACH. Likewise, for modular
and U-tube infrared (high and medium temperature respectively) units with good
reflectors and proper location such that no direct infrared radiation impinges
on the walls, this oversizing is up to 5% at 0.5 ACH. If the infiltration
rate is at 4 ACH, the oversizing can be up to 17%.
The conclusion of this investigation is that the air infiltration rate is
the only variable in the design heat loss calculation which affects in a
meaningful way the results for the sizing of radiant heating units. In
Figure 28, the percent reduction of standard design load is plotted against
the air infiltration rate for panel heating, heated floor and infrared modular
and U-tube units-. A single line has been drawn through the data to represent
all four types of radiant heating systems such that at 0.5 ACH the reduction
in heating unit size is 4%, at 1.0 ACH it is 5 1/2%, at 1.5 ACH it is 7 1/2%,
at 2.0 ACH it is 9 1/2%, at 3.0 ACH it is 13% and at 4.0 ACH it is 16%. The
variation shown in Figure 28 is recommended as the only reduction factor to
use for sizing radiant heating systems.
It should be pointed out that the ASHRAE Standard Design Heating Load
procedure (HLD) also does some overestimating (up to 7%) for forced air heat
ing systems. However, as the infiltration rate increases and a temperature
gradient is present the HLD procedure can underestimate the required heater
-133-
r
o
20
18
' 16
g 14 M W Q
H CO
b O
O M H O & P W
H 25 W
u w
12
10
~ • i • r Panels at 130, 150, 170 °F
Infrared, High Temperature Unit, 1700°F
Infrared, Medium Temperature Tubes, 750°F
AIR CHANGES PER HOUR
FIGURE 28. PERCENT REDUCTION OF STANDARD DESIGN LOAD AS A FUNCTION OF AIR INFILTRATION RATE
size by up to 16% for a 25' high room.
From these results, it is recommended for radiant heating systems that the
ASHRAE design heat loss calculation procedure presented in Chapter 25 of the
Handbook of Fundamentals [1] be used with a reduction in the final value made
according to the estimated infiltration rate as presented in Figure 28. There
is not any overwhelming evidence to reduce the design heat loss values for any
of the other parameters in the radiant heating situation. The following
procedures are suggested for designing radiant heating and cooling systems.
6.1 - Radiant Ceiling Panel Heating Systems
In Chapter 8 of the 1984 ASHRAE Systems Handbook [2], there is a section
on "Panel Heating System Design" and this includes data and examples for metal
ceiling panels, warm water panels with embedded pipe (plaster ceiling and
concrete ceiling), and electric ceiling panels. APPENDIX D contains a repro
duction of Chapter 8 from the 1984 ASHRAE Systems Handbook.
The current design steps and procedures given in Chapter 8 of 1984 Systems
have been reviewed and compared with the original work in the literature which
is presented in APPENDIX B - ANNOTATED BIBLIOGRAPHY. There were no serious
problems or difficulties with the assumptions made in compiling the curves and
tables used in the procedures. In addition, various manufacturers have used
these procedures for many years and have not reported any deficiencies in- the
design procedures.
The following are the recommended design steps for panel heating systems.
The only change from what appears in Chapter 8 of the 1984 Systems Handboook
is that the design heat loss is reduced as a function of infiltration rate as
given in Figure 28. Some additional recommendations are also added.
Panel HeatingrSystem Design Steps
1. Calculate the hourly rate of heat loss for each room using the
procedures given in Chapter 25 of the 1985 ASHRAE Handbook of
-135-
Fundamentals fll. Reduce this heat loss by the amount given in
Figure 28 for the specific estimated air infiltration rates.
2. Determine the available area for panels in each room.
3. Calculate the required unit panel output.
4. Determine the required panel surface temperature.
5. Select the means of heating the panel and the size and location
of the heating elements.
6. Select insulation for the reverse side and edge of panel
7. Determine panel heat loss and required input to the panel.
8. Determine any other temperatures that are required.
9. Design the system for heating the panels according to
conventional practice.
In the design steps, the effect of each assumption or choice on
comfort should be considered carefully. Always consider the manufacturers
recommendations for pre-engineered heating panel systems. The following
general rules should be followed:
1) Place panels near cold areas where the heat losses occur.
2) Do not use high temperature ceiling panels in very low ceilings.
3) Keep floor temperatures at or below 85 F (29 C).
Specific design examples are given in Chapter 8 of the 1984 ASHRAE Systems
Handbook which appears in APPENDIX D.
6.2 - Radiant Ceiling Panel Cooling Systems
The computer procedure which was developed and discussed in SECTION - 5.0
and presented inAPPENDIX-C has been applied to several cases of radiant panel
cooling. The calculation procedure was not able to calculate the actual out
side wall or glass temperature for summer conditions since it did not consider
solar effects on the wall or glass. The procedure was developed basically for
heating design load calculations where solar effects would not be considered
-136-
at the design time. The analysis for the procedure for sizing radiant panel
cooling systems involved examination of the original ASHRAE research work and
the procedure for panel cooling given in Chapter 8 of the 1984 ASHRAE Systems
Handbook. This appears to be sufficient for the cooling situation since in
that case the portion of sensible heat removed by radiation is significantly
less since surface temperature differences are less in the cooling mode than
in the heating mode. In addition, the infiltration load for summer design is
expected to be significantly less since the inside - outside air temperature
difference (stack effect) is much smaller and the summer design wind velocity
(typically 7 1/2 mph) is typically half of the winter design wind velocity (15
mph) . Also, the radiant cooling system is not able to absorb the latent load
so that the ventilation air brought to the cooling space absorbs this latent
load and at the same time absorbs some of the sensible load. For these rea
sons, the correction to the design load given in Figure 28 does not apply for
the design cooling load. It is recommended that the ASHRAE design cooling
load procedures (for commercial buildings and residential buildings) presented
in Chapter 26 of the 1985 ASHRAE Handbook of Fundamentals" [1] be used
directly. There are no engineering and/or comfort reasons to expect a reduc
tion in the design cooling load calculations. In addition, there has not been
any research located since the ASHRAE work in the 1950's that would invalidate
the current design procedures.
The procedures given in Chapter 8 of the 1984 ASHRAE Systems Handbook [2]
were checked and verified with the references and original work done by
ASHRAE. Also, several manufacturers have been using this procedure for
many years and have not reported any deficiencies in the procedure. The
procedure is as follows.
-137-
PANEL COOLING SYSTEM DESIGN
1. Determine the room design dry-bulb temperature, relative humidity
and dewpoint.
2. Calculate room sensible and latent heat gains using the ASHRAE
procedure given in Chapter 26 in the 1985 ASHRAE Handbook of
Fundamentals.
3. Select mean water temperature for cooling.
4. Establish minimum supply air quatity.
5. Calculate the latent cooling available from the air.
6. Calculate the sensible cooling available from the air.
7. Determine panel cooling load.
8. Determine required panel area.
Now, design for the heating situation.
9. Calculate room heat loss using the procedures given in Chapter 25
of the 1985 Handbook of Fundamentals f 11 . Reduce this heat loss
by the amount given in Figure 28 for the specific estimated air
infiltration rate.
10. Select mean water temperature for heating.
11. Determine panel area for heating.
12. Determine water flow rate and pressure drop.
13. Design the panel arrangement.
14. Always consider the manufacturers recommendations for placement,
sizing and insulation of pre-engineered panel heating and cooling
systems.
Specific design examples are given in Chapter 8 of the 1984 ASHRAE Systems
Handbook which is reproduced in APPENDIX D.
In the evaluation of this design procedure, it is seen that the location
of the air diffusers relative to the panel sections does not enter into the
-138-
design procedure. It would be expected that if the air is diffused in close
proximity to the ceiling panels that the cooling and heating performance would
be altered slightly. For typical design situations, this should not alter the
design procedure. It should also be pointed out that the lighting load for
the cooling case should be carefully evaluated since it will be a major con
tributor to the cooling load. The design procedure does not account for dif
ferent types of lighting fixtures and these loads should be incorporated into
the design heat gain calculation. Also, it should be emphasized that the
latent heat gain must be absorbed by an independent source. The source spe
cified in the design procedure is the ventilation air which is dehumidified
separately. It would also be possible to operate various types of dehumidi-
fiers within the space to absorb this latent load or to use chemical dehumi-
dification.
6.3 - Heated Floor Systems
A design procedure for heated floors (concrete floor panels for slab-on-
grade, concrete floor panels for intermediate slabs, and electric floor slab
heating) is presented in Chapter 8 of the 1984 ASHRAE Systems Handbook T21.
The only modification to be made to that procedure is the reduction in the
design heat loss calculation shown in Figure 28 for high infiltration air
changes per hour, as indicated in SECTION 6.1. There are several design -
examples given in Chapter 8 of the 1984 ASHRAE Systems Handbook f21 which is
included in APPENDIX - D.
There were some papers obtained in the literature search which discussed
the physical parameters (slab thickness, tube spacing and lower insulation) in
the design process and their effect on the upward heat delivered by the sys
tem. The actual design process for radiant systems was not part of this pro
ject. Here, we are only interested in the design heat loss calculation and
the placement of the heating surfaces.
-139-
Grammling [19] in a 1985 ASHRAE paper pointed out that in the German
standard DIN 4725, methods were developed for testing the thermal performance
of hydronic floor-heating systems. In addition, numerous measurements have
been made for these types of heating systems. Most of the systems have been
tested with the so-called plate apparatus. The results of the tests show that
measured values of performance differ significantly from figures published in
the literature or company catalogs. It is clear from these results that exact
performance measurements under controlled thermal conditions are necessary for
designing and laying out unique floor-heating systems.
Hogan [20] in a MS thesis reviewed and evaluated the ASHRAE design recom
mendations given in Chapter 8 of the 1984 ASHRAE Systems Handbook f21. This
was done using a steady state and a transient numerical model of the heated
floor slab. The ASHRAE panel heating model does not represent the panel heat
loss mechanisms correctly but the design recommendations are adequate and
slightly conservative for designing both bare and covered radiant floor-
heating panels with no infiltration and an AUST equal to the room air tempera
ture. These design recommendations are conservative because both the downward
and edgewise heat loss and panel thermal resistance are over estimated. These
conclusions were also presented in an ASHRAE paper by Hogan and Blackwell
[21].
Some similar conclusions were presented by Shamsundar, Lienhard and Tez-
duyar [22] in an UNPUBLISHED private report in 1985. They have shown that the
ASHRAE procedure is erroneous (using numerical simulation) and that it can be
modified to make it more correct. Some of the error that they note is on the
conservative side and some of it is underestimating the requirements so that
the errors appear to cancel each other for most conditions. This is most
likely why it has not been detected in existing designs. However, since this
is an unpublished report it is not a valid source of information for changing
-140-
the current design procedure. They indicate that the ASHRAE procedure can
also be used for systems with plastic pipe by using simple multipliers for
various pipe diameters.
6.4 - Hiph and Medium Temperature Infrared Systems
The design guidelines provided by manufacturers of infrared heating sys
tems (14 were made available) have been reviewed. They cover both gas and
electric as well as various intensity levels (porous refractory, radiant tube,
quartz tube, and metal sheath electric). The design guidelines for all of
these units are very similar, with minor variations between manufacturers.
These begin with a heating survey taking note of building materials, design
temperatures, usage schedules, combustible or potentially toxic vapors in the
building and restrictions for moisture level requirements. A standard ASHRAE
heat loss calculation is suggested along with a reduction recommendation ran
ging from 0 to 25% with the usual value being about 15%. Various reasons are
given for this reduction in design heat load: unvented gas fired units are
more efficient, less heat loss due to a reduction in air temperature strati
fication in the building, and a lower air temperature required for comfort
when radiant energy is used for heating.
The size of the heaters are then selected from the manufacturers published
data. These units are usually mounted along the perimeter where the high heat
losses occur. They have specific types of reflectors and are mounted at var
ious angles so as to prevent radiating the wall. Specific details are given
concerning mounting height, lateral spacing of heaters, and distance from
combustible materials.
At some point in the design of the unvented gas units, consideration of
minimum dilution air to control the CO2 and condensation possibilities must be
taken into account. Providing the necessary make-up air and exhaust for these
systems is an extremely important consideration.
-141-
A design procedure is presented below for gas and electric infrared heat
ers. It must be stated, however, that the designer or engineer must follow
the manufacturers design and layout suggestions in order to be protected by
their guarantee. The following basic steps are suggested as a design proce
dure.
1. Determine if the building and/or operations are suitable for
infrared heaters. Do not install units where combustible vapors
are present.
2. Calculate building transmission losses using ASHRAE design procedures
in Chapter 25 of the 1985 ASHRAE Handbook of Fundamentals [1].
3. Compute the air infiltration and any forced ventilation loads using
procedures given in Chapter 22 of Ref. 1.
4. Calculate total heat
loss by adding together the transmission losses and infiltration and
ventilation losses. Reduce this number based on the information
given in Figure 28 and the estimated infiltration air changes
per hour.
5. Select heater size or sizes and type of control. This should take
into account the mounting height, reflector style, clearance to
Combustible materials and general layout of the building. Take
into account the manufacturer's recommendations and requirements.
Select type of control suitable for the heaters and the specific
application.
6. Determine the number of heaters by dividing the total load from
(4) by the heater size selected in (5).
7. Determine heater placement using the manufacturers suggestions
regarding mounting height, distance from combustibles, reflector
design (they should be designed and placed so that no direct
-142-
infrared radiation falls on the walls and that the floor is covered
with direct infrared radiant energy in proportion to the building
heat loss), and building dimensions. Perimeter mounted heaters are
usually angled toward the interior of the building (at about 30 deg.)
and heaters in the interior of the building are usually mounted
horizontal with appropriate reflectors. They should avoid interior
obstructions such as cranes sprinkler systems, storage racks, fork
truck travel, and light fixtures.
8. Determine the method of mounting heaters using manufacturers
recommendations. Use manufacturers recommendations concerning
mounting devices, brackets, flexible gas lines and flexible electrical
conduit. Avoid having heaters too close to structural members.
Always conform to local codes.
9. Select and locate thermostats to control zone loads and provide uni
form heating. They should be mounted according to the manufacturers
recommendations. Generally, they should be about 5 feet from the
floor, out of direct view of the heaters, and not in direct contact
with cold outside walls.
10. Determine minimum air needed to dilute CO2 in unvented units to a safe
level. In unvented gas fired systems check for condensation
possibilities. Check design values of inside surfaces and roof with
the deWpoint temperature of the inside air.
11. Comply with all of the manufacturers installation and operation
instructions.
6.5 - Other Design Procedures
A few other suggestions were found in the literature for modification of
the design heat load for a radiant heating system when the design heat load is
calculated for a forced air system. Mclntyre [23 and 24] used a theoretical
-143-
computer program and a simplified method in the CIBS Guide to compare the
difference in design heating load for radiant and warm.air systems. In addi
tion, he also looked at estimated energy requirements for the various types of
systems. He came to the following conclusions: (1) the simplified CIBS Guide
method and the computer model show very good agreement, (2) theoretical
studies showed little difference (5%) in the power required to maintain com
fortable conditions in residential size and types of rooms with either radiant
or warm air heating, (3) radiant heating was more economical (5-20%) than
forced air heating systems in large spaces with high infiltration rates.
In another paper by Harrison [25], the differences between design heating
loads for convective and radiant systems in discussed. These contain multi
plying factors to be applied to transmission losses and air change losses.
These are all based on theoretical calculations and show the same trend of a
decrease in the design heat loss as infiltration increases and an increase in
the design heat loss as more radiant energy falls on the walls, floors or
ceiling. For about the greatest change illustrated, a combined modifier of
0.82 was given. This is comparable with the value of a 17% reduction given in"
Figure 28.
In another unpublished discussion of this problem, the author calculates a
percent reduction in the design heat loss of about 12% at one air change per
hour. This is somewhat higher than what others have calculated, however, this
model is considerably different than what others propose.
-144-
7.0 - SUMMARY OF MANUFACTURERS SURVEY
Over 320 letters have been sent to the following groups: members and
corresponding members of TC 6.5; equipment and instrument manufacturers or
suppliers as well as installers of radiant heating/cooling equipment; labora
tories, consultants, and trade associations. Approximately 75 positive res
ponses have been received from these inquiries. The majority of the informa
tion has been specific details about products and some information on design
procedures. Approximately 15 to 20 responses were for medium and high tem
perature electric and infrared units. About 10-15 responses concerned radiant
ceiling/floor panels. The remainder of these responses discussed various
aspects of the radiant heating field such as controls, materials of construc
tion, measurement instrumentation and miscellaneous items. This information
has been reviewed and has been incorporated into this report.
-145-
8.0 - SYSTEM DYNAMICS
System dynamics for heating and cooling enter into the calculations only
when actual operation is considered and not when design heating loads are
being calculated. The dynamics of the system are important during transient
load situations in order to estimate comfort conditions and energy require
ments. This is discussed in several references [19,21,26,27,28]. The heated
concrete floor and embedded heaters in plaster ceilings were found to present
the slowest response times and require more sophisticated control systems to
account for temperature lag. The hydronic metal ceiling panels and high tem
perature ( infrared) systems do not appear to present any dynamic problems if
properly designed and controlled.
-146-
9.0 - RESEARCH NEEDS
There have been six areas which were identified as requiring additional
research information. Four of these areas enter directly into the
calculations which have been made in this investigation.
9.1 - Convection Coefficients.
There is insufficient information available on the effect of geometrical
considerations for convection coefficients from surfaces at temperatures dif
ferent than the air temperature. When surfaces (panel heating units or glass
surfaces) at different temperatures than a surrounding surface in the same
plane are present, reliable data does not seem to be available. It is esti
mated that this unknown effect could influence the results by 5 to 10%. There
is also a lack of reliable data for the size effect of surfaces exchanging
heat by convection with the air (tall walls, long thin heating surfaces).
Further research needs to be done in these areas.
9.2 - Air Temperature Stratification.
There appears to be a lack of information on what effect heating and cool
ing surfaces have on air temperature stratification. Some measurements have
been made [18] for specific situations , but no general calculation criteria
appear to be available.
9.3 - Surface Emissivities.
There do not appear to be reliable data for surface emissivities of
radiant heating and cooling surfaces. Most manufacturers do not present this
data and some estimate it at between 0.87 and 0.95. This value has a strong
effect on the area of heaters required for radiant systems.
9.4 - Comfort During Radiant Temperature Asymmetry.
There is some data available [7,14] on the effects on comfort due to
radiant temperature asymmetry , but additional information is needed. It
would be useful to have design charts giving values of radiant temperature
-147-
asymmetry for specific geometries or giving limiting values of heat flux
(Btu/hr ft^) for comfort conditions. Some additional work is needed in this
area.
9.5 - Radiant System Dynamics.
Very little information is available on radiant heating system dynamics
and how this affects the energy requirements for radiant heating and cooling
systems. Some applied types of heat transfer analyses involving the interac
tions between convection and radiant exchanges at surfaces with the material
having a heat capacity are needed.
9.6 - Heated Floor System
With the reported discrepancies in the ASHRAE design procedure [20,21,22]
additional work should be done to correct the estimations of upward heat floWj
downward heat flow, and edge heat losses for different geometries and tube
spacings.
-148-
10.0 - REFERENCES
ASHRAE, Handbook of Fundamentals. American Society of Heating,
Refrigerating and Air Conditioning Engineers, Atlanta, GA, 1985.
ASHRAE, 1984 Systems Handbook. American Society- of Heating,
Refrigerating and Air Conditioning Engineers, Atlanta, GA, 1984.
ASHRAE, 1983 Equipment Handbook. American Society of Heating,
Refrigerating and Air Conditioning Engineers, Atlanta, GA, 1983.
Fanger, P. 0., Theraml Comfort - Analysis and Applications in
Environmental Engineering. McGraw-Hill Book Co., New York, NY, 1972.
McAdams, W. H. , Heat Transmission. McGraw Hill Book, Co., 1954.
Incropera, F. P. and DeWitt, D. P., Fundamentals of Heat and Mass
Transfer. Second Edition, John Wiley and Sons, NY, 1985.
Olesen, B. W. and Nielsen, R. , "Radiant Spot Cooling of Hot Working
Places", ASHRAE Trans., V. 87, Pt. 1, 1981.
McNall, P. E. , Jaax, J. Rohles, F. H. , Jr., and Springer, W. E.,
"Thermal Comfort Condition for Three Levels of Activity", ASHRAE
Trans.. Vol. 73, Pt. 1, 1967.
Olesen, S., Bassing, J. J. and Fanger, P. 0., "Physiological Comfort
Conditions at Sixteen Combinations of Activity, Clothing, Air Velocity
and Ambient Temperature", ASHRAE Trans.. Vol. 78, Pt. II, 1972.
McNall, P. E. , Jr. and Schlegel, J. C , "The Relative Effects of
Convection and Radiation Heat Transfer on Thermal Comfort for
Sedentary and Active Human Subjects", ASHRAE Trans.. Vol. 74,
Pt. II, 1968.
Rohles, F. H., Jr., Woods, J. E., and Nevins, R. G. , "The
Influence of Clothing and Temperature on Sedentary Comfort",
ASHRAE Trans., Vol. 79, Pt. II, 1973.
149-
12. Rohles, F. H. , Jr., Woods, J. E. , and Nevins, R. G., "The Effects
of Air Movement and Temperature on the Thermal Sensations of
Sedentary Man, ASHRAE Trans.. Vol. 80, Pt. I, 1974.
13. Griffiths, I. D. and Mclntyre, D. A., "Radiant Heating and Comfort",
Building Services Engineer, Vol. 40, June, 1972.
14. Fanger, P. 0., "Radiation and Discomfort", ASHRAE Journal. February,
1986.
15. Iso, 7730, "Moderate Thermal Environments - Determination of the
PMV and PPD Indices and Specification of the Conditions for Thermal
Comfort", International Standards Organization, Geneva, 1984.
16. ASHRAE, Procedure for Determining Heating and Cooling Loads
for Computerized Energy Calculations - Algorithms for Building
Heat Transfer Subroutines. American Society of Heating,
Refrigerating and Air Conditioning Engineers, Atlanta, GA, 1975.
17. Altmayer, E. F. , Gadgil, A. J., Baumann, F. S. and Kammerud, R. S.,
"Correlations for Convective Heat Transfer from Room Surfaces", ASHRAE
Trans.. Vol. 89, Pt. 2A, 1983.
18. Min, T. C. , Schutrum, L. F. , Parmelee, G. V. andVouris, J. D., "Natural
Convection and Radiation in a Panel Heated Room", Heating. Piping and Air
Conditioning. May, 1956.
19. Grammling, F. J., "Methods for Testing Hydronic Floor Heating Systems",
ASHRAE Trans., V. 91, Pt. 2, 1985.
20. Hogan, R. E., Jr., "Heat Transfer Analysis of Radiant Heating Panels -
Hot Water Pipes in Concrete Slab Floor", M.S. Thesis, Louisiana Tech.
University, August, 1979.
21. Hogan, R. E. , Jr., and Blackwell, B. , "Comparison of Numerical Model with
ASHRAE Design Procedure for Warm Water Concrete Floor Heating Panels",
ASHRAE, Trans., Vol. 92, Pt. 1, 1986.
-150-
22. Shamsundar, N. , Lienhard, J. H. and Tezduyar, T. E., "Performance of
Polybutylene Pipe in Concrete Heating Panels", Report No. 2, Department
of Mechanical Engineering, University of Houston, Houston, TX, 1985
(UNPUBLISHED).
23. Mclntyre, D.A., "Warm Air and Radiant Heating: Steady State Power
Requirements", Electricity Council Research Centre, Capenhurst,
England, Dec, 1980, (NTJS - PB83 - 231506)
24. Mclntyre, D.A., "Warm Air or Radiant Heating?", Building Research and
Practice. Vol. 17, No. 1, Pg. 48-57, Jan-Feb., 1984.
25. Harrison, E., "The Calculation of the Heat Requirements of Rooms",
The Building Services Engineer. Vol. 43, pg. 19-23, May, 1975.
26. Algren, A. B. and Ciscel, B., "Heating Panel Time Response Study",
Heating. Piping and Air Conditioning. March, 1949.
27. Berglund, L. , Rascati, R. , and Markel, M. , "Radiant Assisted Comfort
Heating For Energy Conservation in Intermittently Occupied Spaces",
Proceedings of the CIB W67 Third International Symposium. Dublin,
Ireland, 1982.
28. Boyar, R. E., "Room Temperature Dynamics of Radiant Ceiling and Air
Conditioning Comfort Systems", Trans. ASHRAE, V. 69, 1963.
-151-
APPENDIX A
BIBLIOGRAPHY
II
A - l
Adlam, T. N. , Radiant Heating. The Industrial Press, New York, NY, 1949.
Aiulfi, D., Fort K., Ottin, T., "Modelization of Floor Heating and Oil Furnaces for the Unilization of Microprocessors in DDC," Clima 2000 -Heating. Ventilating and Air-Conditioning Systems. Vol. 6, W S Kongres - W S Messe, 1985.
Albert!, M. and Rugger, R. , "A Method to Check Thermal Comfort Conditions in High Industrial Buildings Provided with Radiant Panels Heating Plant", Clima 2000 - Indoor Climate. Vol. 4, W S Kongres - W S Messe, 1985.
Alexander, J. C , "Calculations of Direct Energy Losses from Ceiling Mounted Radiant Heating Panels to Fenestrated Areas," Energy Engineering: Journal of Association of Energy Engineers, Vol. 78, No. 5, pg. 35-48, August-September 1981.
Algren, A. B., and Ciscel, Ben, "Heating Panel Time Response Study," Heating Piping & Air Conditioning. March 1949.
Algren, A. B. , Snyder, E. F., Jr., Head, R. R., "Field Studies of Floor Panel Control Systems - Part II,", Heating. Piping and Air Conditioning. April 1954.
Algren, A. B., Snyder, E. F., Jr., Locke, J. S., "Field Studies of Floor Panel Control Systems," Heating. Piping and Air Conditioning. February 1953.
Altmayer, E. F. , Gadgil, A. J., Bauman, F. S., and Kammerud, R. C , "Correlations for Convective Heat Transfer from Room Surfaces," ASHRAE Trans., Vol. 89, pt. 2A, 1983.
Ashley, J. L. , Correa, E. and Canfield, K., "Energy Conservation: Heating Navy Hangars", Technical Report R-910, Naval Civil Engineering Laboratory, July 1984.
ASHRAE, "High Intensity Infrared Radiant Heating - Chap. 18", 1984 Systems Handbook. American Society of Heating, Refrigerating and Air Conditioning Engineers, Atlanta, GA, 1984.
ASHRAE, "High Intensity Infrared Heaters - Chap. 30", 1983 Equipment Handbook. American Society of Heating, Refrigerating and Air Conditioning Engineers, Atlanta, GA, 1983.
ASHRAE, "Panel Heating and Cooling Systems - Chap. 8", 1984 Systems Handbook. American Society of Heating, Refrigerating and Air Conditioning Engineers, Atlanta, GA, 1984.
ASHRAE, Procedure for Determining Heating and Cooling Loads for Computerizing Energy Calculations - Algorithms for Building Heat Transfer Subroutines. American Society of Heating, Refrigerating and Air Conditioning Engineers, Atlanta, GA., 1975.
ASHRAE Task Group, "High-Intensity Infrared Heaters," ASHRAE Journal, December 1963.
A-2
Bahnfleth, Donald R. , "Building Heat with Natural Gas Infrared," ASHRAE Journal, June 1968.
Bahnfleth, D.R., "Physiological Effects of High Intensity Radiant Beam Heating", ASHRAE Jnl.. Nov., 1964.
Bahnfleth, Donald R., "Symposium: Field Performance of Infrared Heating Systems," ASHRAE Journal, June 1968.
Bailey, H. R., "An Experimental Comparison of Energy Requirements for Space Heating with Radiant and Convective Systems", ASHRAE Trans.. V. 86, Pt. 1, 1980.
Bak, Richard, "Hydronic Radiant Floor Heating Staging a Comback", Air Conditioning. Heating and Refrigeration News. March 1985.
Baker, Merl," Effectiveness and Temperature Requirements for Cooling Panels Removing Internal Radiation," Heating. Piping and Air Conditioning. June 1952.
Baker, Merl, "Improved Comfort Through Radiant Heating and Cooling," ASHRAE Journal, February 1960.
Baker, Merl, "Removal of Internal Radiation by Cooling Panels," Heating. Piping. & Air Conditioning. November 1949.
Ball, H. D. and Green, D., "The Impact of Lighting Fixtures on Heating and Cooling Loads - Mathematical Mode", ASHRAE Trans.. V. 89, pt. 2A, 1983.
Banhidi, L. , "Ther Thermotechnical Dimensioning of Radiant Strips and Borders for the Heating of Communal Buildings", Building Science. Vol. 9, 1974.
Banhidi, L., Dr., Somogyi, A., Kintses, G., Besnyo, J., "About Local Discomfort Effects Caused by Asymmetric Radiation Occurring During Winter in Dwelling Houses", Cliraa 2000 - Indoor Climate. Vol. 4, W S Kongres -W S Messe, 1985.
Bauman, F., Gadgil, A., Kammarud, R., Altmayer, E. , Nansteel, M., "Convective Heat Transfer in Buildings: Recent Research Results" ASHRAE Trans, Vol. 89, pt. 1A, 1983.
Baumanns, H., "Gas IR Heaters for Heating Large Spaces," Warme Gas International. Vol. 27, No. 4, Gewea GmbH and Co. Monchengladbach, April, 1978.
Becker, Sidney, "Surface Temperatures of Plaster Ceiling Panels," Heating. Piping. & Air Conditioning April, 1950.
Beier, Richard A. and Gorton, Robert L., "Thermal Stratification in Factories - Cooling Loads and Temperature Profiles," ASHRAE Trans., Vol. 84, Pt. 1, 1978.
Belsey, D. C. and Benseman, R. F., "Spatial Radiation Patterns for Infra-Red Heaters", Building Science. Vol. 3, 1969.
A-3
Benzinger, Theodore H. , Manglum, B. W., and Hill James, "The Design Construction and Operation of a Scanning Radiometer for Measurement of Plane Radiant Temperature in Buildings," ASHRAE Trans., Vol. 82, Pt. 2, 1976.
Berglund, Larry, "Mathematical Models for Predicting the Thermal Comfort Response of Building Occupants," ASHRAE Trans., Vol. 84, Pt. 1, 1978.
Berglund, L. G. and Fobelets, A. P. R., "Subjective Human Response to Low-Level Air Currents and Asymmetric Radiation", ASHRAE Trans.. V. 93, Pt. 1, 1987.
Berglund, L. G. and Gagge, A. P., "Human Response to Thermal Conditions Maintained in an Office by Radiant Ceiling, Baseboard, Forced Air and Floor Heating Systems", ASHRAE Trans., V. 91, Pt. 2, 1985.
Berglund, L. G. and Gagge, A. P., "Thermal Comfort and Radiant Heat", Proceedings of the Third National Passive Solar Conference. American Section of the International Solar Energy Society, Univ. of Deleware, 1979.
Berglund, L. G. , Gagge, A. P. annd Banhidi, L. J., Performance of Radiant Ceiling and Other Heating Systems Controlled for Equal Comfort with an Operative Temperature Sensor", ProceedinEs of the Third International Conference on Indoor Air Quality and Climate. Stockholm, August, 1984
Berglund, L. , Rascati, R. and Markel, M. L. , "Radiant Heating and Control for Comfort during Transient Conditions", ASHRAE Trans.. V. 88, Pt. 2, 1982.
Berglund, L., Rascati, R., Markel, M., "Radiant Assisted Comfort Heating for Energy Conservation in Intermittently Occupied Spaces," Energy Conservation in the Built Environment: Proceedings of the CIB W67 Third International Symposium, Dublin, Ireland, 1982, pg. 98
Best, W. H. "Spot Heating and Condensation Control Using Gas Infrared Systems," ASHRAE Journal, June 1968.
Billington, N. S., "Floor Panel Heating - Some Design Data", Journal of the Institution of Heating and Ventilating Engineers. London, Oct., 1953.
Blossom, J. S., "High Temperature Water Heats New School", Heating Piping and Air Conditioning. March, 1959.
Blossom, J. S., "Pipe HTW Through Classrooms", Heating. Piping and Air Conditioning. July, 1964.
Bohgaki, K., "Effect of Floor Heating on Man's Comfort and Thermal Sensations," Clima 2000 - Indoor Climate. Vol. 4, W S Kongres - W S Messe, 1985.
Borresen, B. A., "Thermal Room Models for Control Analysis", ASHRAE Trans.. V. 87, Pt. 2, 1981.
Boyar, R. E., "The Influence of Radiant Energy Transfer on Human Comfort", Heating. Piping and Air Conditioning. June 1966.
A-4
Boyar, R. E. , "Room Temperature Dynamics of Radiant Ceiling and Air Conditioning Comfort Systems", Trans. ASHRAE, V. 69, 1963
Boyd, R. L., "Control of Electric Infrared Energy Distribution", Electrical Engineering. Feb. 1963, p. 103.
Boyd, R. L. , "What Do We Know About Infrared Comfort Heating?", Heating. Piping and Air-Conditioning. Nov., 1960, p. 133.
Boyd, Robert, "Application and Selection of Electric Infrared Comfort Heaters," ASHRAE Journal, October 1962.
Boyer, L. L. , "Radiant Panel Effects of Floor - Ceiling Assemblies Incorporating Static, Return, and Supply Penums," ASHRAE Trans., Vol. 74, Pt. 2, 1968.
Braun, D. L. and McNall, P. E., Jr., "A Radiometer for Environmental Applications", ASHRAE Trans.. V. 75, Pt. 1, 1969.
Bruce, H. H. , "Off-Peak Floor Heating-Research, Design and Development: Some Controversial Factors", Electricity and Space Heating. Proceedings of Symposium of the Institution of Electrical Engineers in London, March, 1964.
Bryan, W. L., "Comparative Energy Requirements of Radiant Space Heating", ASHRAE Trans.. V. 87, Pt. 1, 1981.
Bryan, W. L. , "Gas-Fired Radiant Heat", Mech. Engr.. Vol. 87, March, 1965.
Buckley, N. A. , "Engineering Principles Support an Adjustment Factor When Sizing Gas-Fired Low-Intensity Infrared Equipment", ASHRAE Trans.. V. 93, Pt. 1, 1987.
Buckley, N. A., and Seel T., "Gas-Fired Medium - Intensity Radiant Heating Provides a Cost-Effective, Efficient Space Conditioning Alternative," ASHRAE Trans., Vol. 92, pt. 1, No. 4, 1986.
Bull, L. C. , "The Practical Approach to Heat Loss Calculation", Building Services Engineer. Vol. 41, March, 1974.
Cambel, A. B., "Model Study of Radiant Heating", Ph.D. Dissertation, University of Iowa, 1950.
Carroll, J. R., "Radiant Systems for Heating", New Methods of Heating Buildings. BRI-760, National Academy of Sciences - National Research Council, Washington, D. C , 1960.
Carroll, Joseph A., An "MRT Method" of Computing Radiant Energy Exchange in Rooms. Systems Simulation and Economic Analysis Conference Proceedings, January 1980, San Diego, CA.
Chapot, Jean Robert, Manuel Techniques "Avadis". Utilization Chauffage Par Le Sol, Pont-A-Mousson, S. A.
A-5
Chrenko, F. A., "Heated Ceilings and Comfort", Journal of the Institution of Heating and Ventilating Engineers. London, April, 1952.
Chrenko, F. A., "Heated Floors and Comfort". Journal of the Institution of Heating and Ventilating Engineers. London, April, 1955.
Chrenko, F. A., "Radiant Heat and Thermal Comfort", Electricity and Space Heating. March, 1964.
Cohn, Lisa, "Radiant Heating Units Net Big Savings in Special Cases," Energy User News, Vol. 7, No. 47, November 1982.
Collicott, H. E., Fontaine, W. E., and Grosh, R. J., "Free Convection and Radiation Heat Transfer from Cylindrical Fins," ASHRAE Journal, Dec. 1965.
Collicott, H. E. , Fontaine, W. E. , and Witzell, 0. W. "Radiation and Free Convection Heat Transfer from Wire and Tube Heat Exchanges," ASHRAE Journal, Dec. 1963.
, "Concrete Code Restricts Panel Heating," Heating. Piping. & Air Conditioning. October 1951.
, "Conserving Energy with Infrared Heating," Plant Engineering. Vol. 31, No. 8, pg. 155 +, Technical Pub. Co., April, 1977.
Correa, Edward L., Design Guidelines for Heating Aircraft Hangars with Radiant Heaters. Naval Civil Engineering Lab, December, 1983.
Danter, E. , "Heat Exchanges in a Room and the Definition of Room Temperature", The Building Services Engineer. Vol. 41, Feb., 1974.
DeHeer, T. and Erkerlens, H. J., "Heat Transmission by Radiation", Journal of the Institution of Heating and Ventilationg Engineers. London, Jan., 1963.
DeWerth, D. W. , "A Study of Infra-Red Energy Generated by Radiant Gas Burners", Research Bulletin No. 92. American Gas Association, Nov. 1962.
DeWerth, Douglas W., "Gas-Fired Radiant Heat", Mechanical Engineering. Vol. 86, Nov., 1964.
DeWerth, D. W. , "Literature Review of Infrared Energy Produced with Gas Burners", Research Bulletin 83, American Gas Association Laboratories, Cleveland, Ohio, 1960.
Diamant, R. M. E., "Radiant Heating" The Heating and Air-Conditioning Journal ASHRAE Journal, Vol. 50, No. 577, pg. 26-28, Troop Publ., February, 1980.
Dunkle, R.V., "Configuration Factors for Radiant Heat Transfer Calculations Involving People", ASME Trans.. Jnl. of Heat Transfer, Feb., 1963.
Endreb, B. Von H., and Mommertz W., "Untersuchung von Flachenheiz - Systemen Mit Der Thermoelektrischen Analogie," Elektrowaerme. Vol. 39, No. 4, pg. 203-210, August, 1981.
A-6
Eno, Burton E., "Combined Convection and Radiation from Rectilinear Fins," ASHRAE Trans., Vol. 73, pt. 1, 1967.
Faithfull, E. W., "Electric Floor Warmings in Commercial Buildings", Electricity and Space Heating Proceedings of Symposium of the Institution of Electrical Engineers in London. March, 1964.
Fanger, P. 0., "Calculation of Thermal Comfort: Introduction to a Basic Comfort Equation", ASHRAE Trans.. Vol. 73, Pt. II, 1967.
Fanger, P. 0., "Radiation and Discomfort" ASHRAE Journal, February 1986.
Fanger, P. 0., Thermal Comfort - Analysis and Applications in Environmental Engineering. McGraw-Hill Book Co., New York, 1972.
Fanger, P. 0., Angelius, 0., and Kjerulf - Jensen, P., "Radiation Data for the Human Body," ASHRAE Trans., Vol. 76, Pt. 2, 1970.
Fanger, P. 0., Banhidi, L., Olesen, B. W. and Langkilde, G., "Comfort Limits for Heated Ceilings" ASHRAE Trans., Vol. 86, pt. 2, 1980.
Faucett, J. W. , "Field Evaluation of High Intensity Infrared Space Heating Systems - Research Project RP-98", Final Report, ASHRAE, 1972.
Faust, Frank H., "New Electric Heating Systems", New Methods of Heating Buildings. BRI-760, National Academy of Sciences - National Research Council, Washington, DC, 1960.
Field, A. A. - "Direct Fired Radiant Heating Systems", Heating and Ventilating Engineer, " Vol. 49, No. 571, pg. 6-9, Technitrade Journals Ltd., February 1975.
Ford, J. K., "Deriving Radiation View Factors within a Triangular Cross-Section Residential Attic", ASHRAE, Trans., Vol. 89, pt. 1A, 1983.
, "Fuel Bills Halved After Switch to Infrared Heating System," Energy Management Technology. Vol. 7, No. 8, pg. 44-45, Walker-Davis Publications, November 1983.
, Gas Engineers Handbook. Industrial Press, NY, 1977.
Gagge, A. P., Fobelets, A. P. and Berglund, L. G., "A Standard Predictive Index of Human Response to the Thermal Environment", ASHRAE Trans.. V. 92, Pt. 2, 1986.
Gagge, A. P., Hardy, J. D., and Rapp, G. M., "Exploratory Study on Comfort for High Temperature Sources of Radiant Heat," ASHRAE Trans., Vol. 71, Pt. 2, 1965.
Gagge, A. P., Rapp.G. M., and Hardy, J. D., "The Effective Radiant Field and Operative Temperature Necessary for Comfort with Radiant Heating," ASHRAE Journal, 1967.
A-7
Gagge, A. P., Rapp, G. M. , and Hardy, J. D., "Mean Radiant and Operative Temperature for High-Temperature Sources of Radiant Heat," ASHRAE Journal, October 1964.
Gagge, A.P., "Final Progress Report: RP-41-Physiological Effects of High Intensity Radiant Beam Heating", ASHRAE Jnl.. April, 1968.
, "Gas Radiant Heating Installation Achieves 90% Overall Efficiency" The Heating and Air Conditioning Journal ASHRAE Journal, Vol. 48, No. S60, pg. 60-61, Troup Publ., September 1978.
Gebhart, Benjamin, "A New Method for Calculating Radiant Exchanges," Heating. Piping. & Air Conditioning. Ithaca, New York, July 1958.
Gluck, B., Strahlungsheizung Theorie und Praxis. Verlag C. F. Muller, Karlsruhe, Germany, 1982.
Gorton, R. L. and Leard, A. T. , "A Computer Program for Air Temperature and Cooling Load Determination for Stratitifed-Cooled Industrial Buildings", ASHRAE Trans.. V. 90, Pt. IB, 1984.
Gorton, R. L. and Sassi, M. M., "Determination of Temperature Profiles and Loads in a Thermally Stratified, Air-Conditioned System: Part 1-Model Studies", ASHRAE Trans.. V. 88, Pt., Pt. 2, 1982.
Gorton, R. L. and Sassi, M. M., "Determination of Temperature Profiles and Loads in a Thermally Stratified, Air Conditioned System: Part 2 -Program Description and Comparison of Computed and Measured Results", ASHRAE Trans.. V. 88, Pt. 2, 1982.
Grammling, F. J., "Methods for Testing Hydronic Floor-Heating Systems," ASHRAE Trans., V. 91, Pt. 2, 1985.
Griffiths, I.D. and Mclntyre, D.A., "The Balance of Radiant and Air Temperature for Warmth in Older Women", Environmental Research. 6, 1973.
Griffiths, I. D. and Mclntyre, D. A., "Radiant Heating and Comfort", The Building Services Engineer. V. 40, June, 1972.
Griffiths, I. and Mclntyre, D., "Subjective Response to Overhead Thermal Radiation", Human Factors. 16 (4), 1974.
Guillaume, M., "Integration of Different Energy Saving Possibilities in. Dwellings," Energy Savings in Buildings. Commission of the European Communities, Edited by Ehringer H. and Zito V., November 1983, D. Reidel Publ. Co.
Hannay, J., Liebecq, G., Nusgens, P., "Convective Heat Transfers within a Large Open Plan Office Area: Experimental Results for Dynamic Buildings Simulations," Clima 2000 - Building Design and Performance. Vol. 2, W S Kongres - W S Messe, 1985.
Hardy, James D., "Physiological Effects of High Intensity Infrared Heating," ASHRAE Journal, November 1962.
A-8
Harrington, Keith and Lydon, R. T., Comparison of Building Thermal Analysis Methods. Systems Simulation and Economic. Analysis Conference Proceedings, January 1980, San Diego, CA.
Harris, Robert L., Jr., "Computer Simulation of Radiant Heat Load and Control Alternatives", Journal of American Industrial Hygiene Assoc.. V. 35, Feb., 1974.
Harrison, E., "Calculation of the Heat Requirements of Rooms," The Building Services Engineer. Vol. 43, pg. 19-23, Batiste Publications Ltd., May 1975.
Harrison, E., "Environmental Temperature and the Calculation of Heat Loss", Building Services Engineer. Vol. 41, March, 1974.
Hart, Gordon H., "Heating the Perimeter Zone of an Office Building: An Analytical Study using the Proposed ASHRAE Comfort Standard (SS-74R)," ASHRAE Trans., Vol. 87., pt. 2, 1981.
Hazard, W. G., "Radiant Heat Control in Industrial Plants," ASHRAE Journal, November 1959.
Heath, George, "Basic Infrared Heating Applications," ASHRAE Journal, Dec. 1967.
Hedgepeth, Lloyd M., and Sepsy, Charles, "A Thermodynamic Simulation of a Building Environmental Control System," ASHRAE Trans., Vol. 78, Pt. 2, 1972.
Heerwagen, D. R. , Kippenhan, C. J., Emery, A. F. , and Varey, G. B, "Developing Office Building Design and Operation Strategies Using VWENSOL and the Comfort Routine", ASHRAE Trans.. V. 86, Pt. 1, 1980.
Herrington, L. P., and Lorenzi, R. J., "Effect of Panel Location on Skin and Clothing Surface Temperature, "Heating. Piping. & Air Conditioning. October 1949.
Hogan, Roy Edward, Jr., Heat Transfer Analysis of Radiant Heating Panels - Hot Water Pipes in Concrete Slab Floor. MS Thesis, Louisiana Tech. University, August 1979.
Hogan, R. E., Jr. and Blackwell, B., Ph.D., "Comparison of Numerical Model with ASHRAE Design Procedure for Warm Water Concrete Floor Heating Panels," ASHRAE Trans., Vol. 92, No. 2, Pt. 1, 1986.
Holden, T.S., "Calculation of Incident Low Temperature Radiation," ASHRAE Journal, April 1961.
Houghten, F. C , et. al., "Heat Loss Through Basement Floors and Walls", ASHVE Trans.. Vol. 48, 1942.
Houghten, F. C , Gunst, S. B. and Suciu, J., "Radiation as Factor in the Sensation of Warmth", ASHVE Trans.. Jan., 1941.
A-9
Howarth, E.S., Huddleston, S.C., and Koch, R.M., "Aluminum Ceiling Panels for Heating and Cooling," Heating. Piping. & Air Conditioning. September 1951.
HPAC Engineering Data File, "Design and Control of High Temperature Hot Water Heat Consumers", Heating Piping and Air Conditioning. June, 1960.
Hulbert, L.E., Nottage, H.B., and Franks, C.V., "Heat Flow Analysis in Panel Heating or Cooling Sections," Heating. Piping & Air Conditioning. April 1950.
Humphreys, C. M., Franks, C.V., Schutrum, L.F., "Field Studies of Heat Losses from Concrete Floor Panels," Heating. Piping and Air Conditioning. January 1951.
Humphreys, C. M., Franks, C. V., Schutrum, L. F., "Laboratory Studies of the Thermal Characteristics of Plaster Panels," Heating, Piping and Air Conditioning, July 1951.
Humphreys, CM., Nottage, H.E., Franks, C.V., Huebscher, R.G., Schutrum, L.F., and Locklin, D.W. "Laboratory Studies on Heat Flow Within a Concrete Panel," Heating. Piping & Air Conditioning. April, 1950.
Hunter, Robert K., "Controllable and Efficient Infrared Radiant Heating", Automation. October 1972.
Hutchinson, F. W. , A Graphical Design Procedure for Radiant Panel Heating. Revere Copper & Brass, Inc., New York, NY, 1948.
Hutchinson, F. W., Design of Heating and Ventilating Systems. The Industrial Press, New York, 1955.
Hutchinson, F. W., "Influence of Gaseous Radiation in Panel Heating", ASHRAE Trans.. Vol. 53, 1947.
Hutchinson, F. W. , "Influence of Gaseous Radiation in Panel Heating", Trans. ASHVE. 1947.
Hutchinson, F.W. and Baker, Merl, "Optimum Panel Surface Distribution Determined from Human Shape Factors," Heating, Piping and Air Conditioning, June 1951.
Hutchinson, F.W., Mills, D. L., and La Tart, L. J., "Losses from a Floor-Type Panel Heating System," Heating. Piping & Air Conditioning. December 1950.
, "In Old Building - New Infrared Heating System Solves Old Problems", Air Conditioning Heating and Refrigeration News. April, 1985.
IHVE, Estimation of Plant Capacity. A9, IHVE Guide, 1975.
IHVE, "Thermal and Other Properties of Building Structures", A3, IHVE Guide, 1977.
Irwin, R.R., "Panel Cooling for a Residence," Heating. Piping & Air Conditioning. May 1955.
A-10
Janssen, John E., "Field Evaluation of High Temperature Infrared Space Heating Systems," ASHRAE Trans., Vol. 82, Pt. 1, 1976.
Johnson, T. E., "Radiation Cooling of Structures with Infrared Transparent Wind Screens", Proceedings of the Second Workshop on the Use of Solar Energy for the Cooling of Buildings, UCLA, August, 1975.
Khudenko, A.A., "Radiation Characteristics of Gas Infrared" Journal of Engineering Physics. Vol. 29, No. 5, Pg. 1395-1400, Consultants Bureau.
Korsgaard, V., "Necessity of Using a Directional Mean Radiant Temperature to Describe the Thermal Conditions in Rooms", Heating. Piping and Air Conditioning, June, 1949.
Korsgaard, V., "A New Radiometer Measuring Directional Mean Radiant Temperatures", Heating Pining and Air Conditioning. July, 1949.
Kweller Esher, "Criteria for Mechanical Energy Saving Retrofit Options for Single Family Residences," New and Existing Single Family Residences. ACEEE 1984 Summer Study on Energy Efficiency in Buildings.
Langkilde, G., Gunnarsen, L.,.Mortensen, N., "Comfort Limits During Infrared Radiant Heating of Industrial Spaces," Clima 2000 - Indoor Climate. Vol. 4, W S Kongres - W S Messe, 1985.
Lebrun, Jean J., and Marret, Dominique J., "Thermal Comfort and Energy Consumption in Winter Conditions -- Continuation of the Experimental Study" ASHRAE Trans., Vol 85, Pt. 2, 1979.
Leopold, C. S., "The Mechanism of Heat Transfer Panel Cooling Heat Storage", Refrigerating Engineering. July 1947.
Leopold, C. S., "The Mechanism of Heat Transfer Panel Cooling Heat Storage -Part II, Solar Radiation, Refrigerating Engineering. June, 1948.
Leopold, Charles S., "Design Factors in Panel and Air Cooling Systems," Heating. Piping and Air Conditioning. May, 1951.
> "Le Procede Calendal: de chauffage par rayonnement a' basse temperature", Reuve de 1'aluminium. Vol. -- No. 438, pg. 141-1565, Sebal, March 1975.
Lorenzi, R. J., and Schreiber, J. F., "Performance of an Electrical System of Panel Heating with Four Stages of Insulation," Heating. Piping. & Air Conditioning. January, 1949.
Macey, H. H., "Heat Loss Through a Solid Floor", Institute of Fuel Journal. 22-128, p. 369.
MacLeod, G. S., and Eves, C.E., "Baseboard Radiation Performance in Occupied Dwellings," Heating. Piping. & Air Conditioning. February 1950.
Madsen, T. L., "Definition and Measurement of Local Thermal Discomfort Parameters", ASHRAE Trans. V. 86, Pt. 1, 1980.
A-ll
Madsen, T. L. , "Thermal Comfort Measurements", ASHRAE Trans.. Vol. 82, Part I, 1976, pp. 67-70.
Madsen, T. L. , Olesen, B. W. , Kristensen, N. K., "Comparison Between Operative and Equivalent Temperature Under Typical Indoor Conditions", ASHRAE Trans.. V. 91, Pt. IB, 1985.
Mclntyre, D. A., "Eight Hour Floor Warming: A Feasibility Study", Electricity Council Research Centre, Capenhurst, England, March, 1977 (NTIS-PB 277 115).
Mclntyre, D. A., "Evaluation of Thermal Discomfort", Electricity Council Research Centre, Capenhurst, England, October, 1984, (NTIS-PB 85-189975).
Mclntyre, D. A., "Overhead Radiation and Comfort", Electricity Council Research Centre, Capenhurst, England, May, 1976, (NTIS-PB 277-428).
Mclntrye, D.A., "Overhead Radiation and Comfort", The Building Services Engineer. Vol. 44, Pg. 226-34, Batiste Publ. Ltd., January 1977.
Mclntyre, D. A., "Radiant Heat from Lights and Its Effect on Thermal Comfort", Illuminating Engineering Society. London. V. 8. No. 3. 1976.
Mclntyre, D. A., "Sensitivity and Discomfort Associated with Overhead Thermal Radiation", Ergonomics. V. 20, No. 3, pp. 287-296, 1977.
Mclntyre, D.A. , "The Thermal Radiation Field", Building Science. Vol. 9, 1974.
Mclntyre, D. A., "Thermal Radiation from Lighting Installations", The Building Services Engineer. V. 41, April, 1973.
Mclntyre, D. A., "Warm Air and Radiant Heating: Steady State Power Requirements", Electricity Council Research Centre, Capenhurst, England, Dec, 1980, (NTIS-PB83-231506).
Mclntrye, D. A. , "Warm Air Or Radiant Heating?" Building Research and Practice Vol. 17, No. 1, Pg. 48-51, Centre Scientifique et technique du bAatiment, January-February 1984.
Mclntyre, D. A. and Brailsford, J. R. , "The Efficiency of Radiant Heat Sources", The Building Services Engineer. Vol. 40, Feb., 1973.
McNall, Preston E., Jr., "A Manufacturer's View of Radiant Heater Control," ASHRAE Trans., Vol. 81, Pt. 1, 1973.
McNall, P. E., Jr. and Biddison, R. E., "Thermal and Comfort Sensations of Sedentary Persons Exposed to Asymmetric Radiant Fields", ASHRAE Trans.. V. 76, Pt. 1, 1970.
McNall, P. E. and Schlegel, J. C , "The Relative Effects of Convection and Radiation Heat Transfer on Thermal Comfort (Thermal Neutrality) for Sedentary and Active Human Subjects", ASHRAE Trans.. V. 74, Pt. 2, 1968.
A-12
Michaels, K. B. , Nevins, R. G. and Feyerherm, A. M. , "The Effect of Floor Surface Temperature on Comfort Part III, College Age Females", ASHRAE Trans.. V. 70, 1964.
Mills, Clarence A., "Residential Cooling by Reflective Radiation," Refrigerating Engineering. Vol. 58, No. 11, November, 1950.
Mills, C.A., "Reflective Radiant Conditioning Can Provide More Comfort at Less Cost", Refrigerating Engineering. Jan., 1955.
Mills, C.A. , "Sensible vs Latent Heat Removal in Radiant Cooling," Refrigerating Engineering. March 1958.
Mills, Clarence A., "Year-Round Residential Conditioning By Reflective Radiation," Refrigerating Engineering. November 1950.
Min, T.C., Schutrum, L.F., Parmelee, G.V., and Vouris, J.D., "Natural Convection and Radiation in a Panel-Heated Room," Heating. Piping & Air Conditioning. Cleveland, Ohio, May 1956.
Moisan, A. and Lebrun, J., "Comfort in Damp Cold Air with Radiant Spot Heating", Proceedings of, the Second Int. CIB/RILEM Symposium on Moisture Problems in Buildings. Rotterdam, 1974.
Morant, M.A. and Strengnart, M., "Simulation of a Hydronic Heating System; Radiator Modelling," Clima 200 - Heating. Ventilating and Air-Conditioning System. Vol. 6, W S Kongres - W S Meese, 1985.
Morelli, A.J., "An Economic Study of an Electric Infrared Space Heating Installation," ASHRAE Journal, June 1968.
Nakanishi, E. , Pereira, N. C , Fan, L. T. and Hwang, C. L. , "Simultaneous Control of Temperature and Humidity in a Confined Space - Part I, Building Science. Vol. 8, 1973, p. 39-49.
Naumov, A. L., Solovyov, A. J., Shilkrot, E.O., "Heating of Industrial Buildings with the Help of Suspended Radiating Panels," Clima 2000 -Heating. Ventilating and Air-Conditionign Systems. Vol. 6, WS. Kongres -W S Messe, 1985.
, National Fuel Gas Code. ANSI Z 223.1, 1984, American Gas Association, 1984.
NEMA, Infrared Application Manual. Stardard Publication No. HE 3-1983, National Electrical Manufacturers Association, Washington, DC, 1983.
Nevins, R. G., "Criteria for Thermal Comfort", Building Research. July-August, 1966.
Nevins, R. G. and Feyerherm, A. M. , "Effect of Floor Surface Temperature on Comfort Part IV: Cold Floors", ASHRAE Trans.. V. 73, Pt. 2, 1966.
Nevins, R. G. and Flinner, A. 0., "Effect of Heated-Floor Temperatures on Comfort", Heating Piping and Air Conditioning. Oct., 1957, p. 149-153.
A-13
Nevins, R. G. , Michaels, K. B. and Feyerherm, A. M. , "The Effect of Floor Surface Temperature on Comfort Part 1, College Age Males", ASHRAE Trans.. V. 70, 1964.
Nottage, H.B., Franks, C. V., Hulbert, L.E., Schutrum, L.F., "Heat Flow Analysis in Panel Heating or Cooling Sections," Heating, Piping and Air Conditioning, May 1953.
Olesen, B. W. , "A Simplified Calculation Method for Checking the Indoor Thermal Climate", ASHRAE Trans.. V. 89, Pt. 2B, 1983.
Olesen, B. W. , Mortensen, E. , Thorshauge, J. and Berg-Munch, B. , "Thermal Comfort in a Room Heated by Different Methods", ASHRAE Trans.. V. 86, Pt. 1, 1980.
Olesen, B. W. and Nielsen, R., "Radiant Spot Cooling of Hot Working Places", ASHRAE Trans.. V. 87, Pt. 1, 1981.
Olivieri, J. B., How to Design Heating-Cooling Comfort Ceilings. Business News Publishing Company, Birmingham, MI, 1971.
Olivieri, J. B. and Singh, T., "A Computer Program for Radiant Cooling of High Bay Buildings", ASHRAE Trans.. V. 93, Pt. 1, 1987.
Pam, R. L. and Kesselring, J. P., Burner Survey for a High Efficiency Gas-Fired Heating Unit. Alzeta Report No. 84-706-104, Teledyne Laars, January 1984.
Parmelee, G. V. and Huebscher, R. G., "Forced Convection Heat Transfer from Flat Surfaces", ASHRAE Trans.. Vol. 53, 1947.
Peach, J., "Radiators and other Convectors", J.I.H.V.E., Vol. 39, Feb., 1972.
Pedersen, C. D., Spitler, J.D., Bunkofske, R.J., Leverenz, D.J., "Experimental Study of Radiation and Convection Heat Exchange in Rooms for Energy Analysis Program Models," Clima 2000 - Building Design and Performance. Vol. 2, W S Kongres - W S Messe, 1985.
Perry, E. H., Cunningham, G. T. , and Scesa, S., "An Analysis of Heat Losses through Residential Floor Slabs," ASHRAE Trans., V. 91, Pt. 2, 1985.
Pfafflin, J.R., "Space Heating Dynamics", IEEE Trans, on Indust. Appls.. Vol. IA-19, No. 5, Sept/Oct, 1983.
Pierce, J.D., "Application of Fin Tube Radiation, "ASHRAE Journal, Feb. 1963.
Plattls, R. E., "Where Polyethylene Pipe Challenges Metal for slab Radiant Heating", Canadian Builder, April, 1963, Pg. 55.
Prince, Fred, J., "Causes and Prevention of Air Temperature Stratification", Plant Engineering. May 1982.
Prince, F. J., "Selection and Application of Overhead Gas-Fired Infrared Heating Devices," ASHRAE Journal, October 1962.
A-14
Prince, Fred J., "Infrared Heating for Overall Comfort," ASHRAE Journal, Dec. 1968.
Raber, B. F. and Hutchinson, F. W. , "Optimum Surface Distribution in Panel Heating and Cooling Systems", ASHVE Trans.. 1944.
Raber, B. F. and Hutchinson, F. W. , Panel Heating and Cooling Analysis. John Wiley & Sons, NY, 1947.
, "Radiant Cooling Panel will get Tryout in U.S.", Air Conditioning. Heating and Refrigeration News. October 21, 1985.
t Radiant Floor Heating. Plasco Manufacturing Ltd., Janca Enterprises Ltd., March 1985.
Rapp, George, "Analysis of Free convection and Radiation Heat Transfer in Valance Heat Exchangers," ASHRAE Trans., Vol. 72, Pt. 1, 1966.
Rapp, George and Gagge, A. P., "Configuration Factors and Comfort Design in Radiant Beam Heating of Man by High Temperature Infrared Sources," ASHRAE Trans., Vol. 73, Pt. 2, 1967.
Rickman, James D., "Selected Segment Hydronic Heating System," Energy Conversion Mgmt., Vol. 25, No. 1, Pg. 73-83, 1985.
Rohles, F. H., Jr., "Temperature or Temperament: A Psychologist Looks at Thermal Comfort", ASHRAE Trans.. V. 86, Pt. 1, 1980.
Ronge, Hans, E., and Lofstedt, Borje E., "Radiant Drafts from Cold Ceilings," Heating, Piping, & Air Conditioning, Uppsala, Sweden, September, 1957.
Roots, W.K., "Electric Space Heating with Active Boundary Members", Institution of Electrical Engineering Proc.. Vol. 114, No. 7, July, 1967.
Sanford, Len, "No Problem with Radiation" The Heating and Air Conditioning Journal. Vol. 54, No. 630, Pg. 14-18, 18, 21, Troup Publ., July-August 1984.
Sartain, E. L., and Harris, W. S., "Heat Flow Characteristics of Hot Water Floor Panels," Heating. Piping. & Air Conditioning. January 1954.
Sartain, E. L. and Harris, W. S., "Performance of Covered Hot Water Floor Panels - Part I - Thermal Characteristics", Heating. Piping. & Air Conditioning. October, 1955.
Sartain, E. L. , and Harris, W. S.f "Performance of Covered Hot Water Floor Panels - Part II - Room Conditions," Heating. Piping. & Air Conditioning. November, 1956.
Sauer, Harry J., "Configuration Factors for Radiant Energy Interchange with Triangular Areas," ASHRAE Trans., Vol. 80, Pt. 2, 1974.
Saunders, J. H. and Andrews, J. W., "The Effect of Wall Reflectivity on the Thermal Performance of Radiant Heating Panels", ASHRAE Trans.. V. 93, Pt. 1, 1987.
A-15
Schlegel, J. C and McNall, P. E., "The Effect of Asymmetric Radiation on the Thermal and Comfort Sensations of Sedentary Subjects", ASHRAE Trans.. V. 74, Pt. II, 1968.
Schlegel, J. C. and McNall, P.E., Jr., "The Effect of Asymmetric Radiation on the Thermal and Comfort Sensations of Sedentary Subjects", ASHRAE Trans.. Vol. 74, 1968.
Schutrum, L. F. and Humphreys, C. M., "Effects of Non-Uniformity and Furnishings on Panel Heating Performance," Heating. Piping and Air Conditioning. February, 1954.
Schutrum, L.F. and Humphreys, C M . , "Further Studies of the Thermal Characteristics of Plaster Panels," Heating. Piping and Air Conditioning. June 1953.
Schutrum, L. F. and Min, T. C , "Cold Wall Effects in a Ceiling - Panel-Heated Room," Heating. Piping. & Air Conditioning. Cleveland, Ohio, August, 1956.
Schutrum, L.F., and Min, T. C , "Lighting and Cooled Air Effects on Panel Cooling," Heating. Piping. & Air Conditioning. Cleveland, Ohio, November, 1957.
Schutrum, L.F., Parmelee, G. V., Humphreys, C M . , "Heat Exchanges in a Ceiling Panel Heated Room," Heating. Piping and Air Conditioning. December, 1952.
Schutrum, L.F., Parmelee, .G. V., Humphreys, CM., "Heat Exchanges in a Floor Panel Heated Room," Heating. Piping and Air Conditioning. July 1953.
Schutrum. L.F. and Vouris, J. D. , "Effects of Room Size and Non-Uniformity of Panel Temperature on Panel Performance," Heating. Piping and Air Conditioning. September, 1954.
Schutrum, L. F. , Vouris, John, and Min, TC , "Preliminary Studies of Heat Removal,By a Cooled Ceiling Panel," Heating. Piping & Air Conditioning. Cleveland, Ohio, January, 1955.
Shamsundar, N. , Lienhard, J. H. and Tezduyar, T. E., "Performance of Polybutylene Pipe in Concrete Heating Panels", Report No. 2, Department of Mechanical Engineering, University of Houston, Houston, Texas, 77004, 1985. (UNPUBLISHED)
Simmons, R. C , "Five Years Operation of an Industrial Infrared Heating System," ASHRAE Journal, June 1968.
Singh, T. , and Olivieri, J. B., "Effect of Radiant Cooling Panels on Temperature Stratification under RP-260", Final Report, ASHRAE, August, 1981.
Singh, T. and Olivieri, J. B., "Effect of Radiant Cooling Panels on Temperature Stratification" ASHRAE Trans.. V. 88, Pt. 2, 1982.
A-16
Smith, R. M. and Rae, A., "Patient Comfort and Radiant Ceiling Heating in a Hospital Ward", Building and Environment. V. 12, pp. 143-146, 1977.
Sofrata, H. M. and Al-Hukail, Y., "Spot Cooling System Design", ASHRAE Jnl.. Jan., 1987.
Sowell, E. F. and Walton, G. N. , Efficient Computation of Zone Loads," ASHRAE Trans., Vol. 86, Pt. 1, 1980.
Spangler, A. T., "Industrial Climate Control Versus Radiant Heat", Air Conditioning. Heating and Ventilating. Jan., 1965.
Sparrow, E. M. and Lin, S.L. , "Radiation Heat Transfer at a Surface Having Both Spacular and Diffuse Reflectance Components", Int. Jnl. of Heat & Mass Trans.. Vol. 8, 1965.
Spolek, G. A., Herriot, D. W. and Low, D. M., "Airflow in Rooms with Baseboard Heat: Flow Visualization Studies", ASHRAE Trans.. V. 92, Pt. 2, 1986.
Springer, W. E., Nevins, R. G.,, Feyerherm, A. M. and Michaels, K. B. , "The Effect of Floor Surface Temperature on Comfort: Part III, The Elderly", ASHRAE Trans.. V. 72, Pt. 1, 1966.
Stevens, J. C., Marks, L. E. and Gagge, A. P., "The Quantitative Assessment of Thermal Comfort", Environmental Research. Vol. 2, 1969, pp. 149-165.
Stevens, Joseph C , and Marks, Lawrence, "Subjective Warmth in Relation to the Density, Duration, and Areal Extent of Infrared Irradiation," ASHRAE Trans., Vol. 76, Pt. 1, 1970.
Subcommittee of TAC, "Thermal Design of Warm Water Concrete Floor Panels" ASHRAE Research Laboratory, Trans., ASHRAE, V. 63, 1957.
Subcommittee of TAC, Staff members of the ASHRAE Research Laboratory, "Thermal Design of Warm Water Ceiling Panels," Heating. Piping. & Air Conditioning. December, 1955.
t Thermo Lutz. Planungsunterlage Fur Ingenieure, ThermoLutz GMBH and Co. Heizungstechnik Kg.
Tasker, C., Humphreys, C. M., and Parmelee, G. V., "The ASHVE Environmental Laboratory", Heating. Piping and Air Conditioning. Vol. 24, March, 1952.
Taylor, F. M. H., "Radiant Space Heating", Building Materials. May/June, 1973.
Tenney, A. S., Ill, "Red Hot and Hotter - Industrial Radiation Thermometry", Mechanical Engineering. Oct. 1986.
Tredre, B. E., "Assessment of Mean Radiant Temperature in Indoor Environments", Brit., J. Industr. Med., V. 22, 58, 1965.
Trewin, R. R., Langdon, F. M., Nelson, R. M. and Pate, M. B. , "An Experimental Study of a Multipurpose Commercial Building with Three Different Heating Systems", ASHRAE Trans.. V. 93, Pt. 1, 1980.
A-17
Trewin, R., Pate M., and Nelson, R., "An Experimental Study of an Installed High Temperature Radiant Heater and Enclosure," ASHRAE Trans., Vol. 92, Pt. 1, No. 1, 1986.
, "Underfloor Radiant System Uses 86H Supply Water", Air Conditioning. Heating and Refrigeration News. Oct. 21, 1985.
VanGerpen, J. H. and Shapiro, H. N., "Analysis of Slab-Heated Buildings," ASHRAE Trans., V. 91, Pt. 2, 1985.
Walker, C.A., "Control of High Intensity Infrared Heating," ASHRAE Journal November, 1962.
Walton, George N., "A New Algorithm for Radiant Interchange in Room Loads Calculations", ASHRAE Trans., Vol. 86, Pt. 2, 1980.
Weida, D.E., P.E., "Life-Cycle Cost Analysis of Hydronic Radiant Panel," ASHRAE Trans., Vol. 92, Pt. 1, 1986.
Weigel, R. H. and Harris, W. S., "Heating a Basementless House with Radiant Baseboard," Trans. ASHRAE, V. 55, 1949.
Wilkes, G. B. and Peterson, C.M.F., "Radiation and Convection from Surfaces in Various Positions", ASHRAE Trans.. Vol. 44, 1938.
Zawacki, T. S., Huang, V., and Macriss, R. A., "Development of a Standard Test Method for Measurement of the Radiant Heat Output of Gas-Fired Infrared Heaters", ASHRAE Trans.. V. 93, Pt. 1,, 1987.
Zhang, Z. and Pate, M. B. , "An Experimental Study of the Transient Response of a Radiant Panel Ceiling and Enclosure", ASHRAE Trans.. V. 92, Pt. 2, 1986.
Zhang, Z. and Pate, M. B., "A Numerical Study of Heat Transfer in a Hydronic Radiant Ceiling Panel", Numerical Methods in Heat Transfer. HTD-Vol. 62, ASME, New York, 1986.
Zhang, Z., Liu, T., Pate, M. B., "An Experimental Study of a Residential Solar System Coupled to a Radiant Panel Ceiling", Heat Transfer and Fluid Flow in Solar Thermal Systems. SED-Vol. 1, ASME, New York, 1985.
Zhang, Z., Pate, M. B., and Nelson, R., "A Performance Evaluation of a Resi-dental Solar Hydronic Radiant Heating System", Proceeding of the 1986 ASME Solar Energy Conference. Anaheim, California, April 14-17, 1986.
A-18
APPENDIX B
ANNOTATED BIBLIOGRAPHY
" < ! • '
la
B-l
m
ANNOTATED BIBLIOGRAPHY
For all of the entries given in the BIBLIOGRAPHY, a short discussion
concerning each article has been prepared. In many instances, this is a
reproduction of the abstract or conclusions of the article. The categories
in this Annotated Bibliography are as follows.
A) Load Analysis and Modeling
B) Convection Coefficients
C) General
D) Comfort Conditions
E) Thermal Comfort - Radiant
F) Floor Panels
G) Panel Heating and Cooling
H) Infrared Heating
I) Design Procedures
J) Energy Consumption
K) Transient Effects
L) Instruments
M) Controls
N) Spot Heating and Cooling
It should also be noted that there are some secondary references available
from most of the entries in the Bibliography by examining the references for
that specific entry. All of these secondary references are not listed, but
are obtainable by locating the article given in the Bibliography and then
locating the references in that specific article.
B-2
ANNOTATED BIBLIOGRAPHY
A) Load Analysis and Modeling
1. ASHRAE, Procedure for Determining Heating and Cooling Loads for Computerizing Energy Calculations - Algorithms for Building Heat Transfer Subroutines. American Society of Heating, Refrigerating and Air Conditioning Engineers, Atlanta, GA., 1975.
This ASHRAE publication presents algorithms for calculating the heating and cooling loads for structures. It contains algorithms for calculating the following items: solar load on walls, roofs and glass, external shadow calculations, radiation shape factors, convection heat transfer coefficients, psychrometric property calculations, and other miscellaneous algorithms.
* * *
2. Ball, H. D. and Green; D., "The Impact of Lighting Fixtures on Heating and Cooling Loads - Mathematical Mode", ASHRAE Trans.. V. 89, pt. 2A, 1983.
A mathematical model describing the spatial and time distribution of energy produced by lighting has been constructed. The model is based on well-known heat-exchange computation procedures and contains no empirical coefficients from lighting energy-transfer experiments. A computer program based on this model was used to generate predictions of cooling loads created by lighting for a variety of building-lighting arrangements, airflow paths and rates, and lighting duration. Comparisons of model predictions are made with published experimental steady-state and transient results. The modeling procedures and the transient comparisons are described in detail. This model was used to generate values of room and total loads and to present these in a form or forms suitable for use in design.
* * *
3. Beier, Richard A. and Gorton, Robert L. , "Thermal Stratification in Factories - Cooling Loads and Temperature Profiles," ASHRAE Trans., Vol. 84, Pt. 1, 1978.
This study has produced a numerical solution to the process of transient thermal stratification occurring in factories. The analysis was restricted by considering the light fixtures as defining the interface between the conditioned area and the stratified area. The solution was based on a simplified, physical model with thermal plumes rising from the light fixtures in a stratified layer between the lights and the ceiling.
The results indicate that when varied over a reasonable range, ceiling height is of no importance in the system in that it has no influence on cooling load magnitude. This indicates that the technique of load reduction by stratification can work whenever an undisturbed strata can be formed. Also, exhausting needed ventilation for the cooled space through the roof can
B-3
reduce the cooling load. Comparison of results from the computer model to measured values from one
factory show acceptably close agreement. More extensive field comparisons need to be made and detailed laboratory investigations of some of the basic assumptions used in model development are required before the model can be applied with full confidence.
Example temperature profiles are given in the article.
* * *
4. Carroll, Joseph A., An "MRT Method" of Computing Radiant Energy Exchange in Rooms. Systems Simulation and Economic Analysis Conference Proceedings, January 1980, San Diego, CA.
The "MRT View Factor" method presented in this paper couples each surface in a room to an MRT node, which acts as a clearinghouse for all radiative exchanges. An upward adjustment in the coupling between each surface and the MRT exactly cancels that surface's self-weight in the MRT. The adjustments also happen to improve the accuracy of the conventional view factors implicitly assigned by MRT methods. The effects of surface emittance and air emittance (typically .05-.15 in residences) are modelled without difficulty. For greatest accuracy, radiation coefficients can be varied with temperature.
This method is inherently free from heat balance errors and errors in the overall radiative coupling of each surface to its environment. Errors do occur in the "implicit view factors", but errors such as this are inherent in any method which overlooks the gory details of the enclosure geometry. Coplanar surfaces cause the largest errors, and these errors can be compensated for if necessary.
* * *
5. Dunkle, R. V.,' "Configuration Factors for Radiant Heat Transfer Calculations Involving People", ASME Trans.. Jnl. of Heat Trans fer, Feb., 19 6 3.
It is thought that the point and area configuration factors for people reported in this paper will prove very useful in heat transfer problems involving people. While people of unusually stout or slim build may deviate somewhat from the empirical equations, they are thought adequate for nearly all engineering work. If more precise information is needed for a specific individual, recourse can be had to experimental measurements. Six configuration figures are presented for a "standard person".
* * *
6. Endreb, B. Von H., and Mommertz, W., "Untersuchung von Flachenheiz -Systemen Mit Der Thermoelektrischen Analogie", Elektrowaerme. Vol. 39, No. 4, pg. 203-210, August, 1981.
This is a German article describing the radiant energy balance in a room which is receiving solar radiation through the glass and on its exterior surface. It uses the electrical analogy for surface radiant exchanges.
B-4
7. Ford, J. K., "Deriving Radiation View Factors within a Triangular Cross-Section Residential Attic", ASHRAE, Trans., Vol. 89, pt. 1A, 1983.
A complete set of radiation view factors within a triangular, symmetric, cross-section attic space is derived. These view factors allow nodal radiant energy-transfer equations to be solved for radiant heat exchanges between the attic floor, roof sections, endwalls, and soffits. A simplified analysis of radiant and convective heat transfers within an idealized attic space shows radiative heat transfers to be on the same order of magnitude as convective heat transfers. A computer program is available to solve radiation view factor equations for some representative attic configurations. A simple set of nodal radiation heat transfer equations is also included to demonstrate the application of view factors to an attic floor and two attic roof sections radiating to one another.
* * *
8. Gebhart, Benjamin, "A New Method for Calculating Radiant Exchanges", Heating Piping. .& Air ConditioninE. Ithaca, New York, July 1958.
A new method of thermal radiation analysis for non-black surfaces is demonstrated by calculating various radiant energy transfer rates in a room. The analysis is direct and consistent because of the way in which basic quantities are defined and is similar to the approach used for black surfaces. The same basic technique is used in all calculations with only minor alterations.
In the calculations, as in any calculations which account for many processes simultaneously, the numerical operations are lengthy. The involved calculations arise in the solution of the equations for the absorption factors. For a many surface enclosure, this is necessarily a machine calculation.
Since the major inconvenience in the use of the method is in connection with the calculation of absorption factors, the question arises as to whether or not the values of the absorption factors could be presented graphically. This is the standard practice for the similar quantity, angle factor.
* * *
9. Gorton, R. L. and Leard, A. T., "A Computer Program for Air Temperature and Cooling Load Determination for Stratified-Cooled Industrial Buildings", ASHRAE Trans.. V. 90, Pt. IB, 1984.
A description of a computer program capable of computing time-varying cooling loads for a stratified-cooled space has been presented. The program is structured to accept hourly changes in internal loads and environmental conditions, to account for storage effects in the structure and in the space air, and to determine the resulting space air temperature profiles and space-cooling loads. Results from a number of situations of interest were presented.
B-5
10. Gorton, R. L. and Sassi, M. M., "Determination of Temperature Profiles and Loads in a Thermally Stratified, Air-Conditioned System: Part 1-Model Studies", ASHRAE Trans. . V. 88r Pt., Pt. 2, 1982.
Experiments have been run in a model chamber to provide information on temperature patterns that develop in stratified-cooled systems as a function of various combinations of geometric, flow, and thermal variables. The results are limited to conditions of uniform floor-level loads (no thermal plumes) with "perfect" air distribution (uniform diffusion of air in the cooled space and no interaction with the upper-zone stratification pattern).
The major conclusions reached from a qualitative evaluation of the test results are:
1. Loads originating at the lights reach the cooled space by radiation and by means of air circulation caused by free convection currents around the lights.
2. Loads originating at the roof reach the cooled space by radiation to the lower zone. Convection from the roof is effectively blocked by the stagnant, high-temperature air layer adjacent to the roof.
3. The space temperature profiles and cooling loads are relatively insensitive to the height of the lights and to their position relative to the cool zone and the roof.
4. The space temperature profiles and cooling loads are relatively insensitive to the height of the roof and to its height relative to tKe cool zone.
5. Exhaust of air from high in the structure in the stratified zone provides a temperature reduction there proportional to the exhaust flow rate. There is an accompanying reduction in lower-zone cooling load; however, the reduction is small and high-level exhaust cannot be considered as a significant means of load control.
6. The supply air temperature and flow rate do not influence the temperature patterns in the space. The level of temperature is changed, for a given load, by changes in temperature and flow rate but not the general shape of the temperature profile.
11. Gorton, R. L. and Sassi, M. M. , "Determination of Temperature Profiles and Loads in a Thermally Stratified, Air Conditioned System: Part 2 - Program Description and Comparison of Computed and Measured Results", ASHRAE Trans.. V. 88, Pt. 2, 1982.
A computer program for use in determining space air temperatures in a stratified-cooled scale model space is described. Program logic, equation formulation, and calculational models of the various heat transfer and flow processes are presented. Program results are compared to experimental results
B-6
obtained in the model chamber. These comparisons are presented for a range of geometric and load variables. The program is shown to be capable of producing acceptably accurate estimates of the measured temperatures, comparisons within 1HC being obtained in most cases. Many estimated air temperature profiles are given. These would be useful in load analysis calculations.
* * *
12. Harrington, Keith and Lydon, R. T., Comparison of Building Thermal Analysis "Methods. Systems Simulation and Economic Analysis Conference Proceedings, January 1980, San Diego, CA.
This paper has endeavored to explore the published state-of-the-art of passive solar load models. To that end it has discerned eight basic algorithmic types. Two of these are meant for calculators. The other six are implemented on large computers. All were considered in light of their completeness, efficiency, inputs and appropriateness of output to the various stages of building design. It has been found that all considered models make a trade-off between cost and model completeness, each to serve different ends.
* * *
13. Harris, Robert L., Jr., "Computer Simulation of Radiant Heat Load and Control Alternatives", Journal of American Industrial Hygiene Assoc. V. 35, Feb., 1974.
A digital computer program calculating the exchange of radiant heat between an absorber and a multiplicity of radiators in its surroundings is described. A matrix of absorber locations within a workspace having radiators in its surroundings may be specified, permitting the construction of isopleths of heat load in the workspace for any elevation of interest. The contribution of each individual radiator to the heat load on the absorber is obtained. Repeated computer runs with simple changes in input data may be used to simulate various control alternatives. The program permits mathematical assessment of radiant heat exchange problems for which manual calculations would be prohibitively time-consuming and costly.
* * *
14. Harrison, E., "Calculation of the Heat Requirements of Rooms", The Building Services Engineer. Vol. 43, pg. 19-23, Batiste Publications Ltd., May 1975.
This article describes two factors which can be used to modify the standard design heating load when radiant heating systems are used. One factor relates to the additional heat transfer through the wall due to radiant panels, and the other factor is related to the infiltration difference between air convection systems and radiant systems.
B-7
15. Hedgepeth, Lloyd M., and Sepsy, Charles, "A Thermodynamic Simulation of a Building Environmental Control System", ASHRAE Trans., Vol. 78, Ft. 2, 1972.
This is a discussion on how to simulate a building HVAC System. Very few details or algorithms are given, only discussions on what must be done. The building did have a radiant ceiling heating system, but no information is presented.
* * *
16. Hutchinson, F. W. , "Influence of Gaseous Radiation in Panel Heating", ASHRAE Trans.. Vol. 53, 1947.
A graphical solution is presented from which the equivalent coefficient for gaseous radiant exchange can be evaluated as a function of separating distance, vapor pressure (expressed in terms of room air temperature and of relative humidity) and surface temperatures. Equations are developed to permit evaluation of radiant exchange between certain systems of surfaces when they are reflective and are separated by an absorbing medium.
Gaseous radiation does not appreciably affect either the panel size or panel rating for an ordinary panel heating system, but it does reduce the effectiveness of local (direct transfer) panels by about 10 percent. Because of cumulative absorption as associated with multiple reflections, gaseous radiation is. responsible for reducing the effectiveness of reflective surfaces (when used as room surfacing in an attempt to reduce radiant body heat loss) to a negligible value.
* * *
17. Khudenko, A. A., "Radiation Characteris t ics of Gas Infrared" Journal of Engineering Physics. Vol. 29, No. 5, Pg. 1395-1400, Consultants Bureau.
Results of an experimental study of the radia t ion charac ter i s t ics of commercial models of gas infrared heaters are presented. An analytic expression i s obtained for the d i s t r ibu t ion of irradiance over a f l a t object a t various distances from the heater . This expression i s :
E(w/m2) - [E0(w/m2)/h2] exp[-b x c ]
and values for EQf b and c are given in the paper.
* * *
18. Korsgaard, V., "Necessity of Using a Directional Mean Radiant Temperature to Describe the Thermal Conditions in Rooms", Heating. Piping and Air ConditioninE. June, 1949.
The conclusions reached by this author are:
B-8
1. The directional mean radiant temperature (DMRT) of a room with respect to a surface element, as explained and shown in the equations given, is not a single physical constant for a given room under given thermal conditions, but is a physical quantity that usually varies both with the place and the direction within a room.
2. When running comparative tests on different heating systems --especially when they include such types of heating systems where the thermal conditions in the room are regulated mainly by varying the temperature of larger or smaller parts of the inside surfaces of the room -- it is important to pay proper attention to the DMRT as a design factor.
3. In comfort analyses it is important to consider the directional mean radiant temperature as a variable, for at any single point on a person's body the heat transfer conditions represent the combined influences of radiation exchange with those surroundings exposed to that part of the body, of convection, and of evaporation.
* * *
19. Mclntyre, D. A., "Warm Air and Radiant Heating: Steady State Power Requirements", Electricity Council Research Centre, Capenhurst, England, Dec, 1980, (NTIS - PB83 - 231506).
The effectiveness of radiant and warm air heating systems were studied theoretically. Power requirements and heat loss estimates were determined using a simplified procedure together with a more detailed computer program, both of which showed good agreement. It was found that in a well insulated domestic room there was little to choose between the two types of heating systems. Radiant heating was advantageous when only local warming of part of the room was required; however, the possibility of savings would have to be set against high radiant asymmetries and an uneven temperature distribution over the room. Radiant heating was shown to be more economical than warm air heating in large spaces with high ventilation rates, such as in factories. The efficiency of the warm air system was reduced by temperature stratification if the inlet temperature was high and the warm air was inadequately mixed, causing an increased convective transfer coefficient.
* * *
20. Mclntyre, D. A., "Warm Air or Radiant Heating?" Building Research and Practice, Vol. 17, No. 1, Pg. 48-50, Centre Scientifique et technique du bAatiment, January-February 1984.
In comparing warm air and radiant heating systems, the author comes to the following conclusions:
(1) The simplified calculation method in the CIBS Guide and a more detailed computer program show very good agreement.
(2) Theoretical studies show little difference in the power required to maintain comfortable conditions in a domestic size room with either
B-9
radiant or warm air heating.
(3) Radiant heating can show advantages when only local warming of part of the room is required; the possibility of substantial savings must be set against the increased discomfort due to high levels of radiant asymmetry.
(4) Radiant heating is more economical than warm air heating in large spaces with high ventilation rates.
(5) The performance of the warm air system is reduced by temperature stratification if the inlet air temperature is high and the warm air is inadequately mixed. The increased transfer coefficient found with forced warm air systems will depress their heating performance.
* * *
21. Mclntyre, D. A. and Brailsford, J. R., "The Efficiency of Radiant Heat Sources", The Building Services Engineer. Vol. 40, Feb., 1973.
The radiant efficiency of horizontal cylinders, vertical and horizontal plates is presented graphically as a function of temperature and dimensions. Measured radiant efficiencies of a selection of commercial heaters are given. Efficiencies of several types of heaters, described as a radiant, fall below 50 percent. Higher efficiencies are achieved by conventional electric bar fires (78 percent) and by downward facing panel heaters (up to 80 percent).
* * *
22. Prince,-Fred, J., "Causes and Prevention of Air Temperature Stratification", Plant Engineering. May 1982.
The author performs an analysis of the effect of air infiltration on temperature stratification. He concludes the following:
If present conditions or an analysis of a new heating system design indicates stratification, the following, actions can be taken to reduce substantially or eliminate the problem:
i) Heating system delivery to the floor should be at least 45 percent of the input.
ii) Heaters must be kept as far as possible from the ceiling.
iii) Air must be introduced in the upper portion of the plant or be tempered; in both cases, air pressure should be sufficient to counteract the forces of infiltration.
iv) Systems with high infrared output to the floor, combined with a pressurized air supply (tempered if necessary), offer an economical and practical approach to the elimination of stratification. Warm-air systems that approach elimination of
B-10
s trat i f i cat ion require substantial expenditure of energy to keep the air continually circulated and homogenized.
* * *
23. Raber, B. F. and Hutchinson, F. W., "Optimum Surface Distribution in Panel Heating and Cooling Systems", ASHVE Trans.. 1944.
Basic experimental data has been presented to permit direct determination of fraction of energy received by a seated or standing subject from a wall or ceiling panel of elementary area. From those data information is also obtainable as to the fraction of energy received by a prone subject from an elementary ceiling panel. All experimental determinations are in groups of three based on the position of the subject with respect to the source both as to vertical and horizontal components and for full face, semi-profile and profile attitudes.
Use of the basic data to compute the absolute shape factor of a subject in any position with respect to a panel of any shape and size is described and illustrated; for panels of usual size in rooms having a ceiling height between 8 ft. and 12 ft. the order of magnitude of the shape factor is 1-2 percent; that is, less than two percent of the energy radiated from a conventional panel passes directly to the occupant.
Generalized patterns for the center lines of ceiling panels have been determined for the standing subject in semi-profile position. The recommended patterns are:
1. For 8 ft. ceiling height use a square pattern with center lines spaced at 10 ft. for best results; wherever possible spacings outside of the range from 9 ft. to 14 ft. should be avoided; the use of a single block panel in the center of the room is not recommended.
2. For 10 ft. ceiling height a square pattern similar to that described above is recommended, but with 14 ft. spacing as the optimum and 10 ft. to 15 ft. as the limiting range.
3. For 12 ft. ceiling a square pattern with 20 ft. optimum spacing is recommended (range 17 ft. to 21 ft.). For rooms smaller than 14 ft. x 14 ft. a pattern consisting of centrally located crossing panels running along both axes of the room is satisfactory. Unlike rooms having 8 ft. or 10 ft. ceilings, a 12 ft. ceiling room does have satisfactory uniformity of direct radiant exchange when the entire ceiling is used as a panel.
Limitations of the generalized patterns are discussed and the suggestion offered that the proposed standard panel distributions be used as an initial design arrangement from which to depart in seeking greater simplification, adjustment for localized exposure, or corrections for such diverse effects as inter-reflections, non-diffuse emission and the many other factors which include use of any completely generalized solution for optimum design of all installation.
B-ll
24. Rapp, George and Gagge, A. P., "Configuration Factors and Comfort Design in Radiant Beam Heating of Man by High Temperature Infrared Sources", ASHRAE Trans., Vol. 73, Pt. 2, 1967.
The conclusions of these authors are as follows:
1) The theory of diffuse gray body radiation exchange, when modified to include a new parameter called the Beam Utilization Vector, or ratio of mean directional to mean diffuse radiant intensity in the solid angle connecting a high intensity plane-point radiating source with an absorptive surface, provides a rational basis for calculating the incident radiant flux density and energy absorbed from a directional source by a man standing or sitting in any source-man geometry.
2. Mean body absorptance values for man irradiated by I-R sources having color temperatures of 2500 K (T-3 quartz lamps) and 1200 K(atmospheric gas-fired burners), as determined by reported reflectance measurements on white skin and clothing, may be taken, for purposes of design, at am =• 0.8 and am - 0.9, respectively, when clothed, and at 0.65 and 0.95, respectively, when wearing bathing attire.
3. A method for calculating the effect of heater energy re-radiated from the floor is presented, and, in the specific case cited, is found to amount to 25% of the direct irradiation of the occupants.
* * *
25. Sauer, Harry J., "Configuration Factors for Radiant Energy Interchange with Triangular Areas", ASHRAE Trans., Vol. 80, Pt. 2, 1974.
Graphs of configuration factors for triangular areas have been developed and are presented. These can be used in certain spaces where it is convenient to partition room surfaces into triangular patterns. Various types of checks were applied to the calculation in order to verify their validity.
* * *
26. Saunders, J. H. and Andrews, J. W. , "The Effect of Wall Reflectivity on the Thermal Performance of Radiant Heating Panels", ASHRAE Trans.. V. 93, Pt. 1, 1987.
The abstract for this work is as follows:
"Numerical results are presented for the heat loss from a room, heated by radiant panels, as a function of wall reflectivity. For these calculations, a computer model was developed for the transient heat transfer within an arbitrary closed room with radiant and convective heat exchange between the panel and room surfaces, and conductive heat transfer through the walls. In the model, the room air temperature was initially set back, while the mean radiant temperature was adjusted with the radiant panels to keep the operative temperature, or comfort level, constant. The results suggest that walls reflective in the infrared can reduce the steady state heat loss to the ambient by
B-12
reducing the wall temperatures, yet the walls radiate at higher equivalent blackbody temperatures since radiant heat from the panels is reflected by the walls. Future work may be needed to develop inexpensive wall paints or papers that are reflective in the IR, but appear normal in the visible spectrum."
They also stated that:
"Preliminary calculations showed that IR-reflective walls have the potential for significant reductions of steady-state heat loss. Experiments should be undertaken to determine the actual savings in a test room and an occupied home. The success of this scheme, however, will likely depend upon user acceptance, both in terms of cost, comfort and aesthetic appeal of the wallpaper or paints used. These wall coverings should be selective surfaces, having a low emissivity in the infrared and high emissivity in the visible spectrum (0.35 to 0.75 microns) to give a normal appearance. Some paints and wall papers are available with these properties, and more might be developed using information gained from research on solar selective surfaces, since requirements are similar to the requirements for selective surfaces needed for solar collectors (Agnihotri and Gupat 1981."
-- • * * *
27. Sowell, E. F. and Walton, G. N., "Efficient computation of Zone Loads", ASHRAE Trans., Vol. 86, Pt. 1, 1980.
This paper .addresses the problem of calculation of building space sensible cooling or heating load from sensible cooling or heating load from sensible heat gain or loss in an efficient manner on a digital computer. Two widely used methods are considered, referred to as the heat balance method and the ASHRAE weighting factor method. It is shown that by careful selection of a solution algorithm for the heat balance equations, the former method can be made much more efficient than suggested by early computer codes which used it. Moreover, it is shown that over-all computation times for this method thereby achieved are, for many important cases, no more than 10-15% greater than using the weighting factor approach. Since there is a loss of flexibility and accuracy with weighting factors, the heat balance may be the preferred approach in the future energy analysis computer programs.
* * *
28. Sparrow, E. M. and Lin, S. L., "Radiation Heat Transfer at a Surface Having Both Specular and Diffuse Reflectance Components", Int. Jnl. of Heat & Mass Trans.. Vol. 8, 1965.
A method of analysis has been devised for determining the radiant interchange among surfaces, each of which may have both specular and diffuse reflectance components. The formulation uses and generalizes the exchange factor concept (which was initially devised for specularly-reflecting surfaces) and the radiosity concept (which was initially devised for diffusely-reflecting surfaces). Various forms of the analytical method are presented that are suitable either for overall engineering-type computations or for more detailed local investigations. Specific analytical and numerical consideration is given to radiant interchange in cylindrical and conical
B-13
cavities and to radiant transport through a circular tube. Results are presented for various subdivisions of the surface reflectance into specular and diffuse components. In general, it is found that the radiant efflux from a cavity increases as the specular component becomes a larger fraction of the surface reflectance. A similar statement applies for the transmission of radiant energy through a tube.
* * *
29. Spolek, G. A., Herriot, D. W. and Low, D. M., "Airflow in Rooms with Baseboard Heat: Flow Visualization Studies", ASHRAE Trans.. V. 92, Pt. 2, 1986.
A model study has been performed to visualize the airflow that occurs in a room with baseboard heat. The full field flow pattern, the typical velocities, and temperature distributions in the room are reported for three different test cases. From the results of those tests, the following conclusions were drawn:
1) The overall flow circulation in a room reverses direction between heater "on" and heater "off" cycles. In both cases, a single circulation loop forms.
2. The basic flow pattern demonstrates relatively high air velocities near the walls, ceiling, and floor with stagnant air in the room's center. Temperature stratification can be significant.
3. During extremely cold outdoor conditions or during intermittent heater operation, the airflow pattern changes and includes two circulation loops. The warmest air passes from the heater directly to the center of the room to enhance mixing and reduce temperature stratification.
* * *
30. Walton, George N., "A New Algorithm for Radiant Interchange in Room Loads Calculations", ASHRAE Trans., vol. 86, Pt. 2, 1980.
In this paper, it is shown that the radiation interchange algorithms of the NBSLD and BLAST loads programs can be significantly improved. It is proposed that the radiant interchange in a room can be adequately modeled by assuming that each surface radiated to a fictitious surface which has an area, emissivity, and temperature giving about the same heat transfer from the surface as in the real multi-surface case. This approximation leads to a more accurate heat balance and can be used for large numbers of surfaces without greatly increasing computation time. The algorithm will make it computationally practical to account for previously neglected effects such as nonlinear, nonconstant interior convection coefficients and heat conduction between simultaneously simulated rooms. It would also be possible to do calculations of room loads and temperatures as they are affected by the operation of the air distribution system.
B-14
B. Convection Coefficients
1. Altmayer, E. F., Gadgll, A. J., Bauman, F. S., and Kammerud, R. C , "Correlations for Convective Heat Transfer from Room Surfaces", ASHRAE Trans., Vol. 89, pt. 2A, 1983.
Correlation of the rate of heat transfer from room surfaces to the enclosed air, based on empirical and analytical examinations of convection in two-dimensional enclosures, have been developed. The heat transfer data base from which the correlations were derived was generated by a validated numerical-analysis computer program. The correlations extracted from this data base express the heat transfer rate in terms of boundary conditions relating to room geometry and surface temperatures. The correlations are applicable to a class of room configurations with cold and warm surfaces on opposite vertical walls. The authors compare their correlations for convection coefficients with the ASHRAE coefficients for vertical and horizontal surfaces. There are substantial differences in the heat flows from the various methods. A complicated correlation description is given, which is a possible method for general calculations. The ASHRAE coefficients can give significant errors.
•* -k *
2. Bauman, F. , Gadgil, A., Kammarud, R., Altmayer, E., Nansteel, M. , "Convective Heat Transfer in Buildings: Recent Research Results" ASHRAE Trans, Vol. 89, pt. 1A, 1983.
Recent experimental and numerical studies of convective heat transfer in buildings are described, and important results are presented. The experimental work has been performed on small-scale water-filled enclosures; the numerical analysis results have been produced by a computer program based on a finite-difference scheme. The convective processes investigated in this research are (1) natural convective heat transfer between room surfaces and the adjacent air, (2) natural convective heat transfer between adjacent rooms through a doorway or other openings, and (3) forced convection between the building and its external environment (such as wind-driven ventilation through windows, doors, and other openings).
Results obtained at Lawrence Berkeley Laboratory (LBL) for surface convection coefficients are compared with existing ASHRAE correlations, and differences of as much as 50% are observed. It is shown that such differences can have a significant impact on the accuracy of building energy analysis computer simulations. Interzone coupling correlations obtained from experimental work reported in this paper are in reasonable agreement with recently published experimental results and with earlier published work. Numerical simulations of wind-driven natural ventilation are presented. They exhibit good qualitative agreement with published wind-tunnel data. Finally, future research needs are suggested.
The convection coefficients presently recommended by ASHRAE are internally inconsistent and in disagreement with recent research results. In particular, the transition to turbulence for convection in enclosures occurs at a Rayleigh number about one order of magnitude larger than is generally accepted. This means that a laminar flow correlation is applicable to a much wider range of Rayleigh numbers than previously recognized. More accurate correlations for
B-15
convection coefficients are needed because they have a significant impact on predictions of building energy consumption.
* * *
3. Co l l i co t t , H. E., Fontaine, W. E., and Grosh, R. J . , "Free Convection and Radiation Heat Transfer from Cylindrical Fins", ASHRAE Journal , Dec. 1965,
This inves t iga t ion i s concerned with a cyl indrical f in protruding from a constant temperature source into a constant temperature environment. Some equations are given.
* * *
4. Collicott, H. E., Fontaine, W. E., and Witzell, 0. W., "Radiation and Free Convection Heat Transfer from Wire and Tube Heat Exchangers", ASHRAE Journal, Dec. 1963.
Effective radiation configuration factors and free convection heat transfer characteristics were determined for wire and tube heat exchangers, and presented graphically as a function of the diameter to spacing ratios of the wires and the tube and GrPr. Values of the configuration factor or heat transfer coefficient may be determined for parameter values not shown by the construction of a simple cross-plot or interpolation.
* * *.
5. Danter, E. , "Heat Exchanges in a Room and the Definition of Room Temperature", The Building Services Engineer. Vol. 41, Feb., 1974.
The main purpose of this paper is to give an account of the considerations underlying the 'environmental temperature' introduced in the 1970 edition of the IHVE Guide. It originates in a re-examination of the problem of approximating to the effect on the thermal response of a room of the radiation heat exchanges within the room. Conventionally these are dealt with interns of a combined radiation-convection transfer between room air and surfaces but the underlying assumption is an over-simplification. A systematic analysis of the problem shows that the most consistent approximations are obtained in a revised representation of the convective and radiation exchanges in which the heat transfer within the room is related to an index temperature defined as a weighted average of air and room surface temperature. Appropriate weights are one-third and two-thirds respectively. The index temperature so defined is the 'environmental temperature' of the Guide. The analysis is used to examine the effect of re-interpreting the room temperature in the conventional approach as the arithmetic mean of air and mean surface temperature. This approach shows a less consistent performance and although it can in many cases provide a good approximation it cannot be generally relied upon to do so. The final section considers the overall steady-state heat balance on a room and presents the heat balance equation in a form which provides the framework for developing the analysis of non-steady conditions.
B-16
6. Eno, Burton E., "Combined Convection and Radiation from Rectilinear Fins", ASHRAE Trans., Vol. 73, Pt. 1, 1967.
The author presents an analysis of the combined effect of convection and radiation from rectilinear fins on tubes. It does not consider some of the practical problems of finned tubes, such as fin-tube bond and variability of the convection coefficient on the fin and tube. It could be used as a first approximation for finned radiation tubes.
* * •*
7. Hannay, J., Liebecq, G., Nusgens, P., "Convective Heat Transfers within a Large Open Plan Office Area: Experimental Results for Dynamic Buildings Simulations", Clima 2000 - Building Design and Performance. Vol. 2, W S Kongres - W S Messe, 1985.
Air movements and indoor temperature measurements in a large open-plan office area have shown quite definite convective couplings between a warm central core and cooler peripheral zones in both natural and HVAC conditions. Heat exchanges were estimated and parametric studies performed by computer simulations. The main conclusion drawn from the present work emphasizes the need of a multizone approach of such types of dwellings by computational means.
* * *
8. IHVE, "Thermal and other Properties of Building Structures", A3, IHVE Guide, 1977.
This section of this guide presents various values for convection coefficients for inside surfaces for heat loss/gain calculations. They are numbers which are very close to the values given in the ASHRAE Handbook of Fundamentals. Some of those given are listed below.
Rc, walls --- 1.43 hr ft2oF/Btu
Rc, upward flow to ceilings -- 1.0 hr^oF/Btu Rc, downward flow to floors -- 2.9 hr ft^oF/Btu
The following table is presented for inside surface resistances.
Building Element
Walls
Ceilings or roofs, flat or pitched roofs, floors
Ceilings and floors
Heat Flow
Horizontal
Upward
Downward
Surface Resistance, hr ft^oF/Btu
High Emissivity Surface (E=0.9)
0.7
0.60
0.85
Low Emissivity Surface (E<=0.05)
1.72
1.24
3.16
B-17
9. Min, T. C , Schutrum, L. F., Parmelee, G. V., and Vouris, J. D., "Natural Convection and Radiation in a Panel-Heated Room", Heating. Piping & Air Conditioning. Cleveland, Ohio, May 1956.
The data reported in this paper were obtained with the entire floor area or ceiling area used as a heated panel, a uniform environment wherein all surfaces other than the heated panel were at a uniform temperature, relatively still air conditions (no infiltration), and an empty, unlighted room. Under these conditions the following conclusions may be drawn:
1. On the basis of research done at the ASHRAE Laboratory, the following equations apply in calculating natural convection heat transfer at room surfaces.
A. In a floor-heated space
1. Convection from floor: qc = O ^ ^ t ) 1 - 3 1 / ! ^ 0 * 0 8
2. Convection to walls: qe = 0.29 (At)1-32/**0-05
3. Convection to ceiling: Same as convection from floor.
B. In a ceiling-heated space
1. Convection from ceiling: qe - 0.041 (At)^.25/De
2. Convection to walls: Same as for floor-heated space.
3. Convection to floor: Same as convection from ceiling.
2. Natural convection data given by other investigators for small heated plane surfaces were found to be in good agreement with all of the equations listed in the first conclusion except the equation for natural convection from a heated ceiling. The convection coefficient for small free-edge plates may be 6 to 10 times as great as that for a heated ceiling.
3. Room size in a ceiling-panel-heated room has a significant effect on the unit convection from the ceiling and the floor. However, the convection heat transfer from a heated ceiling is small in comparison with radiation exchange, and therefore, the effect of room size on the total heat transfer is riot important. In a floor-heated room, the effect of room size is not significant.
4. In a completely enclosed space with high emissivity room surfaces, the interchange factor between the heated panel and its enclosure may be approximated by the hemispherical emissivity of the panel surface.
5. The approximate combined film conductances (natural convection and radiation) for the heated panels based on the difference between panel temperature and the room air temperature for a normal size room with high emissivity surfaces are as follows:
h r c = 2 for heat floor panel at about 85 F.
B-18
h r c — 1.1 for heated ceiling panel at about 120 F.
10. Parmelee, G. V. and Huebscher, R. G.f "Forced Convection Heat Transfer from Flat Surfaces", ASHRAE Trans.. Vol. 53, 1947.
Conclusions:
1. The effects of surface length and air velocity on forced convection heat transfer between a smooth flat plate and a parallel stream of air have been measured for Reynolds numbers between 19,000 and 1,200,000.
2. The data for the completely turbulent boundary layer flow were in substantial agreement with a formula based on skin friction measurements of flat plates and were in reasonably close agreement with similar data to be found in the literature.
3. The data for the laminar boundary layer were also placed in agreement with the analogous skin friction formula, after corrections were made for the observed effects of natural convection.
4. Most satisfactory agreement with the friction curves was obtained by defining the length as the length of the heat transfer surface. However, it has been pointed out in the literature that the effect of unheated surface preceeding the heat transfer surface can be significant.
5. The air stream turbulence in the wind tunnel was measured and found to be 1.5 percent. Since turbulence is an important factor in heat transfer, it is suggested that future tests include a measurement of the general turbulence by an accepted method.
6. Examples of the application of the results to problems in the field of heating and air conditioning have been given, based on the assumption that the turbulent boundary layer flow is a fair approximation of conditions to be found in practice.
7. Further study is indicated in such matters as air stream turbulence, unheated surface preceding the heat transfer area, and natural convection effects at low velocities.
• » 11. Pedersen, C. D., Spitler, J. D., Bunkofske, R. J., Leverenz, D. J.
"Experimental Study of Radiation and Convection Heat Exchange in Rooms for Energy Analysis Program Models", Clima 2000 - Building Design and Performance. Vol. 2, W S Kongres - W S Messe, 1985.
The experiments performed so far have produced several interesting results.
A geometric dependence of the local film coefficient has been identified. The dependence is very strong in the vertical direction, and, while not as
B-19
strong, there does seem to be a dependence in the horizontal direction. The range of convective film coefficients produced by the various
experiments includes the standard ASHRAE coefficient (after subtracting the radiation component) . While it is unlikely that the standard ASHRAE coefficient is correct for all building heat transfer conditions, its use in combination with a well-mixed model seems to be reasonable for the conditions covered in these experiments.
An alternative model fo room heat transfer has been considered. Although the question of how an effectiveness model would be implemented remains unresolved, it was shown that, for a given geometry, the effectiveness parameter was highly correlated to mass flow and temperature difference.
* * *
12. Wilkes, G. B. and Peterson, G. M. F., "Radiation and Convection from Surfaces in Various Positions", ASHRAE Trans.. Vol. 44, 1938.
Results are given for radiation and convection coefficients for practical configurations found in HVAC. Horizontal and vertical orientation is considered, as is the surface emissivity. This appears to be the source of convection and radiation coefficients, which are published in the ASHRAE Handbook of Fundamentals.
B-20
C. General
1. Bahnfleth, Donald R., "Symposium: Field Performance of Infrared Heating Systems", ASHRAE Journal, June 1968.
This is a short article describing the activities of the ASHRAE Task Group on Radiant Space Heating and an ASHRAE Symposium that was held in 1968.
* * *
2. Bak, Richard, "Hydronic Radiant Floor Heating Staging a Comeback", Air Conditioning. Heating and Refrigeration News, March 1985.
Discussion of hydronic radiant floor heating systems. Presents the idea of using plastic piping in place of metal pipes. A general discussion of the installation, application and disadvantages is given.
* * *
3. Banhidi, L. , "The Thermotechnical Dimensioning of Radiant Strips and Borders for the Heating of Communal Buildings", Building Science. Vol. 9, 1974.
This paper describes a special form of radiant strip heating which offers a suitable solution for the supply of housing and communal buildings with hot water and radiant heating, taking into consideration also the comfort criteria. The thermotechnical calculations are based on Kollmar's method, the design parameters were determined by measurements.
* * *
4. Blossom, J. S., "High Temperature Water Heats New School", Heating Piping and Air Conditioning. March, 1959.
Discussion of a specific application of a split system in a school. It delivers radiant heat from the pipes that transport high temperature water to fan ventilator units. It looks at the details of the specific design for a school.
* * *
5. Blossom, J. S., "Pipe HTW through Classrooms", Heating. Piping and Air Conditioning. July, 1964.
This article discusses the use of high temperature water pipes through rooms in order to improve comfort. It is the radiant portion of split type of system. It is the discussion of an application for a specific building.
B-21
6. Cambel, A. B. , "Model Study of Radiant Heating", Ph.D. Dissertation, University of Iowa, 1950.
This is a thesis done in 1950 and a copy is not available from Dissertation Abstracts. Apparently a copy could be obtained from the library at the University of Iowa at a rather high cost.
* * *
7. Carroll, J. R. , "Radiant Systems for Heating", New Methods of Heating Buildings. BRI-760, National Academy of Sciences - National Research Council, Washington, D. C , 1960.
This is a discussion of the general characteristics of radiant heating systems. A table of emissivity values for building surfaces is given. General descriptions of radiant exchange; low, medium, and high temperature systems; thermal comfort aspects, and spot heating are presented.
* * *
8. , "Concrete Code Restricts Panel Heating", Heating. Piping, & Air Conditioning. October 1951.
A discussion concerning the possibility of the American Concrete Institute changing their code for structural concrete." such that it would be possible to have embedded pipes for radiant heating systems.
* * *
9. Faust, Frank H., "New Electric Heating Systems", New Methods of Heating Buildings. BRI-760, National Academy of Sciences- -National Research Council, Washington, DC, 1960.
This article discusses electric heating with the greatest emphasis being on the elctric heat pump. It discusses ceiling panels and other electric panels used for radiant heating. The author presents a short discussion on specific benefits, limitations, applications, costs, and construction requirements.
* * *
10. Holden, T. S., "Calculation of Incident Low Temperature Radiation", ASHRAE Journal, April 1961.
The author discusses methods for calculating low temperature radiation to building surfaces from the sky, the ground and from neighboring buildings. Emphasis is placed on a method developed for the case of a building surface at any angle to the horizontal with the ground at any slope.
B-22
11. Hough ten, F. C , et. al., "Heat Loss through Basement Floors and Walls", ASHVE Trans.. Vol. 48, 1942.
Experimental results for basement heat losses are presented. Heat flows, as a function of the time of year, are given as are the ground temperatures at various levels. They observed heat loss reduction as the ground temperature increased, and reduced heat loss from the center of the basement floor.
* * *
12. HPAC Engineering Data File, "Design and Control of High Temperature Hot Water Heat Consumers", Heating. Piping and Air Conditioning:. June, 1960.
This article presents design guidelines for the use of high temperature hot water systems. Part of the use can be by direct radiation to spaces. General design guidelines are given for heat exchanger selection, boiler selection and operation, and control systems.
* * *
13. Johnson, T. E. , "Radiation Cooling of Structures with Infrared Transparent Wind Screens", Proceedings of the Second Workshop on the Use of Solar Energy for the Cooling of Buildings, UCLA, August, 1975.
Energy conserving radiation cooling schemes for dwellings in high humidity climates have usually failed due to the deleterious effect of the wind. In this paper, the cooling mechanisms at work in wind conditions are examined. A radiator system using an infra-red transparent wind screen that doubles as the structural envelope is proposed and supporting experimental results are presented. A one family dwelling built with these radiation panels can carry 50 percent of the 24 hour cooling load. Worst case conditions give radiator coefficients of performance twice that of exisiting appliances.
* * *
14. Kweller Esher, "Criteria for Mechanical Energy Saving Retrofit Options for Single Family Residences", New and Existing Single Family Residences. ACEEE 1984 Summer Study on Energy Efficiency in Buildings.
This paper estimates energy savings, and provides performance and selection criteria, for mechanical equipment options for single-family homes; all from prior studies reported in the literature. Performance and selection criteria are presented as advantages, disadvantages and limitations for each option.
Four broad categories of energy-saving mechanical options were investigated: space heating, water heating retrofit options, heat pump water heaters, and recovery of central air conditioner waste heat by desuperheaters. Gas- and oil-fueled forced-air furnaces and hydronic (hot water) space-heating
B-23
equipment were treated in this report.
* * *
15. Morant, M. A. and Strengnart, M., "Simulation of a Hydronic Heating System; Radiator Modelling", Clima 200 - Heating. Ventilating and Air-Conditioning System. Vol. 6, W S Kongres - W S Meese, 1985.
The number of elements taken into account for the modelization of the radiator is chosen depending on the interest of the heating system simulation. If we are interested in calculating the integration of the energy delivered by the radiators over a period longer than one hour, a model with only one capacitance is quite good. One should keep in mind that the temperature evolution obtained with a logarithmic At is better than that observed with an arithmetic At, particularly if the emission is calculated for small water flow rates. Moreover, an optimization of the control system is better performed if the radiator is divided in several elements.
0
* * *
16. MacLeod, G. S., and Eves, C. E. , "Baseboard Radiation Performance in Occupied Dwellings", Heating. Piping. & Air Conditioning. February 1950.
In order to determine the performance of baseboard radiation in field installations, tests were conducted in several houses of diversified construction. Winter comfort conditions obtained were evaluated by the observation of the interdependent factors of air temperature distribution, room air velocity, mean radiant temperature, and relative humidity, as well as from the comments of the occupants of the houses studied. An outline of field test procedure and a description of test equipment are given as a basis for future work. Their conclusions were as follows:
1. Size, shape or construction materials of the structures had little effect on the overall performance of the baseboard radiation.
2. Air temperature differentials from floor to ceiling and from room to room were less than in houses heated by more conventional systems."
3. Baseboard radiation systems were free from inherent drafts.
4. Indoor relative humidity was observed to be satisfactory without the use of humidification devices.
5. Highly satisfactory results were obtained from the use of simple control systems.
B-24
17. Olivieri, J. B., How to Design Heating-Cooling Comfort Ceilings. Business News Publishing Company, Birmingham, Ml, 1971.
A textbook on the design of HVAC systems. It contains a very short section on the description of radiant heating systems.
* * *
18. Peach, J., "Radiators and Other Convectors", J.I.H.V.E., Vol. 39, Feb., 1972.
This paper is concerned with equipment which emits heat by the combined processes of radiation and convection. If the heat exchanger surface is exposed for all to see it is termed a radiator; if enclosed, a convector. Reference is also made to heat emitters in which air is forced over the heat exchanger surfaces by a fan incorporated into the unit. Heat can be discharged into a room from a unit point, along a line or over an area. In common parlance unitary equipment includes radiators, convectors, fan-coil units, etc. Linear equipment embraces skirting heaters and undersill heaters, whilst area equipment would include underfloor heating, electrically conducting paint, heated wallpaper, etc.
These three categories can exist in either of two forms: where the heat is either transferred to the air by free (natural) convection or by forced convection.
This paper is limited to unitary types under free and forced convection and linear types under free convection. Operation of the equipment on low temperature hot water systems only has been considered.
* • * *
19. Pierce, J. D.f "Application of Fin Tube Radiation to Modern Hot Water Heating Systems", ASHRAE Journal, Feb. 1963.
Discussion of radiant baseboard systems and how they perform. Only general information is presented.
* * *
20. Rapp, George, "Analysis of Free convection and Radiant Heat Transfer in Valance Heat Exchangers", ASHRAE Trans., Vol. 72, Pt. 1, 1966.
The specific fluid flow and heat transfer mechanisms present in valance and baseboard heat exchangers have been identified and quantified by application of modern convection and radiation theory. It was found that conditions strongly favorable to stable laminar free convection exist for temperature differences 50 < At < 300F; and that radiation transfer, apparently neglected heretofore, plays an important role at source temperatures of 200 F and higher.
Mathematical general equations giving the net radiant flux within and heat balance on any valance exchanger enclosure have been derived and their application illustrated numerically by prediction of the radiation-convection
B-25
outputs of a specific design at the source temperature of 200 and 300 F. The predicted results agree quite well with full size room rating tests on a similar valance when both are compared at the 200 F source temperature level, the analysis showed that 100 F increase in source temperature (from 200 to 300 F) approximately doubled the output.
The importance of the valance vs the ceiling on occupant irradiation and body heat loss was evaluated. It has shown that neither the valance (at temperatures up to 300 F) nor the lower temperature ceiling is of much importance in control of body heat loss, as compared with the larger area unheated room surfaces; and that valance temperatures much higher than 300 F would be required for any worthwhile effect in this regard.
* * *
21. Rickman, James D., "Selected Segment Hydronic Heating System", Energy Conversion Mgmt., Vol. 25, No. 1, Pg. 73-83, 1985.
A new type of hydronic forced hot-water heating system is described that provides room-by-room temperature control in a series pipe loop system. It operates by alternately pumping segments of hot boiler water and cool return water in a timed heating cycle. Experiments show that the system delivers at least 49% more BTU's per foot of pipe length to rooms selected for heating than to unheated rooms. Calculations show that temperature setbacks of 3-9^ in selected rooms reduce the yearly heat loss by 5.5-20% depending on the setback schedule.
* * *
22. Roots, W. K., "Electric Space Heating with Active Boundary Members", Institution of Electrical Engineering Proc. Vol. 114, No. 7, July, 1967.
The results of earlier papers permit the building of computer models of electric space-heating processes. In common residential processes, the boundary members (walls, ceiling, floor) have hitherto been passive, and so were representable by available mathematical models. A recent development in North America is the wall panel constructed of glass fibre into which is woven a fine mesh of electric-heating wire, thus combining the functions of thermal insulation and low-density electric heating. These panels are mounted over large areas of the walls and ceiling, and so now many of the boundary members can become active. New mathematical models are required to represent such active members, and these are described.
* * *
23. Weigel, R. H. and Harris, W. S., "Heating a Basementless House with Radiant Baseboard", Trans. ASHRAE, V. 55, 1949.
A summary of the test results of this investiagtion and the conclusions that may be drawn are as follows:
1. The radiant baseboard is particulary adapted to maintaining
B-26
comfortable floor slab temperatures in a basementless structure, since long, low units of this type cover a large percentage of outside exposure.
2. Average air temperatures, as measured three inches above the floor, were approximately 70 to 71 F for all indoor-outdoor temperature differences encountered when the temperature at the 30-rin. level was 72 F.
3. Temperature differences between the occupancy zone, as measured 3 in. and 30 in. above the floor, were only of the magnitude of a degree and a half to two degrees.
4. The hot-water heating system using radiant baseboard responded quickly to sudden changes in load, thus maintaining constant room air temperature even while the outdoor temperature was changing rapidly.
5. The problem of maintaining adequate indoor humidities for comfort in winter cannot be separated from consideration of good building construction. At 10 F outdoor temperature, 22 percent relative humidity was obtained in the basementless home.
6. Fuel consumption was*not affected by setting of the adjustable differential on the thermostat. However, water temperatures in the radiant baseboards, and consequent fluctuations in room air temperatures, were affected. The longer lengths of the on and off periods of the burner and circulator resulted in greater fluctuations of room air temperatures.
7. Faint dirt patterns were observed on some of the walls above the radiant baseboard after nine months of operation. Dirt patterns of this type can be eliminated by limiting the water temperature to a maximum of 200 F.
B-27
D. Comfort Conditions
1. Berglund, Larry, "Mathematical Models for Predicting the Thermal Comfort Response of Building Occupants", ASHRAE Trans., Vol. 84, Pt. 1, 1978.
Three comfort models are reviewed and equations are presented. They all use the heat balance equation together with some physiological parameters to predict the thermal sensation of a person in an environment. These models are the Pierce, KSU, and Fanger. These models are used when simulating a building's thermal behavior and coupling it to comfort conditions.
* * *
2. Bull, L. C , "The Practical Approach to Heat Loss Calculation", Building Services Engineer. Vol. 41, March, 1974.
In this paper, the author tries to show, through simple numerical examples, that the environmental temperature method
Te - (2/3) MRT + (1/3) T a i r
of calculating heat transmission in buildings gives identical answers to those arrived at through use of tedious heat balance equations. He suggests, however, that the proper determination of room heat requirements involves the use of resultant temperature as a comfort index and environmental temperature for heat flow. A step by step procedure is presented.
* * *
3. de Heer, T. and Erkerlens, H. J., "Heat Transmission by Radiation", Journal of the Institution of Heating and Ventilating Engineers. London, Jan., 1963.
The authors have developed a new method for calculating radiant exchange by subdividing the surfaces into small equal squares. The size of the squares is dictated by the accuracy one wants in the results. This technique allows local intensities to be more accurately predicted. This is applicable in evaluating the exchange of energy between a person and his surroundings, where local radiation intensities are important.
* * *
4. Fanger, P. 0., "Calculation of Thermal Comfort: Introduction of a Basic Comfort Equation", ASHRAE Trans.. Vol. 73, Pt. ii, 1967.
This paper presents the development of the Fanger Comfort Equation and the Fanger Comfort Equation and the Fanger Comfort Charts, which are given in the ASHRAE Handbook of Fundamentals. It provides much of the data and figures which exist in the current comfort literature of ASHRAE. Many references are also provided for the development of the comfort equation.
3-28
5. Fanger, P. 0., Thermal Comfort - Analysis and Applications in Environmental Engineering. McGraw-Hill Book Co., New York, 1972.
This is a classic textbook discussing all of the parameters of human comfort and how they are evaluated and interrelated. It contains the following chapters: Introduction, Conditions for Thermal Comfort, The Influence of Certain Special Factors on the Application of the Comfort Equation, Practical Assessment of Thermal Environments, Calculation of Mean Radiant Temperature, Radiation Data for the Human Body, Thermal Environmental Analysis and References. It is very useful for making calculations and finding references.
B-29
6. Gagge, A. P., Fobelets, A. P. and Berglund, L. G., "A Standard Predictive Index of Human Response to the Thermal Environment", ASHRAE Trans.. V. 92, Pt. 2, 1986.
Temperature and sensory indices of human response to the thermal environment are often expressed in terms of the known in a controlled laboratory environment, as a standard. The three rational indices of this type to be considered are (1) ASHRAE's Standard Effective Temperature (SET*) Index, defined as the equivalent dry bulb temperature of an isothermal environment at 50% RH in which a subject, while wearing clothing standardized for activity concerned would have the same heat stress (skin temperature Tg^) and thermo-regulatory strain (skin wettedness, w) as in the actual test environment; (2) Fanger's Predicted Mean Vote (PMV) Index, defined in terms of the heat load that would be required to restore a state of "Comfort" and evaluated by his Comfort Equation; and (3) Winslow's Skin wettedness Index of "Thermal Discomfort" (DISC) defined in terms of the fraction of the body surface, wet with perspiration, required to regulate body temperature by evaporative cooling. The classic difference between PMV and DISC as predictors of warm discomfort occurs at very high and very low humidity but both lead to essentially the" same judgment at average humidities (40-60% RH or 1-2 kPa) . A new index PMV* is proposed for any dry or humid environment by simply replacing operative temperature TQ in Fanger's Comfort Equation with SET*. The use of PMV* as a sensor of heat stress and strain, is illustrated for typical HVAC situations and with a new Comfort-Humidity psychometric chart for indoor environments.
• * * *
7. Harrison, E. , "Environmental Temperature and the Calculation of Heat Loss", Building Services Engineer. Vol. 41, March, 1974.
The nature, nomenclature and relative numerical values of the various temperatures involved in load calculation for rooms have been variously described and discussed and there is confusion in some of the IHVE Guide calculations. The author points out that much of the confusion arises from a failure to distinguish clearly between the temperature used in calculating the heat requirements and the actual, measureable temperatures produced in the room when the rate of heat supply suggested by the calculation has been used.
The discussion centers around the definition of the environmental temperature, Te and its definition. The IHVE Guide defines this as
Te - (2/3) MRT + (1/3) Tair
Corrections are proposed to the definition of Te based on comfort conditions and calculations. The author presented these for convective heating systems and in a later paper for radiant heating systems.
B-30
8. Heerwagen, D. R., KIppenhan, C. J., Emery, A. F., and Varey, G. B., "Developing Office Building Design and Operation Strategies Using VWENSOL and the Comfort Routine", ASHRAE Trans.. V. 86, Pt. 1, 1980.
This article discusses the development of a "comfort condition computer program" for use with an energy analysis computer program. They have applied this to several situations in buildings. They consider the effect of cold or warm surfaces radiating to the occupant (strong influence), as well as window size and number of glazing panes (strong influence). Similar calculations were carried out for the effect of exterior envelope on comfort effects and alternate temperatures, and clothing conductances on comfort and energy requirements.
* * *
9. Madsen, T. L. , Olesen, B. W., Kristensen, N. K., "Comparison Between Operative and Equivalent Temperature Under Typical Indoor Conditions", ASHRAE Trans.. V. 91, Pt. IB, 1985.
In some of the new standards for thermal environments, a certain operative temperature range is given as a requirement for the thermal environment. In other standards the requirement is given as an interval for the PMV-index. This index takes the activity level, clothing, air humidity and the equivalent temperature into account. In this paper it is discussed whether the operative temperature is satisfactory or the equivalent temperature should be used as a better expression because it also takes the air velocity into account. It is shown that the use of the equivalent temperature can save energy during summer conditions but also that it can be necessary to increase the temperature during winter conditions in order to keep the thermal comfort at an acceptable level.
Their conclusions were as follows: The operative temperature is a good one to use when evaluating the heating
and cooling loads of a room or building, but it is only useful for describing the general thermal comfort at air velocities < 0.1 m/s.
The equivalent temperature is more correct, even at higher air velocities, because it takes into account the total dry heat loss from a person, i.e., it measures the integrated influence of air temperature, mean radiant temperature, and air velocity. A sensor for measuring the equivalent temperature directly has been described.
The general thermal comfort may be described more correctly and accurately by using an index like PMV or by using the equivalent temperature than by using the operative temperature alone. Only then is it possible to minimize the expenses for heating and air-conditioning buildings without sacrificing an acceptable thermal environment.
B-31
10. Mclntyre, D. A., "Evaluation of Thermal Discomfort", Electricity Council Research Centre, Capenhurst, England, October 1984 (NTIS - PB 85-189975).
This paper examines the techniques which have been employed to assess the degree of thermal discomfort, and the criteria which have been used in setting acceptable limits to environmental variables, by both researchers and official bodies. It is shown that there are many inconsistencies. Some are internal, where subjects' assessments on different scales during the same experiment apparently disagree. Other inconsistencies exist between different experimenters, where apparently similar comfort recommendations have been arrived at by very different reasoning. The paper considers the problems involved in transferring subjective judgements made in the laboratory to the real world, and discusses to what extend field studies of comfort and behavior can contribute.
* * *
11. Michaels, K. B. , Nevins, R. jG. and Feyerherm, A. M., "The Effects of Floor Surface-Temperature on Comfort Part III, College Age Females", ASHRAE Trans.. V. 70, 1964.
The author's conclusions were as follows:
1. For college-age females undergoing 3-hr test periods at rest with air temperature at 75 F and floor temperatures ranging from 75 to 100 F, the data show that there exists a statistically significant effect of floor temperature on foot comfort vote. With increasing floor temperatures means for foot comfort scores, ranging from 2.03 to 2.63, moved away from an ideal "2" for comfortable toward "3" for hot. At the same time, the means for thermal sensation, ranging from 3.38 to 3.76, moved from slightly cool "3" roward an ideal "4" for comfortable.
2. Results for tests with college-age females standing while performing light work also showed significant effects of floor temperature on foot comfort but not on thermal sensation. Sample means for thermal sensation scores ranged from 3.07 to 3.24 for floor temperatures from 75 to 100 F. The mean for foot comfort scores ranged from 1.87 to 2.59.
3. Based on foot comfort, floor surface temperatures as high as 85 F do not cause serious discomfort when the air temperature is 75oF.
* * *
12. Moisan, A. and Lebrun, J., "Comfort in Damp Cold Air with Radiant Spot Heating", Proceedings of the Second Int. CIB/RILEM Symposium on Moisture Problems in Buildings. Rotterdam, 1974.
This study found a positive effect of air humidity on thermal sensation: for same operative temperature, a higher air moisture causes a warmer
B-32
sensation. This effect is much higher than that predicted by the comfort equation in
the thermal neutrality region and with RH > 67% . But it appears that there is some discomfort directly related to humidity itself.
Perhaps other effects would appear if the time of permanency was increased by some hours in high humidity conditions. On the other hand, no other criteria concerning physiological or hygienic aspects have been considered.
* * *
13. Nevins, R. G., "Criteria for Thermal Comfort", Building Research. July-August, 1966.
A general review of thermal comfort conditions. Presents a history and some of the early references in this area.
* * *
14. Olesen, B. W., Mortensen, E., Thorshauge, J. and Berg-Munch, B., "Thermal Comfort: in a Room Heated by Different Methods", ASHRAE Trans.. V. 86, Pt. 1, 1980.
The present experiments have shown that all nine heating methods investigated are able to create an acceptable thermal environment in a well insulated room with one frontage including a double-glazed window exposed to steady-state winter conditions (outside temperature down to -5oC, and air infiltration rates up to 0.8 air-changes/h).
When the temperature level in a room provides thermal neutrality (PMV=0) for sedentary person near the frontage, there will be only a small likelihood of local discomfort and the thermal conditions will be acceptable in the entire occupied zone.
Only with a radiator at the back wall did the predicted percentage of dissatisfied (PPD-value) at a position near the radiator increase significantly from the optimal value (from 5 to 12%) .
In all tests, vertical air temperature differences, radiant temperature asymmetry and floor temperatures were inside established comfort limits.
There was a risk of mean air velocities higher than 10 cm/s along the floor in the occupied zone nearest to the frontage when the down-draft along the window and from the air infiltration was not counteracted by an upward convection from the heating system. In general, the highest measured air velocities were in the test with the two floor heating systems (approx. 15 cm/s).
* * *
15. Ronge, Hans. E., and Lofstedt, Borje E., "Radiant Drafts from Cold Ceilings", Heating, Piping, & Air Conditioning, Uppsala, Sweden, September, 1957.
Contains a statement of methods of study and results of 20 tests made on 5 persons who, after spending from 1/2 to 1 hr. beneath a warm ceiling, moved to beneath a cold ceiling in an experimental room and remained there until
B-33
temperature measurements on various body surfaces indicated equilibrium with surroundings. Tests included condition of (1) upper body naked and subject at rest, (2) one layer clothers and subject at rest and (3) 2 to 3 layers of clothes on upper body and subject doing light work. Air temperature covered range from 60.8 to 68 F. Good correlation is shown between skin temperature equilibrium values and mean of air and ceiling temperature. A comfort chart based on the results is included. Experimental results were checked against measurements made in an underground factory room with cool celing.
* * *
16. Rohles, F. H., Jr., "Temperature or Temperament: A Psychologist Looks at Thermal Comfort", ASHRAE Trans. . V. 86, Pt. 1, 1980.
Five studies are reviewed which address the psychology of thermal comfort. They are summarized as follows: (1) under identical temperature, adding wood panels to the walls, carpet, and comfortable furniture made people feel warmer than when they were in the stark, sterile setting of the room before it was modified; (2) at 65oF (18.3oC),. secretaries who were informed that a radiant heater was operating in the modesty panel of their desks, felt warmer than those who were not informed that it was operating; (3) when people were told that the temperature of a room was 74<>F (23.3<>C) when it actually was 72oF (22.2oC), 70oF (21.1oC), or 68oF (20oC), they were just as comfortable as when the room temperature was 74oF (23.3oC); (4) in a study to determine if comfort was related to the season of the year, it was found that cool temperatures are preferred over warm temperatures in the summer and the opposite is true in the winter; (5) based on a questionnaire in which temperatures were ranked as cooler-than-comfortable, comfortable, and warmer-than-comfortable, a Preferred Comfort Envelope was proposed that ranges from 70oF (21.1oC) to 76oF (24.4<>C).
* * *
17. Springer, W. E. , Nevins, R. G., Feyerherm, A. M. and Michaels, K. B., "The Effect of Floor Surface Temperature on Comfort: Part III, The Elderly", ASHRAE Trans.. V. 72, Pt. 1, 1966.
Conclusions reached in this investigation were the following.
1. For elderly females and males undergoing 3-hr test periods at rest with the air temperature at 80 F and floor temperatures ranging from 75 to 100 F, the data show that there exists a statistically significant effect of floor temperature on foot comfort vote. A statistically significant effect of floor temperature on thermal sensation did not exist for elderly males but did for elderly females. With increasing floor temperatures, the means for foot comfort ranged from 2.12 to 2.57 for the elderly males and from 2.25 to 2.64 for the elderly females.
At the same, the thermal sensation votes ranged from 4.28 for the elderly males and from 4.04 to 4.60 for the elderly females.
2. With an air temperature of 80 F, the data show that floor surface
B-34
temperatures as high as 85 F did not cause serious discomfort for elderly females or males.
3. With a floor surface temperature of 90 F the individuals are apparently experiencing a discomfort due to the inability to reject heat readily via conduction and radiation, but at a floor temperature of 95 F the greater stress has induced the body to dissipate a greater portion of the heat via moisture evaporation thus providing lower thermal sensation and foot comfort votes. Increasing the floor temperature to 100 F overcomes the advantage of increased evaporative losses and once more higher thermal sensation and foot comfort votes result.
18. Stevens, J. C. , Marks, L. E. and Gagge, A. P., "The Quantitative Assessment of Thermal Comfort", Environmental Research. Vol. 2, 1969, pp. 149-165.
Discomfort aroused by lowering or raising the operative temperature of a subject's environment was found to follow the "power law" that governs many dimensions in the domain of"sensory psychophysics. To a first approximation, discomfort caused by cooling grows as the 1.7 power of shifts downward in temperature from the level that feels comfortable; discomfort caused by heating grows as the 0.7 power shifts upward from the level that feels comfortable. One group of 8 subjects matched numbers to the degree of discomfort (magnitude estimation); another group of 12 subjects adjusted the loudness of a white noise to match the discomfort (cross-modality matching). These verbal and nonverbal methods gave approximately the same result with regard to the quantification of thermal discomfort.
B-35
E. Thermal Comfort-Radiant
1. Albert!, M. and Rugger, R. , "A Method to Check Thermal Comfort Conditions in High Industrial Buildings Provided with Radiant Panels Heating Plants", Clima 2000 - Indoor Climate. Vol. 4, W S Kongres - W S Messe, 1985.
A procedure was developed using the Fanger Comfort equations for determining the PMV and PPD indices in radiant heated, high ceiling industrial buildings. The method uses computerized procedures to evaluate the thermal environment at any position in an industrial building. This technique and calculation is applied to the Italian heating energy savings regulations. Very few details are given on how the equations are solved, what convection coefficients were used or what emissivities were used. Sample results of PMV are given as are some of the resultant surface temperatures.
* * *
2. Bahnfleth, D. R. , "Physiological Effects of High Intensity Radiant Beam Heating", ASHRAE Jnl.. Nov., 1964.
This is a description of ASHRAE activity in radiant heating during the early 1960's. It also discusses ASHRAE Research Project-41-Physiological Effects of Radiant Beam Heating.
* * *
3. Baker, Merl, "Improved Comfort Through Radiant Heating and Cooling", ASHRAE Journal, February 1960.
A description of the advantages of panel heating and cooling are presented along with how this affects the sensation of comfort. The article is not very in-depth.
* * *
4. Barihidi, L. , Dr., Somogyi, A., Kintses, G. , Besnyo, J., "About Local Discomfort Effects Caused by Asymmetric Radiation Occurring During Winter in Dwelling Houses", Clima 2000 - Indoor Climate. Vol. 4, W S Kongres - W S Messe, 1985.
This article deals with the laboratory analysis of a special case of a symmetric radiation that commonly occurs in everyday life. Namely, in winter what are the acceptable surface temperature boundaries for the inner side of a main front wall, in relation to the local discomfort feeling of a person sitting next to the wall? Sitting face to face with a cooled panel brought about in both cases of gradual and random temperature change more unfavorable reactions (20% discontent) of thermosensitivity than when the participant sat with back to the cooling wall (temperature limits were 22oC, 20oC and 16oC for facing with continuous temperature, facing with shorter random temperature
B-36
periods, and back to wall respectively).
* * *
5. Berglund, L. G. and Fobelets, A. P. R., "Subjective Human Response to Low Level Air Currents and Asymmetric Radiation", ASHRAE Trans. V. 93, Pt. 1, 1987.
"The responses of 50 subjects wearing winter clothing (0.86 clo) to two-hour-long exposures of various kinds of winter indoor conditions were studied. The conditions included air speeds between 0.05 and 0.5 m/s (10 and 100 fpm) and asymmetric radiation to a cold wall that produced radiant temperature asymmetries ranging from 0. to 20 K (0 to 36 F). The study was done at neutral or preferred temperatures and at conditions 3°C (5.A F) lower. Some of the conclusions are:
The operative temperature concept for combining air and mean radiant temperatures into a single temperature scale is an effective means of characterizing and controlling complex environments, although the coefficient A in the operative temperature equation of ASHRAE Standard 55-81 may be too low at high air speeds. - •
The neutral operative temperature, calculated according to ASHRAE Standard 55-81 from the experimentally determined neutral conditions, for velocities of 0.25 m/s (50 fpm) or less were unaffected by radiant temperature asymmetries of 10 K (18 F) or less.
Thermal acceptability at neutral conditions was unaffected by air speeds of 0.25 m/s (50 fpm) or less and RTAs of 10 K (18 F) or less.
Thermal acceptability decreased when radiant temperature asymmetry increased beyond 10 K (18 F) .
Thermal acceptability decreased when air speed increased from above 0.25 m/s (50 fpm) even at neutral conditions.
An operative temperature 3°C less than neutral is probably too low for human sedentary occupancy as thermal acceptance of such conditions was only 63% in this study.
There were differences in the subjective responses between the men and woemen of this study.
The perception of draft was a linear function of air speed and temperature and independent of radiant temperature asymmetry.
The sensation of local cooling was related to RTA and independent of air speed.
There was no interaction between velocity and RTA on the subjective responses of this study. That is, effects from velocity and radiant asymmetry are independent and additive.
Relationships were found relating thermal sensation with thermal preference, comfort, and thermal acceptability."
* * *
6. Berglund, L. G., Gagge, A. P., and Banhidi, L. J., "Performance of Radiant Ceiling and other Heating Systems Controlled for Equal Comfort with an Operative Temperature Sensor", Proceedings of the Third International Conference on Indoor Air Quality and Climate. Stockholm, August, 1984.
B-37
A typical office space was heated separately by (1) radiant ceiling panels, (2) forced air, (3) baseboard and (4) floor heating systems. Each system was controlled with the same "operative temperature" sensor, whose set point was held constant at 23<>c. Twenty subjects experienced each environment for 3 hours. The comfort and whole body thermal sensations were not statistically different for all four systems, although air and radiant temperatures and other thermal characteristics differed widely. The present experiments demonstrate that operative temperature, as a single control input, is an effective way to regulate heat for comfort in complex environments.
* * *
7. Berglund, L. G. and Gagge, A. P., "Human Response to Thermal Conditions Maintained in an Office by Radiant Ceiling, Baseboard, Forced Air and Floor Heating Systems", ASHRAE Trans., V. 91, Pt. 2, 1985.
A typical office space was heated separately by (1) radiant ceiling panels, (2) forced-air, (3) baseboard, and (4) carpet heating systems. Each system was controlled for equal comfort with the same "operative temperature" thermostat whose set point was unchanged throughout the test series. The office was contained within an environmental chamber. The steady-state power consumption per unit floor area for all systems averaged 9.1 Btu/hft^ (98 W/m^). The radiant system used the least power. Air temperatures in the occupied space were most uniform with the floor heating system. Twenty subjects experienced each environment for three hours. Periodically, subjects indicated their thermal sensations for whole body, head, and feet, local discomfort, comfort, and whether the environment was thermally acceptable or not. Comfort and whole body thermal sensations were not statistically different for all four systems and thermal acceptability averaged 94%. Though subjects indicated their feet were slightly warm with the carpet heating, they preferred this system to the others tested.
The experiments demonstrate the usefulness of operative temperature as a control parameter for complex environments, such as those produced by radiant ceilings. The thermal environmental characteristics of the four heating systems made it difficult to test each system at exactly the same comfort level. However, the wall-mounted operative temperature sensor-controller, whose set point was unchanged, maintained the comfort, thermal acceptability, and thermal sensation responses of the occupants in the test space at similar, if not identical, levels.
The air temperature in the occupied zone was most uniform with the floor heating and the least so with the baseboard system. It was also very uniform with the forced-air system. The difference between air and globe temperatures was the largest, with the radiant ceiling system using the least power. The floor system used the most power.
In these tests, the forced-air system caused the most local discomfort, followed closely by the radiant ceiling system. The occupants reported the least local discomfort with the floor system, even though the mean thermal sensation for their feet was slightly warm.
The occupants' preference ranking of the four systems from preferred on down was: floor,, baseboard, forced-air, and radiant. From the thermal sensation, comfort, acceptability, and local discomfort responses, one would have expected the radiant to be preferred over the forced-air system.
The radiant panel system used 89% of power per unit floor area as the
B-38
forced air-system. The baseboard system used 109%, and the carpet heating system used 116%.
Berglund, L. G. and Gagge, A. P., "Thermal Comfort and Radiant Heat", Proceedings of the Third National Passive Solar Conferencer American Section of the International Solar Energy Society, Univ. of Delaware, 1979.
The conclusions of the authors follow. In passive solar buildings, the air and mean radiant temperatures are
seldom equal. Comfort conditions in such buildings can be conveniently described in terms of the operative temperature of the environment. Operative temperature is approximately the average of the air and mean radiant temperatures present. Techniques for predicting the mean radiant temperature from the expected surface temperatures of the space are clearly described by Fanger. Fortunately for the passive building, humidity control is rather unimportant for comfort as is air movement below 30 fpm. An environment expected to be thermally acceptable to 80% or more of its sedentary occupants would have operative temperatures between 68 and 80oF. Of course over this large temperature range appropriate clothing adjustments (from 1 to .3 clo) would be necessary.
9. Bohgaki, K., "Effect of Floor Heating on Man's Comfort and Thermal Sensations", Clima 2000 - Indoor Climate. Vol. 4, W S Kongres -W S Messe, 1985.
The purpose of this study was to show the thermal effect of the posture in a room during sitting on the floor or sitting in a chair. From the experimental results, it was found that the case of sitting on the heated floor was more comfortable than sitting in a chair. Comparing the results of this experiment with other experiments indicated that the thermal environment was slightly warm. The comfort conditions for floor heating found in this study are as follow: air temperature, 18-20oC, floor surface temperature, 26-28oC, air velocity, under 0.1 m/s.
10. Boyar, R. E., "The Influence of Radiant Energy Transfer on Human Comfort", Heating. Piping and Air Conditioning. June 1966.
The author shows the variations in MRT that can be expected from different methods of providing heat (forced air or radiant) to a space. The radiant ceiling panels provided higher MRT near the glass surfaces than did the forced convection system. There was also a surprisingly low floor temperature with the forced air system. Some temperatures are presented from some experimental results for a forced air system and a radiant ceiling panel system.
B-39
11. Chrenko, F. A., "Heated Ceilings and Comfort", Journal of the Institution of Heating and Ventllatine Engineers. London, April, 1952.
The paper describes a series of experiments on five subjects, and further experiments on a group of 150 subjects, designed to obtain quantitative assessments of the risks of producing unpleasant conditions in rooms with heated ceilings and includes tables, based on the results of these experiments, which give maximum desirable surface temperatures of panels of various dimensions embedded in ceilings of different heights. The results of a separate series of experiments investigating the effects of a cold wall on persons exposed to radiation from a heated ceiling, did not modify the conclusions' as to the maximum desirable ceiling temperatures reached in the earlier tests. Ancillary observations made in the field are included in a discussion of the relation between the experimental results and existing British practice, and it is concluded that the recommended temperatures are, on the whole, in accord with current practice. The risk of discomfort, the causes of which are discussed, was most closely related to the elevation of mean radiant temperature at head-level due to the heated ceiling. A method of computing the mean radiant temperature at a point is given.
* * *
12. Chrenko, F. A., "Heated Floors and Comfort", Journal of the Institution of Heating and Ventilating Engineers. London, April, 1955.
The paper describes experiments designed to obtain quantitative assessments of the risk of producing discomfort in rooms with heated floors. The experiments were carried out in the laboratory on five men and three women who were (a) sitting and (b) walking about. The subjects wore their usual clothing and footwear. Measurements of the skin temperatures of the subjects' feet were made and the results showed that subjective reactions of the men and women to thermal stimuli were very similar, and such differences as were found between the responses of the two groups of subjects at various floor temperatures were due to differences in footwear. Discomfort was closely associated with the floor-surface temperature and with the temperature of the skin of the sole of the foot. It was concluded that the floor temperatures recommended here were undertaken in accordance with those accepted in current practice.
* • * *
13. Chrenko, F. A., "Radiant Heat and Thermal Comfort", Electricity and Space Heating. Proceedings of Symposium of the Institution of Electrical Engineers in London. March, 1964.
The authors conclusions were stated as follows. (1) Existing psychophysical scales of warmth are of no use in designing
new systems of radiant heating or the assessment of old systems.
(2) Studies of localized sensations of warmth have enabled design criteria
B-40
to be given in ordinary physical terms.
(3) Studies of the sensitivity of the skin to radiant heat have yielded results which have practical application in design.
* * *
14. Fanger, P. 0., Angelius, 0., and Kjerulf - Jensen, P., "Radiation Data for the Human Body", ASHRAE Trans., Vol. 76, Pt. 2, 1970.
The purpose of this study was to determine experimentally geometrical radiation data necessary for calculation of the radiant heat exchange between humans and their environment. This data is necessary when analyzing panel heating or cooling systems, infrared heating systems, and effects due to cold surfaces. Many figures are given for specific geometries.
* * *
15. Fanger, P. 0., Barihidi, L. , Olesen, B. W. and Langkilde, G. , "Comfort Limits for Heated Ceilings" ASHRAE Trans. , Vol. 86, pt. 2, 1980.
A curve has been established, showing the percentage of people feeling discomfort due to overhead radiation, as a function of the radiant temperature asymmetry. The curve applies for sedentary people who feel thermally neutral for the body as a whole. It is recommended that a heated ceiling should not provide a radiant temperature asymmetry exceeding 4<>C in spaces with high standards for the indoor climate. Less than 5% of the population are then predicted to feel uncomfortable due to overhead radiation. The corresponding limit for the ceiling temperature can be found from a figure for different sizes and heights of the heated ceiling. Increasing discomfort due to increasing overhead radiation with lowered air temperature, can be attributed to warmer head and colder feet.
* * *
16. Fanger, P. 0., "Radiation and Discomfort" ASHRAE Journal, February 1986.
This study presents results that indicate that people are not particularly sensitive to asymmetric radiation from surfaces. It shows that in practice the limits will rarely be exceeded. Clothed subjects were exposed to radiant asymmetry from walls and floors for 3 1/2 hours. In order to have 10% dissatisfied occupants it requires from 7 to 25oC radiant temperature asymmetry. A useful diagram relating percent dissatisfied to the angle factor yields the allowable temperature difference between the air and wall.
B-41
17. Gagge, A. P., "Final Progress Report: RP-41-Physiological Effects of High Intensity Radiant Beam Heating", ASHRAE Jnl.. April, 1968.
The general aim of the project has been the development of a comfort standard for high temperature sources of thermal radiation in the spectral range 0.7 to 20 microns. Briefly stated the conclusions of RP-41 are as follows:
1. For environmental conditions with varying ambient temperatures and radiant heat, comfort may be described by the operative temperature.
2. Comfort is not described by any single temperature level, but usually falls in a wide range of operative or ambient temperatures.
3. Physiological factors, such as metabolic rate, evaporative loss and the vascular regulation of peripheral blood flow, can each affect the level of thermal equilibrium and perhaps comfort by 2 to 4 degrees fahrenheit.
4. A normal physiological state is not necessarily the most comfortable to an individual.
5. A method has been presented showing how much radiant heat may be required to balance out the discomfort of low ambient air temperatures. A practical level of operative temperature for comfort useful for this method is 80oF (unclothed) and 72<>F (clothed).
* * *
18. Gagge, A. P., Hardy, J. D. , and Rapp, G. M. , "Exploratory Study on Comfort for High Temperature Sources of Radiant Heat", ASHRAE Trans., Vol. 71, Pt. 2, 1965.
For environmental conditions with varying ambient temperatures and radiant heat, comfort may be described by the operative temperature. A constant level of comfortable operative temperature implies either that there is thermal equilibrium with the environment or that there is a constant but small rate of body heating or body cooling.
A method has been presented, showing how much radiant heat may be required to balance out the discomfort of low ambient air temperatures. A practical level of operative temperature for comfort useful for this method is 80 F (unclothed) and 72 F (clothed).
* * *
19. Gagge, A. P., Rapp, G. M., and Hardy, J. D., "Mean Radiant and Operative Temperature for High-Temperature Sources of Radiant Heat", ASHRAE Journal, October 1964.
The objective of this paper was to outline a method of standardizing measurements of high intensity radiation by any radiometer in order to derive a mean radiant temperature (MRT) and an operative temperature (TQ), which can
B-42
describe with greater accuracy man's thermal response to his environment. In the study, the radiant energy received by the unclothed subject from a high-temperature source has been evaluated without specifying its quality by the increase in his sweat rate. The experimental conditions have been so chosen that the increase in the subject's radiation load was balanced by an equal increase in his evaporation loss. The MRT and T0, used here to describe any radiant environment, have been defined in terms of the radiation (hr) and convection (hc) coefficients that would have applied under corresponding conditions of equal wall and air temperature. Thus, it has been possible to describe a complex radiant environment in terms of a temperature scale well associated with everyday experience.
* * *
20. Gagge, A. P., Rapp, G. M. , and Hardy, J. D., "The Effective Radiant Field and Operative Temperature Necessary for Comfort with Radiant Heating", ASHRAE Journal, 1967.
A new hypothetical variable, effective radiant field (ERF) has been introduced that has useful application to radiant heating (and cooling). The effective radiant field is defined as the radiant heat exchanged by an occupant with his surrounding environment when his black body skin (or clothing) temperature is hypothetically equal to the ambient air temperature. The ratio of the ERF to the environmental constant represents the environmental temperature change caused by the radiant field. The sum of this change and the ambient air temperature describes the operative temperature by which comfort and discomfort may be judged in accordance with known physiological and comfort standards. For high temperature radiant heat, the ERF is equal to the actual heat absorbed by the occupant from the radiating source. ERF is a summative term and may include the radiant heating from a high temperature source, the re-radiation from warmed floors or ceiling, and the cold radiation from window surfaces. For constant ambient air temperature, comfort and physiological response are directly proportional to ERF and to the difference between mean radiant temperature and ambient air temperature.
* * *
21. Griffiths, I. and Mclntyre, D. , "Subjective Response to Overhead Thermal Radiation", Human Factors. 16 (4), 1974.
Forty subjects in two experiments experienced conditions in which ceiling temperatures varied between 26.5oC and 45©C. Air temperature, mean readiant temperature, air velocity, and humidity were held constant. Experimental variables consisted of two levels of seat height, subject baldness, and environmental temperature. The subjects appraised the environmental conditions by use of a 34-item semantic differential questionnaire. Baldness and seat height were unimportant factors, but ceiling temperature significantly affected warmth assessment. Conditions of higher ceiling temperature were perceived as cooler than those with the same mean radiant temperature and lower ceiling temperatures. Raising the air temperature did not increase sensitivity to overhead radiation, and raising the ceiling temperature did not cause discomfort. The results indicate greater
B-43
snesitivity to radiant exchange with walls than with the ceiling, and that European upper limits for ceiling temperature are unduly restrictive.
* * *
22. Griffiths, I. D. and Mclntyre, D. A., "Radiant Heating and Comfort", The Building Services Engineer. V. 40, June, 1972.
What the authors have sought to do is to evaluate the evidence which exists concerning radiant temperature and comfort. It seems that the standard recommendations are based on poor evidence where satisfactory evidence, in fact, exists and that the environmental temperature formula should be something like
Te = 0.56Ta + 0.44Tr
if the air velocity is equal to or less than 0alm/s. It also seems that there is no evidence of a preference for a radiant rather than a convective environment and that those designing installations should decide between radiant and convective systems on their individual merits.
* * *
23. Griffiths, I. D. and Mclntyre, D. A., "The Balance of Radiant and Air Temperature for Warmth in Older Women", Environment Research. 6, 1973.
Fifty-six women over 55 years of age (mean age 67.5 yr. standard deviation 6.4 yr) experienced three environmental conditions of equal predicted subjective warmth, but different mean radiant and air temperatures (air temperature 26.9 C, mean radiant temperature 17.3©C; air temperature 23.0oC, mean radiant temperature 23.7»C; air temperature 19.2oC, mean radiant temperature 26.8 C). After 40 minutes, exposure subjects rated the environment on a number of subjective scales, there were no significant differences between conditions. This supports a previous finding with young men as subjects, that radiant and warm-air environments are not perceived differentially and also suggests that the relative importance of air and mean radiant temperature for warmth is not affected by age.
* * *
24. Hardy, James D., "Physiological Effects of High Infrared Heating", ASHRAE Journal, November 1962.
Discussions concerning the physiological properties of skin, as related to radiant energy, are given. Caution should be used when infra-red radiant energy is used for concentrated heating.
B-44
25. Hart, Gordon H., "Heating the Perimeter Zone of an Office Building: An Analytical Study using the Proposed ASHRAE Comfort Standard (55 -74R)", ASHRAE Trans., Vol. 87, pt. 2, 1981.
Using ASHRAE Standard 55-74R along with fundamentals of radiant and convection heat transfer, an analytical study was performed to demonstrate the dependence of operative temperature on outdoor temperature. Three different types of heating systems were analyzed: baseboard convection, all-air, and radiant panel. One particular building wall construction, which conforms to the ASHRAE Standard on Energy Conservation in New Building Design (90-75R), was evaluated. The results suggest that it would be necessary to raise the air temperature in the space to a slightly higher value to maintain a constant operative temperature. An alternative solution would be the use of a wall section with less glazing area. The results of this study should be useful in designing and operating office buildings with uniform comfort in the perimeter zones.
* * *
26. Herrington, L. P., and Lorenzi, R. J., "Effect of Panel Location on Skin and Clothing surface Temperature", Heating. Piping. & Air Conditioning. October 1949.
Since room comfort is closely related to the surface temperature of an occupant's skin and clothing, the authors conducted experiments recording these variables under comparable conditions. The purpose was to find what effect the location of the radiant heating panel had on the human body. The authors conducted tests for five days each on the effect of floor heating conditions and ceiling heating conditions. Measurements were made of the effect of various temperatures of floor surface, room center and ceiling surface on subjects, as indicated by head temperatures (an average of cheek, upper hair surface, and dorsal neck area) and mean exposed skin temperatures and clothing surface temperatures. Then, mean comfort votes and mean temperature votes were taken.
One of the primary conditions of thermal comfort is a skin temperature ranging from approximately 80 F on the toes and sole of the foot to approximately 95 F on the trunk and certain facial areas, with an overall average for the skin surface of 90-92 F.
Their conclusions were as follow: 1. At a room center black body temperature of 75 F, radiant
floor panels operating at 79 F produce a detectable increase in the temperature of the clothing surface of the lower extremities. Under these conditions the gradient between extremities and environment is increased about 28 percent over comparable heating with ceiling panels operating at 95 F.
2. Foot temperatures under comfortable conditions requiring no house heating are about 10 deg above shoe surface temperatures which are near 74 F when floors are at 71 F, other surfaces and air temperatures being within 3 deg of this value.
3. Physiological considerations are reported which support the view that floor temperatures above 75 F are not desirable.
B-45
4. Since ce i l ing location of the radiant panel does not produce a significant effect on head temperature, this location for a radiant panel i s preferred to the floor location.
* * *
27. Houghten, F. C., Gunst, S. B. and Sue in, J., "Radiation as a Factor in the Sensation of Warmth", ASHVE Trans. . Jan. 1941.
This is a discussion about a set of experiments relating radiation to comfort conditions. It considers such factors as wet and dry bulb temperature, air velocity and mean radiant temperature. The authors develop a relationship between MRT and effective temperature. Changes in MRT dictate changes in effective temperature resulting in no change in feeling of warmth.
* * *
28. Langkilde, G. , Gunnarsen, L., Mortensen, N. , "Comfort Limits During Infrared Radiant Heating of Industrial Spaces", Clima 2000 -Indoor Climate, Vol. 4, W S Kongres - W S Messe, 1985.
Thirty-two subjects were exposed to overhead radiant heating from gas fired infrared heaters. The duration of exposure was 3 1/2 hours and subjects had different qualities of clothing, and some were standing and some seated. Relationships between percentage of dissatisfied and radiant temperature asymmetry were established for people exposed to infrared heaters. Accepting 5% feeling uncomfortable, a radiant temperature asymmetry of 10 to 14HC was found permissible. These limits are less restricitve than existing recommendations and standards. No effect of wearing a helmet was found.
•*• * *
29. Lebrun, Jean J., and Marret, Dominique J., "Thermal Comfort and Energy Consumption in Winter Conditions -- Continuation of the Experimental Study" ASHRAE Trans., vol. 85, Pt. 2, 1979.
Radiant heating systems by floor and/or ceiling panels are examined by the detailed measuring of the inside microclimate in an experimental room in relation with heat transmission through an exposed wall and with ventilation enthalpy flow. The efficiency of the system relates to thermal comfort conditions as well as to energy consumption. The Predicted Percentage of Dissatisfied is computed from every distribution of internal temperatures. What is presented is a rational interpretation of differences observed between experimental and theoretical values of the overall heat transfer coefficient of the exposed wall as of the volumetric heat loss coefficient. The results are compared with those previously obtained with a warm air system. Some information is given about comfort experiments with subjects, performed to confirm the physical diagnostic of the inside micro-climates realized by all the heating systems previously studied.
For a "badly" insulated building (with large, single-glazing areas), a satisfactory comfort is only achieved by radiators located below the windows or by warm air. heating; but this second solution is much more energy consuming
B-46
(due to the increase of in terna l heat exchange coefficient) . For a "well" insula ted building, radiant heating i s as comfortable as the
other systems; nevertheless there i s no miracle: radiant heating does not allow a s ign i f i can t lowering of a i r temperature and corresponding vent i la t ion l o s s .
For a "very well" insulated building, warm a i r heating would probably become the most energy saving one, because of the in ternal isothermy i t could produce. Ver t ica l a i r temperature prof i les are also given.
* * *
30. Mclntyre, D. A., "Overhead Radiation and Comfort", The Building Services Engineer. Vol. 44, Pg. 226-34, Batiste Publ. Ltd., January 1977.
This series of experiments has investigated the reactions of people to overhead heating, generally from a heating panel extending over the ceiling. A measure of asymmetry has been developed, termed the v.r.t.
A vector radiant temperature of 20K does not increase the mean discomfort vote. However, this level of radiation is noticeable, and may be blamed for causing discomfort. It is recommended that for normal indoor situations a v.r.t. of 10K be regarded as the upper limit. There is no difficulty in meeting this criterion in well insulated buildings, but buildings with poor insulation and large windows may require further insulation to reduce the power loading.
There is evidence to suggest that the direction of the radiation is important. Cold radiation on the back and warm radiation on the face are apparently disliked.
* * *
31. Mclntyre, D. A., "Overhead Radiation and Comfort", Electricity Council Reasearch Centre, Capenhurst, England, May 1976, (NTIS-PB 277 428).
The Environmental Section at ECRC has conducted a series of experiments aimed at an understanding of the effects of thermal radiation on comfort. This report summarizes the work on overhead radiation, and makes recommendations on the maximum loading of ceiling heating systems which can be used without risking complaints of discomfort. Five experiments are summarized. From this evidence, and the published results of other workers, it is shown that asymmetric thermal radiation characterized by a vector radiant temperature of greater than 10K produces noticeably non-uniform conditions, and is likely to lead to complaints by occupants. The vector radiant temperature (v.r.t.) is a measure of asymmetry, and may be thought of as the average surface temperature of one half of the room minus that of the other. This result is applied to the design of ceiling heating systems. The v.r.t. is related to both the size and temperature of the heated area, and hence to the heat output. In practice this sets an upper limit to the heat emission that may be provided without increasing the v.r.t. above the recommended limit of 10K. For rooms of normal height, the maximum installed load Pmax(W) is related to the heated area A(m squared) by Pmax - 700 + 95A. This simple relation should be used at the design stage to check the
B-47
acceptability of a system.
* * *
32. Mclntyre, D. A., "Radiant Heat from Lights and Its Effect on Thermal Comfort", Illuminating Engineering Society, London, V. 8, No. 3, 1976.
The thermal radiation received from a lighting system may be estimated from a knowledge of the lamp type and the illuminance. Fluorescent lighting produces an irradiance of about 8 W/mz per 1000 lux. This is unlikely to produce any adverse effect on comfort at conventional levels of illuminance below 4000 lux. Tungsten filament sources produce considerably more radiation, at 70 W/m^ per 1000 lux. Any installation using 1000 lux or more is liable to produce uncomfortable conditions.
* * *
33. Mclntyre, D. A., "Sensitivity and Discomfort Associated with Overhead Thermal Radiation", Ergonomics, V. 20, No. 3, pp. 287-296, 1977.
One hundred and forty-eight subjects each experienced one of four levels of overhead radiation, up to a maximum ceiling temperature of 45oC. The degree of asymmetry is characterized by the vector radiant temperature (v.r.t.); the four levels were 0, 5, 9 and 14 K. Air and wall temperatures were held equal to each other, and reduced to compensate for the raised ceiling temperatures, so that perceived warmth was constant across the conditions. After 15 min. exposure, the subjects rated the environment on seven scales. Scales of general evaluation showed a slight improvement with increasing asymmetry. However, a scale which asked whether the hot ceiling caused discomfort showed a steady increase in discomfort with increasing asymmetry. It appears that people are ready to attribute discomfort to unusual aspects of the environment. A maximum asymmetry of v.r.t. = 10K is therefore suggested as a design criterion; this level did not actually increase discomfort, but was noticeable and in practice levels greater than this are likely to produce complaints.
* * *
34. Mclntyre, D. A., "The Thermal Radiation Field", Building Science. Vol. 9, 1974.
This paper has presented a comprehensive view of the thermal radiation field. The concept of the field is probably unfamiliar to most workers in the field of comfort or thermal physiology, where it is more usual to deal in terms of energy exchanges. However, the field concept provides a simpler and more powerful way of describing the thermal radiation environment, and it is to be hoped that some of the measures, particularly the radiation vector and its equivalent, the vector radiant temperature, will be used when specifying the thermal environment. Fortunately, several differential radiometers are already on the market, and the measurement of the radiation vector presents no
B-48
problem. It is still not possible to obtain a reliable instrument for measuring the
mean radiant temperature and this fact more than any other has delayed the understanding and acceptance of the importance of the radiation field in determining comfort and warmth.
* * *
35. Mclntyre, D. A., "Thermal Radiation from Lighting Installations", The Building Services Engineer. V. 41, April, 1973.
The thermal radiation from a number of different light sources was measured as a function of the illuminance. A linear relation was obtained between total thermal radiation and illuminance, which implies that the thermal irradiance may be predicted from illuminance. The constant of proportionality varied from 0*07 W/m2 per lux for tungsten filament lamps to 0.006 w/m2 per lux for low pressure sodium lamps. The figure is 0.008 w/m2
per lux for fluorescent lamps, which implies that the increase in mean radiant temperature due to radiation from a fluorescent lighting installation is up to 0.7oC per 1000 lux. This may be compensated by reducing the air temperature 0.3oC per 1000 lux. Radiant" asymmetry per se will not cause discomfort at illuminances up to 15000 lux for fluorescent or 1700 lux for filament installations; the level at which discomfort occurs has not been established. The typical modern fluorescent lighting installation of 1000 lux will produce no comfort problems from thermal radiation.
* * *
36. McNall, P. E. and Schlegel, J. C , "The Relative Effects of Convection and Radiation Heat Transfer on Thermal Comfort (Thermal Neutrality) for Sedentary and Active Human Subjects", ASHRAE Trans.. V. 74, Pt. 2, 1968.
The following shows the relative influence of convection and radiation heat transfer, determined by the ratio of the convection heat transfer coefficient (hc) to the radiation heat transfer coefficient (h^), for people wearing clothes with an insulation value of 0.59 clo in equilibrium with environments that have a partial pressure of water vapor of 0.435 in. Hg:
1. Sedentary (relative air velocity = 25-30 fpm):
a. Males (metabolic rate = 389 Btuh), => 1.51
b. Females (metabolic rate = 301 Btuh) , he/ti£ = 1.37
c. Males and Females combined (metabolic rate = 345 Btuh), = 1.43, recommended value 1.4
2. Active (relative air velocity = 45 fpm):
a. Males and females combined (metabolic rate = 741), = 1.2, recommended value 1.4
B-49
The results show generally good agreement with Fanger's comfort equation In the environments investigated, including the validity of Fanger's equation for predicting thermally neutral environments.
The same ratio for sedentary and active subjects is felt useful for engineering purposes.
37. McNall, P. E., Jr^ and Biddison, R. E. , "Thermal and Comfort Sensations of Sedentary persons Exposed to Asymmetric Radiant Fields", ASHRAE Trans.. V. 76, Pt. 1, 1970.
The results of the statistical analyses performed on the votes of thermal comfort of sedentary male and female subjects wearing clothing with an insulation value of 0.59 clo in equilibrium with environments with a partial pressure of water vapor of 0.435 inches Hg and air velocity of 20-30 fpm indicate that:
1. The thermal sensations of subjects exposed with radiation shape factors of 0.20 to a wall 20 F cooler than the balance of enclosure surfaces and thermal sensations of subjects exposed to uniform enclosure surface temperatures belong to the same regression plane. Therefore the "Thermally Neutral Zone" developed in an earlier study for enclosure surfaces of uniform temperature is applicable for environments of the former type.
2. The regression planes, relating thermal sensation with air temperature and mean radiant temperature, developed for subjects exposed with radiation shape factors of 0.20 to a wall at 130 F and for subjects exposed to uniform enclosure surface temperatures were found to be significantly different, producing thermal sensation votes about 0.5 higher than expected in the case of the 130 F wall. Although the previously mentioned "Thermally Neutral Zone is not applicable for thermal environments of the former type, it is felt it applies for less severe exposure to heated panels.
3. Thermally "neutral" subjects exposed with radiation shape factors of 0.12 to ceiling panels at 50 and 130 F and radiation shape factors of 0.20 to wall panels at 50 F experienced no significan discomfort which could be attributed to the symmetry of the radiant field.
4. Thermally "neutral" subjects exposed with radiation.shape factors of 0.20 to wall panels at 130 F experienced significant discomfort which was found to be caused by the asymmetry of the mean radiant temperature.
B-50
38. Nevins, R. G. and Feyerherm, A. M., "Effect of Floor Surface Temperature on Comfort Part IV: Cold Floors", ASHRAE Trans.. V. 73, Pt. 2, 1966.
Limited tests involving cold floors, indicate that thermal sensations of both male and female college-age subjects are not seriously affected by floor surface temperatures as low as 60 F, air temperature 75 F. Subjects were clothed, seated at rest and exposed to the test conditions for 3 hr. Foot comfort votes indicated that both male and female subjects objected to a floor temperature of 60 F. In addition, a floor temperature of 65 F may be too cool for female subjects. Foot skin temperatures, measured on the bottom and top of the foot inside the shoe, roughly correlated with foot comfort vote.
Based on theoretical calculations, floor surface temperatures of 100 F may reduce the radiant heat transfer to 20% of that occurring in a uniform environment. For a 60 F floor temperature, the heat transfer may increase to 150 %. The radiation heat transfer in a uniform environment (air temperature equal to MRT) is approximately 25 to 30% of the total. The influence of the floor temperature varies with the size of the floor, the temperature and the location of the man.
* * *
39. Nevins, R. G. and Flinner, A. 0., "Effect of Heated-Floor Temperatures on Comfort", Heating Piping and Air Conditioning. Oct., 1957, p. 149-153.
For the first phase of a study to determine the effect of floor temperatures on comfort, comfort data were obtained by subjecting college age students, male and female, seated at rest, to various floor-panel-heated environments for periods of 60 min. A total of 108 male and 21 female students were used. Correlation coefficients were calculated for the correlation of comfort vote with air temperature, operative temperature, effective temperature, air temperature plus mean radiant temperature, floor temperature and relative humidity. The coefficients show that the comfort vote correlates with those parameters in which air temperature is a predominant factor. The comfort vote did not show a significant correlation with floor temperature or relative humidity.
Floor surface temperature over a range of 65 to 95 F were found to have a negligible effect on the comfort vote when the air temperature was 75 F. During the 1957 tests, using male subjects, it was found that floor temperature of 100 F significantly affected the comfort vote whereas floor temperatures of 80 to 95 F did not. It was concluded that 95 F is the maximum floor temperature for comfort under the conditions of these tests. Foot temperatures recorded at the end of the 60-min exposure indicate that 88 to 91 F is the maximum foot temperature for comfort under the conditions of these texts.
B-51
40. Nevins, R. G., Michaels, K. B. and Feyerherm, A. M., "The Effect of Floor Surface Temperature on Comfort-Part 1, College Age Males", ASHRAE Trans.. V. 70, 1964.
1. A statistically significant effect of floor temperatures on thermal sensation and foot comfort does exist for college-age males undergoing 3-hr test periods at rest with air temperature at 75 F and floor temperatures ranging from 75 to 100 F. With increasing floor temperatures, means for foot comfort scores, ranging from 2.08 to 2.36, moved away from an ideal "2" for comfortable toward "3" for hot. At the same time the means for thermal sensation, ranging from 3.64 to 3.95, moved from slightly cool "3" toward an ideal "4" for comfortable.
2. Results for tests with college-age males standing while performing light work also showed significant effects of floor temperature on comfort. Sample means for thermal sensation scores stayed close to the 4.0 mark for floor temperatures up to 95 F. The mean for 100 F was 4.43. The mean for foot comfort scores ranged from 2.09 to 2.55.
3. Based on foot comfort, floor surface temperatures as high as 85 F do not cause serious discomfort when the air temperature is 75 F.
* * *
41. Olesen, B. W. , "A Simplified Calculation Method for Checking the Indoor Thermal Climate", ASHRAE Trans.. V. 89, Pt. 2B, 1983.
At the design stage, it is desirable to be able to predict the indoor thermal climate that will result from a given combination of building construction, heating system, and outdoor climate. Several large computer programs can be used to predict heating and cooling loads, indoor air temperatures, surface temperatures, and humidity, but these programs are often difficult and expensive and are used mainly for large buildings. A simplified method of calculation for evaluating the thermal indoor environment of the design stage has been presented. It uses the operative temperature, floor surface temperature and radiant temperature asymmetry.
* * *
42. Schlegel, J. C. and McNall, P. E., "The Effect of Asymmetric Radiation on the Thermal and Comfort Sensations of Sedentary Subjects", ASHRAE Trans.. V. 74, Pt. II, 1968.
Experiments were conducted by exposing sedentary subjects for 3 hrs to environments of symmetric and asymmetric MRT that were in and around the thermally neutral zone. The air velocity was held constant at 25-30 fpm and the vapor pressure of moisture in the air was maintained at 0.435 in. Hg (45% relative humidify at 78 F, for example).
For the symmetric MRT tests reported previously, the MRT was separated up to 12 F from air temperature. In the asymmetric MRT tests, one entire wall of the test chamber was cooled or heated approximately 12 F different from the
B-52
balance of the walls, ceiling, and floor, and the subjects were seated relative to that wall such that a sphere in the position of each subject would have a shape factor of 0.2 with respect to that wall. The surface temperatures were adjusted so that MRT was equal to air temperature for a sphere in the position of each subject. A two-sphere radiometer developed by Honeywell was used to measure the mean radiant temperatures.
The conclusions are: 1. No significant discomfort was noticed by the subjects due to the
asymmetric MRT of the magnitudes tested. However, since the subjects used in the present tests were not allowed to participate in repeated exposures, it cannot be concluded that sensations of comfort or neutrality are sufficient to rule out harmful effects that may exist for unilateral cooling, even for the relatively mild temperature differences employed in these tests.
2. The seated subjects could be approximated as spheres for mean radiant temperature calculations, as evidenced by the radiometer readings of mean radiant temperatures which accurately predicted the subjects' thermal sensations.
3. The experiments employed only low temperature radiation near room temperature levels. They do not apply to infrared sources.
43. Spangler, A. T. , "Industrial Climate Control Versus Radiant Heat", Air Conditioning. Heating and Ventilating. Jan., 1965.
This article considers the effects of radiant heat on the heat dissipation from the human body. A chart and basic equations are presented for calculating these effects along with the necessary adjustments in ambient air temperature required to maintain the same relative degree of human comfort. Various methods are also presented for reducing radiation effects and mean radiation temperature. Use of these methods has made climate control practical in applications that would have otherwise been impossible.
44. Smith, R. M. and Rae, A., "Patient Comfort and Radiant Ceiling Heating in a Hospital Ward", BuildinE and Environment. V. 12, pp. 143-146, 1977.
An experimental survey of the limitations placed by patient comfort considerations on the size and surface temperature of infra-red ceiling heating panels in a hospital ward is described.
It was found that at ceiling surface temperatures of up to 50oC patients suffered no additional discomfort with angle factors, based on a parallel planes measure, of up to 0.31 to the heated surface, and that at ceiling surface temperatures of up to 60oC angle factors of up to 0.12, where the air temperature was limited to 23©C, and of up to 0.31, where the air temperature was limited to 21oC, were permissible.
B-53
45. Stevens, Joseph C., and Marks, Lawrence, "Subjective Warmth in Relation to the Density, Duration, and Areal Extent of Infrared Irradiation", ASHRAE Trans., Vol. 76. Pt. 1, 1970.
The conclusions from this experimental investigation were as follows. 1. Subjective warmth aroused by infra-red irradiation of the skin grows
in magnitude as a power function of the flux density of the irradiation. The exact size of the exponent of the power function depends on the duration of the exposure and its areal extent.
2. Subjective warmth correlates well with flux density but poorly with flux duration. Since skin temperature depends on both density and duration, the proximal stimulus for subjective warmth cannot be skin temperature per se. A number of other common hypotheses concerning the nature of the proximal stimulus are also at odds with the data on apparent warmth. Two theories, one involving a two-layer receptor system, the other a single receptor system possessing properties of adaptation, are reasonably consistent with the data on warmth.
3. Flux density and areal extent can be traded for each other to preserve the same sensation of apparent warmth. At very low sensation levels, the rule of trading is virtually complete reciprocity. With increasing sensation level, area becomes a progressively less effective determinant of warmth than flux density, and as a result, at higher sensation levels it takes a much larger percentage change in area to offset a given percentage change in density.
* * *
46. Tredre, B. E., "Assessment of Mean Radiant Temperature in Indoor Environments", Brit, J. Industr. Med, V. 22, 58, 1965.
The conclusions of the aughor are quoted below. Attention has been drawn to the fact that in non-uniform environments (1)
the mean radiant temperature varies throughout the occupied space, (2) the average surface temperature is a dubious index of mean radiant temperature, (3) errors will arise in assuming that the mean radiant temperature at a point is equivalent to the mean radiant temperature as its affects a man, and (4) the heat load on a man is dependent on his orientation towards the various radiating surfaces.
While the mean of the estimates of m.r.t.s from two or three globe thermometers gave good agreement with the results obtained from sitting or standing metal models of men respectively, the error in the use of one globe at 45 in. above the floor was only small. Differences between m.r.t.s obtained with the thermopile at 45 in. above the floor were slightly greater but still did not exceed 0.62 C and 1.17 C for the standing and sitting models respectively. The results suggest that the 45 in. reference point should be placed in the region of the knees rather than the trunk when a man is seated with his legs near a heated panel.
Hence it seems that in indoor environments warmed by any of the more usual forms of heating installation but with no intense radiation from sharply localized sources, the simple estimation of mean radiant temperatures with a globe thermometer at 45 in. above the floor will generally give a very
B-54
satisfactory indication of the heat load on a standing or sitting man.
I
1 B-55
F. FLOOR PANELS
1. Billington, N. S., "Floor Panel Heating - Some Design Data", Journal of the Institution of Heating and Ventilating Engineers. London, Oct., 1953.
The use of the network analyser to study some probleems of floor heating is described. Data are presented relating to:
a) Surface temperature for different tube spacings and depth of the tube in the concrete floor.
b) Downward losses for different tube spacings and depths.
c) Emission of panel - from 1.2 to 0.5 Btu/ft* hr HF. This was extremely sensitive to tube spacing and depth.
2. Bruce, H. H., "Off-Peak Floor Heating - Research, Design and Development: Some Controversial Factors", Electricity and Space Heating. Proceedings of Symposium of the Institution of Electrical Engineers in London, March, 1964.
Discussion of floor warming in order to take care of loads occurring at a later time. Suggests that the recommended value of 22.5 w/ft^ for an intermediate floor is too high for comfort conditions. It will produce floor surface temperatures in excess of 85 F. He suggests that in some cases, 14 w/ft^ with a charge time of 12 hours in 24 will produce floor surfaces up to 85 F.
3. Faithfull, E. W., "Electric Floor Warmings in Commercial Buildings", Electricity and Space Heating. Proceedings of Symposium of the Institution of Electrical Engineers in London. March, 1964.
The author discusses several systems that can be used for warming floors. These are; hollow-pot floor, pre-stressed plank floor, hollow concrete beam floor, and cast in situ floor. He discussed the advantages of warmed floors for comfort and heating.
4. Grammling, F. J., "Methods for Testing Hydronic Floor-Heating Systems", ASHRAE Trans., V. 91, Pt. 2, 1985.
For the German standard DIN 4725, methods have been developed for testing the thermal performance of hydronic floor-heating systems, and numerous measurements have been made. Most of the systems have been tested with the
B-56
so-called place apparatus. The results of the tests show that measured values of performance differ significantly from figures published in the literature or company catalogs. I t i s therefore clear that exact performance measurements under controlled thermal conditions are necessary for the planning and optimal operation of floor-heating systems.
* * *
5. Hogan, Roy Edward, Jr., Heat Transfer Analysis of Radiant Heating Panels - Hot Water Pipes in Concrete Slab Floor. August 1979.
The ASHRAE design recommendations for radiant floor heating panels are reviewed and evaluated using the results of a numerical model. The numerical model is described in detail the results are compared to prior experimental data. Both bare and covered panels are considered. Particular areas of interest are the downward and edgewise heat loss, the panel thermal resistance, and the required mean water temperature. A transient simulation of the panel performance over a typical winter day is presented and a control system is discussed. Isotherms are plotted for the temperature field in both the panel and the earth. The ASHRAE panel model is acceptable for the geometry considered even though it does not represent the panel heat loss mechanisms correctly. Further studies could be made for other panel geometries, different infiltration rates, and an AUST not equal to the room air temperature. The numerical results agree in trend with the prior experimental results. The ASHRAE design recommendations are adequate and slightly conservative for designing both bare and covered radiant floor heating panels with no infiltration and an AUST equal to the room air temperature. These design recommendations are conservative because both the downward and edgewise heat loss and the panel thermal resistance are over estimated. The steady state design water temperature appears to be more than adequate for transient operation.
* * •*
6. Hogan, R. E. , Jr. and Blackwell, B., Ph.D., "Comparison of Numerical Model with ASHRAE Design Procedure for Warm Water Concrete Floor Heating Panels", ASHRAE Trans., Vol. 92, No. 2, Pt. 1, 1986.
The ASHRAE design recommendations for radiant heating panels are reviewed and evaluated using the results computed from a numerical model. The panel configuration that was considered consists of hot water pipes buried in either a bare or a covered concrete slab floor with a concrete footing and a perimeter insulation. The numerical model is described in detail and uses a finite control volume based solution method. Particular parameters of interest are the downward and edgewise heat loss, the panel thermal resistance, and the required mean water temperature. Results of the numerical model were compared with prior experimental results and agree qualitatively. The ASHRAE design recommendations are shown to be adequate and are slightly conservative for designing both bare and covered floor heating panels with no infiltration and,with an area-weighted average unheated surface temperature, AUST, equal to the room air temperature. The ASHRAE design recommendations are conservative because both the downward and edgewise heat loss and the panel thermal resistance are overestimated. .
B-57
7. Hulbert, L. E., Nottage, H. B., and Franks, C. V., "Heat Flow Analysis in Panel Heating or Cooling Sections", Heating. Piping & Air Conditioning. April 1950.
A set of curves applicable to the case of uniformly spaced pipes buried within a solid slab having isothermal surfaces has been developed for the purpose of relating the following quantities: (1) rate of heat release (or pick-up) per linear foot of pipe, (2) thermal conductivity of the slab, (3) difference between the pipe and mean slab surface temperatures, (4) spacing between adjacent pipes, (5) slab thickness, (6) tube outer diameter, and (7) position of the pipe grid between the slab surfaces. Simple equations are given for establishing the division of heat flow between the two surfaces. Data derived from analysis have received experimental confirmation.
* * *
8. Humphreys, C. M. , Franks, C. V., Schutrum, L. F., "Field Studies of Heat Losses from Concrete Floor Panels", Heating. Piping and Air Conditioning. January 1951.
During the summer of 1915, thermocouples and heat flow meters were installed under four houses to study the heat losses from floor panels to the earth. Different kinds and amounts of insulation were placed under the floor slabs. The results of the tests in these four houses during the 1949-50 heating season are reported, and some details of instrumentation are given. Their conclusions were:
1. The greatest part of the heat loss from a floor panel occurs around the perimeter of the panel. It is in this area that insulation will prove most effective.
2. It is particularly important that the panel be separated from the foundation by suitable insulation.
3. Four in. hollow clay tile does not appear to be any more effective as an insulation under a floor panel than an equivalent thickness of gravel. Its use as an edge insulation was not studied.
* * *
9. Humphreys, C. M., Nottage, H. B. , Franks, C. V., Huebscher, R. G., Schutrum, L. F., and Locklin, D. W., "Laboratory Studies on Heat Flow Within a Concrete Panel", Heating. Pining & Air Conditioning. April, 1950.
Equipment for studying heat flow within concrete panels is described and results of tests on three panels are reported.
An electrical analogue, developed to verify and extend the range of thermal test results, is described. Heat flow rates and temperature distribution, as determined by thermal tests and electrical analogue methods, are compared with values predicted from theory. The experimental studies confirm the fundamental theory given in another paper.
B-58
10. Hutchinson, F. W., Mills, D. L., and La Tart, L. J., "Losses from a Floor-Type Panel Heating System", Heating. Piping & Air Conditioning. December 1950.
This paper presents the results of an investigation on heat losses from two floor-type panel heating systems during the 1948-49 and 1949-50 heating seasons. Edge and rear losses were of the order of 30 percent of the total energy supplied so it may be concluded that insulation is necessary with floor-type systems if economical operation is desired. In addition to the loss data, experimental results are also presented giving the actual rating of a unit area of floor panel and the combined film coefficient of heat transfer for such a panel as evaluated from tests on the actual system; the latter results are of particular interest in that they permit conclusive decision as to the applicability of the laboratory tests of other investigators to actual field installations.
* * *
11. , Le Procede Calendal: de chauffage par rayonnement a' basse temperature", Reuve de 1'aluminium. Vol. -- No. 438, pg. 141-1565, Sebal, March 1975.
A French article discussing the specific details for installing floor panel heating systems. Numerical data is given and many photographs of installation procedures are given.
* * *
12. Macey, H. H. , "Heat Loss Through a Solid Floor", Institute of Fuel Journal. 22-128, p. 369.
A formula is derived for the estimation of the heat loss through a floor standing solid on the ground and surrounded by a wall, as in kilns, furnaces and driers. It involves two constants, one depending on the wall thickness and the other on the shape of the floor.
* * *
13. Perry, E. H. , Cunningham, G. T. , and Scesa, S., "An Analysis of Heat Losses through Residential Floor Slabs", ASHRAE Trans., V. 91, Pt. 2, 1985.
Based on the steady-state two-dimensional finite element analysis carried out in the present study, the following conclusions can be drawn:
1. For an unisulated slab about 60% of the total heat loss occurs through the region lying within three feet of the slab edge.
2. The heat loss coefficient is virtually independent of outdoor air temperature and varies only slightly with the deep soil temperature.
B-59
3. Soil thermal conductivity should definitely be taken into account in any estimate of the slab heat losses since the losses are directly proportional to the soil conductivity and since the latter parameter varies considerably from one soil type to another.
The study confirmed that the losses occur primarily near the edge of the slab and are proportional to the product of slab perimeter and the indoor/outdoor temperature difference, a relationship familiar to the HVAC community. However, the constant of proportionality was found to be strongly dependent on both the insulation configuration and the soil thermal conductivity. The latter dependency has seemingly been ignored in past studies. Thus, accurate predictions of slab heat losses must include considerations of the soil underlying the slab.
* * *
14. Plattis, R. E. , "Where Polyethylene Pipe Challenges Metal for Slab Radiant Heating", Canadian Builder, April, 1963, Pg. 55.
The author states, "Both laboratory and field experience show that low density polyethylene pipe should be fully suitable for concrete slab radiant heating systems if proper control is exercised."
He adds: "Pipe must meet recognized standards. Maximum temperature should be positively limited to 130oF and preferably 120oF by means of aquas tat cut-off valves on the feed line from the 'blender' to the slab. Pressures should not exceed 15 psi for sustained used at these temperatures."
* * *
15. , Radiant Floor Heating. Plasco Manufacturing Ltd., Janca Enterprises Ltd., March 1985.
An instruction manual produced by a Canadian (German) manufacturer for hydronic floor radiant heating systems. Gives some design procedure details as well as installation details.
* * *
16. Sartain, E. L. and Harris, W. S., "Performance of Covered Hot Water Floor Panels - Part I - Thermal Characteristics", Heating. Piping & air Conditioning. October, 1955.
The following is a summary of the results obtained for the test conditions investigated:
1. Apparent thermal resistance of the bare concrete panel was about 1.05 (F deg) per Btuh (sq ft).
2. Thermal resistance of the combinations of carpeting and pad ranged from 0.40 (F deg) per Btuh (sq ft) for the rubber pad alone to 1.87 (F deg) per Btuh (sq ft) for the heavy carpet and 40 oz jute pad.
3. The thermal resistance of both the asphalt time and the rubber tile
B-60
was about 0.05 (£ deg) per fituh (sq. ft).
4. Floor coverings, such as asphalt tile or rubber tile, which have a thermal resistance of 0.2 (f deg) per Btuh (sq ft) or less had a neglibible effect on the performance of floor panel systems.
5. At design conditions, covering the floor panels with any type of carpeting had pronounced effects on the water temperatures, reverse loss from the panel, and the required boiler size (see Table 4).
6. Covering a floor panel with carpeting did not appreciably increase the seasonal fuel consumption.
7. Because of the large increase in water temperature required when a carpet is applied to floor panels it may be impossible to balance floor panel systems in which carpeting is used in some rooms only, unless the piping is arranged to permit zoning with the use of more than one water temperature.
8. Major effects of carpeting over a bare floor panel on the design and performance of a floor panel system are shown in Table 4.
* * *
17. Sartain, E. L., and Harris, W. S., "Performance of Covered Hot Water Floor Panels - Part II - Room Conditions", Heating. Piping & Air Conditioning. November, 1956.
1. At design conditions of 80 F indoor-outdoor temperature difference the maximum difference in room-air temperature between the levels 3 in. above the floor and 3 in. below the ceiling was 3.5 F. At a location 2 ft from the north wall, the room-air temperature was 3 F lower than at the center of the room. Cooled air dropped to the floor at the north wall and window and moved in the direction of the south wall with a somewhat high velocity.
2. Addition of floor coverings to bare floor panels reduced the ability of the system to maintain a constant room-air temperature. The greater the thermal resistance of the floor covering, the greater the resulting room-air temperature variation. Carpets and pads retarded the flow of heat from the water to the room-air, resulting in poor response.
3. Glass surface' temperatures measured with floor panel heating were the same as those obtained in the Research Home with conventional radiation.
4. The exposed wall surface temperature was about 8 F lower and the AUST was about 4 F lower than the room-air temperature measured at the center of the room 30 in. above the floor.
5. The AST in Rooms A and B was about 1 F below the room-air temperature while in Rooms C and D the AST was essentially the same as the room-air temperature.
B-61
The MRT as obtained in the center of Room A with a thermo-integrator was the same as the AST for all outdoor conditions encountered.
The effects of the carpeting were to cause an increase in the floor surface temperature along a line toward the center of the room and to smooth out the heat flow profile from the panel to the carpet.
At design conditions of 80 F indoor-outdoor temperature difference the measured panel output was from 7 to 18 percent greater than the calculated panel output. For a given panel minus room-air temperature difference the panel heat output to Rooms C and D, which had more severe exposures, were 15 to 20 percent greater than outputs in Rooms A and B.
The relative humidity in Room D which had a carpet and pad was consistently greater than that in Room B, indicating that there was probably an increase in the rate of transfer of water vapor through the concrete floor slab in room.
* * *
Sartain, E. L. , and Harris, W. S., "Heat Flow Characteristics of Hot Water Floor Panels", Heating. Piping & Air Conditioning. January 1954.
The following is a summary of the results obtained for the test conditions investigated.
1. The air temperatures at the center of the rooms were very uniform, with a variation between the floor and 60-in. level of 0.5 deg. The temperatures of the air 3 in. below the ceiling and 3 in. above the floor were practically the same.
2. The measured heat flow from panel to room ranged from 87 percent of the calculated above floor heat loss in Room A to 101 percent of the calculated above floor loss in Room C.
3. It was found that the fuel savings resulting from the use of insulation under the entire floor slab as compared to the use of edge insulation only was too small to warrant the additional cost.
4. Vertical insulation along the inside edge of the foundation wall was as effective as the L type edge insulation. The savings in material and the ease of installation made the vertical insulation the more desirable of the two types.
5. At design conditions, the reverse loss from panels with edge insulation amounted to 20 to 23 percent of the total panel output. The reverse loss was roughly twice as great as the estimated heat loss through unheated floor slabs using the heat transmission values given in THE GUIDE 1953.
6. Comparisons of the reverse loss from the heated slabs, with figures presented in Chapter 12 of THE GUIDE 1953 for heat losses from
B-62
unheated floor slabs, indicate that at an indoor-outdoor temperature difference of 80 F, the fuel consumption when using floor panels, would exceed the fuel consumption when using radiators or convectors by about 10 percent.
7. It was found that while the mean floor surface temperature was uniform across the panel, the heat emission rate was much higher near the exposed wall and window than at the center of the room. Thus, the floor panel system had the desirable characteristic of automatically increasing the heat output rate in areas adjacent to points of high heat loss from the room. For a given floor surface temperature and room-air temperature, one could expect a somewhat higher output per square foot of panel area in an unisulated room or one with large glass area than in a fully insulated room with limited glass area.
* * *
19. Schutrum, L. F. , Parmelee, G. V., Humphreys, C. M. , "Heat Exchanges in a Floor Panel Heated Room", Heating. Piping and Air Conditioning. July 1953.
The authors qualified their conclusions to the fact that these were laboratory tests. Their observations were as follows:
1. Pending the establishment of data on the radiative and convective components of the total heat output from floor panels, studies of which are currently under way, the values for total panel output obtained in these tests indicate that the values given in THE GUIDE 1952 (Fig. 11, p. 548) are of the right order of magnitude for unheated mean radiant temperature (UMRT) values of around 65 F.
2. Room air temperatures were found to be higher for a floor panel heated room than for a ceiling panel heated room for the same room surface temperatures. For example, with an 85 F panel temperature and a 70 F AUST, with no infiltration, the room air temperature at the 60 in. level would be 74.8 F with a floor panel, and 71.2 F with a ceiling panel.
3. The effect on the room air temperature at the 60-in. level of varying the amount of infiltration air and its entering temperature was about the same in the floor panel tests (see Fig. 10 this paper) as it was shown to be in the ceiling panel tests (see Fig. 9, Reference 1).
4. In none of the tests reported on here did the air temperature gradient in the room between the 2-in. and the 90-in. level exceed 3.5 F deg.
B-63
20. Shamsundar, N., Lienhard, J. H. and Tezduyar, T. E., "Performance of Polybutylene Pipe in Concrete Heating Panels", Report No. 2, Department of Mechanical Engineering, University of Houston, Houston, Texas, 77004, 1985. (unpublished)
The summary and conclusions from this UNPUBLISHED article are the following.
They have shown that the ASHRAE procedure is not only erroneous, but results in inadequate designs. They have shown how this procedure can be corrected to increase its accuracy considerably, and how the modified procedure can be used to design the amount of insulation needed. They have developed computer programs to obtain accurate results and recommend that they be used in critical cases to check the adequacy of the design.
One aspect they did not check was the effect of the spatial oscillation in the surface temperature. The earlier research performed by ASHAVE showed that this needs to be considered only when the tube pitch was larger than twice the slab thickness.
* * *
21. , "Underfloor Radiant System Uses 86*i Supply Water", Air Conditioning. Heating and Refrigeration News. Oct. 21, 1985.
A description of a panel water distribution system for use in floors. This would replace the commonly used plastic or metal tubes. Provides a description of floor panel heating systems and their advantages and disadvantages.
* * *
22. VanGerpen, J. H. and Shapiro, H. N. , "Analysis of Slab-Heated Buildings", ASHRAE Trans., V. 91, Pt. 2, 1985.
Analyses of the performance of slab-heated buildings have been presented. On the basis of the simulations studies, some general conclusions can be stated regarding the feasibility of slab heating in comparison with other direct heating systems.
1. Slab systems share the advantage of other radiant heating methods in that the air temperature can be lower than for convective-type systems.
2. Slab heating actually requires more input energy than conventional heating systems to maintain the same air temperature, due to edge and bottom heat losses.
3. An important design trade-off exists in selecting the depth of the mats. Placing the mats close to the surface allows for quick response to changing loads, but it has the disadvantage of allowing the inside air temperature to drop rapidly when the mats are turned off. Alternatively, deeper mats cannot respond quickly but are needed to provide enough energy storage in order to use off-peak electric power
B-64
exclusively.
4. The control of a slab-heating system is difficult because of the phase shift between when the energy is put into the mats and when it is recovered at the slab surface. This can lead to chronic under - and overheating.
5. The main advantage of slab heating is in its ability to use lower cost off-peak power. However, in a given case, a complete life-cycle cost analysis would be required to assess the feasibility of slab heating.
B-65
G. PANEL HEATING & COOLING
1. Alexander, J. C., "Calculations of Direct Energy Losses from Ceiling Mounted Radiant Heating Panels to Fenestrated Areas", Energy Engineering: Journal of Association of Energy Engineers, Vol. 78, No. 5, pg. 35-48, August-September 1981.
The proportion of energy radiated from a ceiling mounted radiant heating panel directly to a window is calculated, so that the resulting energy losses can be estimated in the design of a radiant space heating system. Both first (geometric) and second (taking into account the optics of glass) order numerical radiometric calculations are done. Results are given for 2 x 8 ft and 4 x 8 ft radiant ceiling panels at 175oF located centrally to the window or off-set from the center for various (5 x 6 ft to 8 x 20 ft) window sizes with the panel located horizontally from the window from 0 to 10 ft. Up to 30 percent of the panel energy can be lost through the glass.
2. Baker, Merl, "Removal of Internal Radiation by Cooling Panels", Heating. Piping & Air Conditioning. November 1949.
Internal radiation emitted from electrical lighting filaments may be efficiently removed from a conditioned space by cooling panels. Because of the high temperature of incandescent filaments, the receiving panel or sink may be operated at a relatively high temperature which may be obtained by use of unchilled water. Fortunately, for most light colored painted surfaces, the visible portion of the radiation spectrum is highly reflected while the infrared, constituting the bulk of filament radiation, is extensively absorbed.
The results of research conducted by the author show that a maximum of approximately 65 percent of the electrical input to a large incandescent filament may be removed directly from the conditioned space. Specially designed reflectors may be used to concentrate the radiation on particular surfaces for removal.
* * *
3. Baker, Merl, "Effectiveness and Temperature Requirements for Cooling Panels Removing Internal Radiation", Heating. Piping and Air Conditioning. June 1952.
This paper presents a simplified theoretical analysis of the removal of internal radiation by use of cooling panels. Equations are developed enabling the determination of the effectiveness of enclosed radiant energy sources in heating the room air, together with expressions for the cooling load requirements. From the established relationships, both the panel cooling and the convective load components may be calculated for any given panel temperature, or the required panel temperature may be computed in accordance with the comfort equation. The theoretical analysis is supplemented by examples and design charts.
B-66
Internal radiation emitted from electric lighting filaments may constitute a major portion of the total cooling load. For structures not equipped with cooling panels, the effectiveness of this energy in heating the room air varies with the equivalent overall conductance of the enclosure and ranges approximately from 96 to 68 percent for good and poor insulation, respectively. By use of cooling panels possessing a conventional finish, these values decrease appreciably ranging from approximately 62 to 48 percent. A further decrease to approximately 52 to 45 percent is accomplished by use of a heat-absorbing panel surface.
* * *
4. Becker, Sidney, "Surface Temperatures of Plaster Ceiling Panels", Heating. Piping & Air Conditioning. April 1950.
The surface temperatures of a plaster panel containing 3/4 in. pipes on 7-in. centers, with water circulating at 119 F, were obtained at equilibrium conditions for four arrangements of insulation at the back of a panel. Imbedding of the pipes in plaster as compared with exposure of pipes on the back of a panel reduced temperature variation at the surface by 10 deg. Temperature differences between various points on the panel surface varied from 26-12 deg, depending on the insulation used behind the panel, the heat lag in the coolest part of the panel surface does not exceed 20 to 25 minutes.
•* * *
*
5. Boyer, L. L. , "Radiant Panel Effects of Floor - Ceiling Assemblies Incorporating Static, Return, and Supply Penums", ASHRAE Trans., Vol. 74, Pt. 2, 1968.
Vertical heat transfer factors for heat removal light troffers acting within a full scale floor-ceiling sandwich system have been determined. At a flow rate of 70 cfm per unit, approximately 65% of the heat from the lighting system is extracted, regardless of conditioning cycle or plenum function. On the cooling cycle about 15% is directed downward while on the heating cycle about 25% flows into the room. The balance in each case flows into the plenum to be distributed by the remaining elements of the floor-ceiling sandwich.
The overall radiant panel effects of the floor-ceiling sandwich have been evaluated. All radiant panel aspects have been combined and compared to the air distribution component of the total conditioning effect. Only with the supply plenum configuration is the net radiant panel aspect compatible with the conditioning system on both cooling and heating cycles. On the heating cycle, about 90% of the conditioning with a supply plenum is due to the radiant environment.
The supply plenum condition provides the most favorable radiant environment for occupant comfort, as indicated by mean radiant temperature, since on the cooling cycle the coolest MRT is obtained and on the heating cycle the warmest MRT is obtained. In neither of these two cases, does the MRT differ from room air temperature by more than 1.5 F at the room center.
B-67
6. Howarth, E. S., Huddleston, S. C , and Koch, R. M., "Aluminum Ceiling Panels for Heating and Cooling", Heating. Piping & Air Conditioning. September 1951.
Results of heating and cooling tests conducted on a variety of brazed aluminum ceiling panels are presented fora number of panel and room operating conditions. An analytical solution concerning the thermal performance of metal heating and cooling panels is developed and this is reduced to a design chart for easy applications to panels differing markedly from those tested. Comparisons drawn between the actual test results and those predicted by the analytical solution indicate good agreement. It can be concluded that reasonably accurate predictions of the thermal performance of a wide variety of aluminum ceiling-type heating and cooling panels (including types in which aluminum panels are fastened to tubes by mechanical means such as clips) can be made through the use of the design chart or equations and properly selected values of all contribution factors.
* * *
7. Humphreys, C. M. franks, C. V., Schutrum, L.-F., "Laboratory Studies of the Thermal Characteristics of Plaster Panels", Heating. Piping and Air Conditioning. July 1951.
This paper presents the results of laboratory studies on four plaster panels, using both non-ferrous tube ferrous pipe, located both above and below the lath. The panels represented conventional types of ceiling panels.
It is shown that the difference between the average tube temperature and the average panel surface temperature can be related to the heat output from the panel by a simple empirical equation. Effective conductance values are given for the different types of panels for tube spacings of 4 to 12 in.
Relationships between total panel output, tube spacing and the amplitude of the temperature wave on the panel surface are shown in a series of curves. Data and curves show the effect of back insulation on panel performance and some information is given on the effect of back plastering.
* * *
8. Hutchinson, F. W. and Baker, Merl, "Optimum Panel Surface Distribution Determined from Human Shape Factors", Heating. Piping and Air Conditioning. June 1951.
This paper presents the results of an experimental investigation of the shape factor of the clothed human body with respect to energy emitted by floor areas. The results complement those of an earlier paper in which similar shape factors were reported for the human body (in standing and in sitting position) with respect to energy emitted by wall and ceiling areas; the present study was limited to the standing position. Human shape factors for approximately 95 percent of the population will agree within plus or minus 7 percent with the data obtained from the dummy which represented an average man; for the majority of people agreement between actual and dummy shape factors will be found to be extremely close. Design charts are presented.
B-68
9. Irwin, R. R., "Panel Cooling for a Residence", Heating. Piping & Air Conditioning. May 1955.
Data obtained and presented in part in this paper seem to indicate that panel cooling can be used successfully to provide summer comfort in the residence tested. The study was carried out in a college apartment unit under actual living conditions. One of the unexpected results was the importance of the radiant cooling effect of the wall and ceiling panels which gave the occupants a comfortable feeling when discomfort might have been expected because of high humidity readings. The system was designed and operated as a year round unit, but this paper deals only with the cooling phase. The cooling of the water was done with an evaporative cooler.
* * *
10. Leopold, Charles S., "Design Factors in Panel and Air Cooling Systems", Heating. Piping and Air Conditioning. May, 1951.
The author has attempted to present the theory of panel and conventional air cooling systems and to indicate possible courses of panel cooling design.
In comparing air conditioning methods, it is essential that the methods under comparison shall not produce an end result which will unduly compromise with the production of optimum conditions.
Assuming that the air conditioning methods to be compared are capable of attaining the same end result, the selection of a particular form of air conditioning is a matter of economics. The air conditioning design should be related to all elements of building construction and use, and the economics be determined not solely on the owning and operating cost of the air conditioning but on the owning and operating cost of the entire building.
* * *
11. Leopold, C. S., "The Mechanism of Heat Transfer Panel Cooling Heat Storage", Refrigeration Engineering. July 1947.
The author points out that designing for panel cooling is not the simple reverse of panel heating. Neither the concept of combined surface conductance, nor the grey body are adequate for the cooling analysis. Phenomena of heat storage and panel cooling are discussed in terms of the mechanism of energy transfer source to atmosphere and enclosure. Test data are presented for a continuous cooled ceiling for various types and sizes of luminaires, with and without supplemental air supply, and for two types of ceiling finish.
* * *
12. Leopold, C. S. , "The Mechanism of Heat Transfer Panel Cooling Heat Storage Part II, Solar Radiation, Refrigerating Engineering. June, 1948.
In this paper, the author's analyses is extended to the solar load. Test
B-69
data are presented for the performance of panels and the overall heat balance for an enclosure with a continuous cooled celling, one window, and various combinations of glass and shading devices. Several basic solar load properties which are now incorporated into the ASHRAE design procedure are brought out here.
* * *
13. Lorenzi, R. J., and Schreiber, J. F., "Performance of an Electrical System of Panel Heating with Four Stages of Insulation", Heating. Piping & Air Conditioning. January, 1949.
A four-room one-floor, occupied residence, completely heated by panels of electrically conductive rubber, was operated through the 1947-48 heating season. Operating results and cost data are reported for the original construction and for three conditions in which the heat loss was progressively reduced by the following steps: (1) change to double glazing of windows, (2) adding 2 in. of mineral wool insulation to ceiling, and (3) addition of 2 in. mineral wool insulation to floor. Comments are made on comfort reaction of occupants and on economic justification of expenditure for reduction of heat loss of the structure.
* * *
14. Mills, C. A, "Reflective Radiant Conditioning Can Provide More Comfort at Less Cost", Refrigerating Engineering. Jan., 1955.
A discussion of the benefits of radiant heating and cooling panels for residential applications. Points out the benefits of reducing air motion and the bringing in of outside air.
* * *
15. Mills, Clarence A., "Residential Cooling by Reflective Radiation", Refrigerating Engineering. Vol. 58, No. 11, November, 1950.
This article reports on the design and first test results of cooling and heating by radiant cooling and heating panels on the ceiling. It presents a description of a single residence set up for radiant cooling and heating.
* * *
16. Mills, C. A., "Sensible vs Latent Heat Removal in Radiant Cooling", Refrigerating Engineering. March 1958.
t
A discussion of radiant cooling in industrial and commercial applications where high latent loads are present. Mostly descriptive in nature.
B-70
17. Mills, Clarence A., "Year-round Residential Conditioning By Reflective Radiation", Refrigeration Engineering. November 1950.
This article is a discussion of radiant cooling and heating for residential applications. Many claims are made but not a great deal of data is given.
* * *
18. Naumov, A. L., Solovyov, A. J., Shilkrot, E. 0., "Heating of Industrial Buildings with the Help of Suspended Radiating Panels", Clima 2000 - Heating. Ventilating and Air Conditioning Systems. Vol. 6, W S Kongres - W S Messe, 1985.
Radiant heating ensures a high level of comfort and low heat consumption. Theoretical and experimental research works being conducted in the Soviet Union gave opportunities to elaborate an effective structure of radiating panels (convective heat transfer from 1 kg of metal equal to 65 watts; a share of radiant heat transfer is about 0.63). A method is suggested for the design of radiant heating systems which make possible the determination of the necessary amount of radiating panels, and a scheme of their location in the premises. Effective field of application of radiant heating systems with suspended radiating panels taking account of the shape of the premises, value of infiltration heat exchange, moisture content of the premises, has been presented in the paper.
* * *
19'. Nottage, H. B. , Franks, C. V., Hulbert, L. E., Schutrum, L. F., "Heat Flow Analysis in Panel Heating or Cooling Sections", Heating. Piping and Air Conditioning. May 1953.
The heat flow has been studied for the case of a row of pipes or tubes imbedded in a slab and tangent to one surface, with a solid conducting medium adjacent to this surface, and under conditions such that the opposite slab surface may be taken as isothermal. Results are presented in the form of the thermal resistance between the pipes and the opposite isothermal slab surface. A general mathematical solution was obtained and electrolytic analogue was employed for establishing application data because of the complexity of the mathematical treatment. A floor slab on the earth is the intended panel heating application.
* * *
20. Olivieri, J. B. and Singh, T., "A Computer Program for Radiant Cooling of High Bay Buildings", ASHRAE Trans.. V. 93, Pt. 1, 1987.
In 1982, the authors reported results of research on the effect radiant cooling panels had on stratification in high bay buildings. That study showed that energy use could be reduced by as much as 40% using radiant cooling panels.
B-71
In this latest study, a computerized calculation method suitable for use on a personal computer is presented. The calculation method used finite differences to predict the amount of heat transferred through the roof and into the floor slab.
21. , "Radiant Cooling Panel will Get Tryout in U.S.", Air Conditioning. Heating and Refrigeration News. October 21, 1985.
A description of a radiant panel cooling system that will be marketed in the U.S. Contains only descriptive information.
22. Schutrum, L. F. , Parmelee, G. V., Humphreys, C. M., "Heat Exchanges in a Ceiling Panel Heated Room", Heating. Piping and Air Conditioning. December, 1952.
For this room and the specific test conditions reported here the following observations can be made:
1. The total heat output of the ceiling panel was much lower than is given in some presently used design methods. This seems to be due principally to the fact that the convection conductances obtained for the. ceiling panel appear to be much lower than presently published data would indicate.
2. The heat flow due to radiant exchange between a given surface and the rest of the room may be opposite in direction to the heat flow due to the convective exchange between the surface and the ambient air.
3. The surface temperatures of neutral walls were not necessarily the same as the temperature of the ambient air but were dependent upon the heat balance between the radiative and convective heat exchanges.
4. For this room, a definite relationship was established between room surface temperatures, infiltration air rate and temperature, and the temperature of the room air at the 60 in. level.
5. With no infiltration the AST minus the room air temperature was directly proportional to the ceiling temperature minus the AUST. For constant values of ceiling temperature and AUST, the difference between the AST and the room air temperature increased as the rate of infiltration increased and as the temperature of the infiltration air decreased. As defined before, the AST is the area-weighted average temperature of all the room surfaces, and the AUST is the area-weighted average temperature of the unheated surfaces of the room.
B-72
23. Schutrum, L. F. and Humphreys, C M . , "Further Studies of the Thermal Characteristics of Plaster Panels", Heating. Piping and Air Conditioning. June 1953.
The authors' conclusions are as follows:
1. The heat transfer within a plaster panel is related to the heat output from the lower and upper surfaces of the panel and can be expressed in terms of an effective conductance. This is defined as the downward heat flow in Btu per (hour) (square foot) divided by the difference between the average tube temperature and the average panel surface temperature in Fahrenheit degrees.
2. For all practical purposes, panel surface temperatures may be used for calculation of the upward heat flow.
3. Insulation on the back of plaster panels serves the following purposes: (a) reduces upward heat flow: (b) increases the average •panel surface temperature for a given tube temperature or conversely: (c) for a given downward heat flow, permits operation with a lower tube temperature than would be required if the panel were not insulated.
4. The addition of back-plastering to panels constructed with tubes above metal lath increases the heat transfer to the panel surface. In general, the heat transfer of a panel with tubes on 6-in. centers without back-plastering is equivalent to that of a panel with tubes on 8-in. centers with back-plastering.
5. It has been demonstrated that for panels having tubes above metal lath, good tube embedment and good contact between tubes and lath are prerequisite to good heat transfer.
24. Schutrum, L. F. and Humphreys, C. M. , "Effects of Non-Uniformity and Furnishings on Panel Heating Performance", Heating. Piping and Air Conditioning. February, 1954.
Tests from which the following conclusions are drawn were madein the Environment Laboratory. However, these tests were made under such a variety of conditions that it seems reasonable to assume that they may be applied with satisfactory accuracy to any ordinary structure.
1. The performance of a panel heating system in a space having a non-uniform surface temperature environment can be predicted with satisfactory accuracy on the basis of the area weighted average unheated surface temperature (AUST) of the space.
2. Furnishings in a panel heated space tend to reduce the heat output of the panel, and increase the room air temperature. However, the net effect of these trends may be considered negligible for design purposes.
B-73
3. When a floor covering is laid over a heated floor panel, the surface temperature of that panel and the surface temperature of the heating medium must be considerably increased to maintain the same heat output to the space that would be obtained from the bare panel. For a heat output of 25 Btu per (hr) (.sq ft), the amount of this temperature increase was found to vary from 27 deg to 60 deg for the various combinations of carpets and pads tested.
4. Coverings for heated floor panels should be selected to provide a minimum of resistance to heat flow from the panel to the space.
* * *
25. Schutrum, L. F. and Vouris, J. D. , "Effects of Room Size and Non-Uniformity Non-Uniformity of Panel Temperature on Panel Performance", Heating. Pining and Air Conditioning. September, 1954.
From laboratory tests, the authors arrived at the following conclusions: 1. The effects of room" size on the performance of floor and ceiling
panels are relatively small so that the heat transfer relationships developed in the standard room and reported in References 2, 3 and 4, may safely be used without correction for the design of panel heating systems for any space of normal size and proportion.
2. Within the range of conditions tested, ceiling or floor panels comprised of heated and unheated sections have the same total heat output and produce the same room air temperature as if the entire area were heated to a uniform temperature equal to the area weighted average of the heated and unheated surfaces. The maximum test temperatures of the heated portions of floor and ceiling panels were 95 F and 140 F respectively.
* * *
26. Schutrum, L. F., Vouris, John, and Min, T. C , "Preliminary Studies of Heat Removal By a Cooled Ceiling Panel", Heating. Piping & Air Conditioning. Cleveland, Ohio, January, 1955.
The preliminary conclusions from this study were:
1. Ceiling panel cooling is an inversion of floor-panel heating, and the performance of a cool ceiling panel can be predicted from the performance of a warm floor panel.
2. An appreciable amount of sensible heat can be removed from a room by a cooled ceiling panel, surface temperature of which is above normal inside dew-point temperatures. In a typical test, a ceiling at 10 deg F below room temperature absorbed 20 Btu per (hr) (sq ft) of ceiling area.
3. Infiltration of warm air and internal, convective, heat sources must be considered for design purposes.
B-74
4. Non-uniform wall and floor temperatures can be represented by area-weighted average temperatures (AUST) in calculating heat pickup by a cooled ceiling.
5. Room furnishings can be neglected as they affect the heat pickup by only approximately 5 percent.
27. Schutrum, L. F. and Min, T. C , "Cold Wall Effects in a Ceiling -Panel Heated Room", Heating. Piping & Air Conditioning. Cleveland, Ohio, August, 1956.
#
The room-side temperature of a wall or all-glass wall exposed to winter outdoor conditions affects the air temperature, the air movement, and the radiant conditions within the room. In a test room heated by a ceiling panel, one whole wall was cooled to simulate the inside temperature of an exposed wall. The ceiling was heated in panels 4, 8, and 12 ft. wide and the panel temperatures selected to maintain a constant 70 F room-air temperature. Room-air temperatures, air velocities, and mean radiant temperatures were measured under steady-state conditions and are reported.
In the living space, variations in room-air temperatures were found to be small; and room air velocities were neglibible except near the floor. Mean radiant temperatures were in general higher than the room-air temperature.
28. Schutrum, L. F. , and Min, T. C , "Lighting and Cooled Air'Effects and Panel Cooling", Heating. Piping & Air Conditioning. Cleveland, Ohio, November, 1957.
The authors reached the following conclusions.
1. The heat pickup by a cooled ceiling-panel and conditioned air may be summarized as follows: (a) The sum of normal heat gain through the room surfaces plus the radiation from lights which falls on the walls and floor is removed in part by the panel and, in part, by the conditioned air, and the division may be determined by Fig. 3, or by Equation 3 of the paper; (b) The convection heat gain from the lights is removed almost entirely by the conditioned air system and may be added to the convected load as determined in (a); (c) the radiation from the lights to the panel can be added to the panel pickup as determined (a).
2. For the particular fixtures used in this study, approximately 10 percent of the energy supplied to the direct lighting system was radiated in the visible wavelengths, 32 percent was emitted as long-wave radiation, and 58 percent was transferred to the room air by convection. For indirect lighting, these values were 10 percent visible radiation, 21 percent long-wave radiation, and 69 percent convection.
B-75
3. Approximately 16 percent of the energy input to the direct fluorescent lighting fixtures was radiated to the cooled ceiling panel, and 24 percent to the walls and floor, for indirect lighting, these changed to 18 and 12 percent respectively.
4. More than 20 percent of the energy supplied to the lighting system can be removed by a cooled ceiling panel. This includes the heat radiated directly to the ceiling, and that which is reradiated from the other room surfaces.
* * *
29. Singh, T. and Olivieri, J. B. , "Effect of Radiant Cooling Panels on Temperature Stratification", ASHRAE Trans. . V. 88, Pt. 2, 1982.
The conclusions from this study were the following:
1. The use of radiant cooling panels produces a slightly greater degree of temperature stratification when compared to those produced by air systems at all loads.
2. The energy consumption for radiant cooling system is 8 percent less, compared to that for the most efficient air system (using combined supply and return air system).
3. The position of. the panel has no significant effect, either on the temperature stratification or the energy consumption.
4. The cooling load factor Fc decreases as the lighting load increases.
* * *
30. Singh, T. , and Olivieri, J. B., "Effect of Radiant Cooling Panels on Temperature Stratification under RP-260", Final Report, ASHRAE, August, 1981.
The conclusions of this ASHRAE project were listed as follows.
1. The use of radiant cooling panels produces a slightly greater degree of temperature stratification as compared to those produced by air systems at all loads.
2. The energy consumption for radiant cooling system is 8% less as compared to that for the most efficient air system (using combined supply and return air system).
3. The position of the panel has no significant effect either on the temperature stratification or the energy consumption.
4. The cooling load factor Fc decreases as the lighting load increases.
B-76
31. Subcommittee of TAC, Staff Members of the ASHRAE Research Laboratory, "Thermal Design of Warm Water Ceiling Panels", Heating. Piping & Air Conditioning. December, 1955.
The design procedure here presented is based on research and provides a reliable means for the thermal design of ceiling-type heating panels using warm water as the heating medium. With its simplicity, it retains engineering accuracy appropriate to the usual applications of panel heating in residential and commercial buildings.
B-77
The basis of this procedure is a body of experimental data obtained at the ASHRAE Research Laboratory in a comprehensive program planned and guided-by Technical Advisory Committee on Panel Heating and Cooling. This work has been reported in the 8 research papers listed in the Bibliography at the end of this paper.
Findings from these studies have been weighed as to importance, and then trimmed to meet the needs of a simple but accurate design procedure.
The simplified procedure provides a panel design to maintain the desired room air temperature for the selected outdoor conditions. Room air temperature is the selected criterion of comfort, and the design procedure is restricted to situations in which the area weighted average temperature of the walls, the floors, and glass does not differ greatly from room air temperature. The room-scale tests, which simulated various conditions of construction and outdoor temperature, showed that this near-equality of the two temperatures normally prevails.
* * *
32. Tasker, C. , Humphreys, CM., and Parmelee, G. V., "The ASHVE Environmental Laboratory", Heating. Piping and Air Conditioning. Vol. 24, March 1952.
The new Environment Laboratory, located in the ASHVE Research Laboratory, provides facilities for the study of panel heating performance and also for further study of human reactions to environment involving various combinations of temperatures of air and interior wall surfaces. Temperatures of all surfaces or portions of them may be varied at will to simulate field operating conditions. The ceiling height was also adjustable and ventilation was variable. The room could also be divided into smaller segments.
* * *
33. Weida, D. E. , P. E. , "Life-Cycle Cost Analysis of Hydronic Radiant Panel", ASHRAE Trans., Vol. 92, Pt. 1, 1986.
A convective heat transfer computer model was utilized to predict energy consumption of a radiant panel heating and cooling system. The panel system responded faster than convective systems and maintained a more uniform mean radiant temperature in the room. Therefore, a system load adjustment should be considered for the radiant panel model. Finally, the radiant panel heating and cooling system can be justifiable on a life cycle cost basis.
* * *
34. Zhang, Z. and Pate, M. B., "A Numerical Study of Heat Transfer in a Hydronic Radiant Ceiling Panel", Numberical Methods in Heat Heat Transfer. HTD-Vol. 62, ASME, New York, 1986.
The heat diffusion equation was used to model numerically a radiant-ceiling panel for both steady-state and transient heat transfer. Three different types of boundary conditions were required: isothermal, adiabatic, and a combination of radiation and convection heat transfer. Several different numberical solution schemes were investigated, and significant insights into
B-78
the advantages and disadvantages of each scheme were obtained by comparing the results. The explicit method was found to be the most effective method for solving both the unsteady-state heat diffusion equation and the steady-state Laplace equation. Specifically, the computation time was less for the explicit method as compared to the implicit method because of the radiation boundary conditions. The output of the transient model was the temperature history of the radiant panel ceiling including the final steady-state temperature distribution and the heat flux from the panel. The heat transfer characteristics of the heating panel as predicted by the transient numerical model are also discussed herein. Several design considerations are investigated using the numerical model, including tube spacing, plaster thickness, and convection heat transfer rate.
B-79
H. INFRARED HEATING
1. ASHRAE, "High In tens i ty Infrared Heaters - Chap. 30", 1983 Equipment Handbook. American Society of Heating, Refrigerating and Air Conditioning Engineers, Atlanta, GA, 1983.
The current chapter in the ASHRAE Handbook describing infrared heaters .
* * *
2. ASHRAE, "High Intensity Infrared Radiant Heating - Chap. 18", 1984 Systems Handbook. American Society of Heating, Refrigerating and Air Conditioning Engineers, Atlanta, GA, 1984.
The current chapter in the ASHRAE Handbook describing high intensity infrared radiant heating.
" • * * *
3. ASHRAE Task Group, "High-Intensity Infrared Heaters", ASHRAE Journal, December 1963.
This information was prepared by the ASHRAE Task Group on Radiant Space Heating, and is the basis of the information presented in the ASHRAE Chapters on Infrared Heaters (Chap. 30 in the 1983 Equipment Handbook) and Infrared Radiant Heating (Chap. 18 - 1984 Systems Handbook). It considers such items as applications, gas infrared energy generators, electric infrared energy generators, system efficiency, complete building heating, spot heating, wind or draft effects, system design principles, controls, and system precautions.
* * *
4. Baumanns, H., "Gas IR Heaters for Heating Large Spaces", Warme Gas International. Vol. 27, No. 4, Gewea GmbH and Co. Monchengladbach, April, 1978.
High temperature radiant heating in large spaces offers physiological advantages. Cool air and comfortable radiation reduce fatigue and the amount of particulates in the air. Because of the low temperature of the air, avoidance of temperature stratification and air motion there is a considerable saving (20%) in energy. (In German). Figures are given for determining the heat requirements for industrial buildings. Tables are given for gas infrared unit efficiencies.
B-80
5. Belsey, D. G. and Benseman, R. F., "Spatial Radiation Patterns for Infra-Red Heaters". Building Science. Vol. 3, 1969.
A emthod for evaluating the radiation intensity distribution is presented. An example case is given. When evaluating high intensity infra-red heaters, the data from a limited number of radiation measurements can be used to compute spatial radiation patterns for a range of mounting heights and tilts. The assumptions used to simplify the computation do not introduce significant errors and the method produces data of adequate accuracy for all practical purposes. Information such as this will assist designers of high intensity infra-red heating installations.
* • * *
6. Boyd, Robert, "Application and Selection of Electric Infrared Comfort Heters", ASHRAE Journal, October 1962.
This article points out that any space can be heated by infrared to improve comfort. Other means should be investigated thoroughly before deciding to use infrared. Where any other means is practical, infrared will not be the preferred method in most cases. Many places impractical to heat any other way can be heated effectively and economically by infrared. Many places where no practical means can provide real comfort can be made tolerable by infrared.
Authoritative definitions, nomenclature, recommended design procedures, and engineering data and standards are needed sorely in this field. Where infrared is indicated for comfort heating* electric infrared is advantageous because: (1) elements approach point or line sources, allowing excellent control of pattern (keeping radiation where it is wanted) . (2) Reliable controls permit balancing output with the heating requirements of the moment. (3) There are no air contaminants produced by the heating system.
* * *
7. Boyd, R. L., "What Do We Know About Infrared Comfort Heating?", Heating. Piping and Air Conditioning. Nov., 1960, p. 133.
This article presents a comprehensive review of what is known about infrared comfort heating by quartz lamps and tubes and metal sheathed heaters, and perhaps more importantly what is thought to be known i.e. those circumstances of position, movement, dress, wind conditions, and others, which cannot be so accurately measured but which can seriously affect the adequate operation of an infrared installation. A method of determining "watts input density" based on the amount of radiation delivered to a surface is also suggested.
B-81
8. Boyd, R. L. , "Control of Electric Infrared Energy Distribution", Electrical Engineering. Feb. 1963, p. 103.
The growth in popularity of infrared heating systems for personnel has been accompanied by overemphasis of some details and characteristics of specific elements. At the same time, other details and characteristics of elements, fixtures and systems, and their application have been neglected or overlooked.
Control of patterns of electric infrared radiation offers opportunity for material improvement in effectiveness for many applications. Although there is a shortage of available data, references given indicate that the nature and efficiency of the elements producing the infrared are not nearly as important in effectiveness of the system as is control of the pattern of the radiation produced. Some technical discussion of control of radiation patterns is given.
* * *
9. Bryan, W. L. , "Comparative Energy Requirements of Radiant Space Heating", ASHRAE Trans.. V. 87, Pt. 1, 1981.
This was primarily an analytical study to determine the fundamental heat transfer characteristics of a radiant heating system. Intrinsic to the analytical model, the heat loss from the building could be supplied by convection heat input directly to the space air or by radiation transfer directly to the internal surfaces of the building. Thus results could be obtained for the model for the extreme conditions of 100% radiant to 100% convection heat input. In these unrealistic extremes the MRT for 100% radiant heat input was 8 deg C greater than 100% convection heat input with the result that the space air temperature required for equal comfort is reduced 8 deg C. This one-to-one reduction results from an increase in radiant coefficient for the clothing surface of 4%. The inside surface temperatures of the space with the exception of the floor remained below space air temperature for both radiant and convection heat input. For radiant input, these surface temperature increases result in a heat loss greater than had been expected. Several assumptions (high surface emissivity, no room air temperature gradient, no radiant energy to occupants) make the results questionable.
* * *
10. Bryan, W. L., "Gas-Fired Radiant Heat", Mech. Engr. . Vol. 87, March 1965.
A comment on the validity of data in an article in Mechanical Engineering and AGA - Bulletin - 92. He objects to the use of GIR factor for rating radiant gas burners.
B-82
11. Buckley, N. A., P.E. and Seel T., "Gas-Fired Medium - Intensity Radiant Heating Provides a Cost-Effective, Efficient Space Conditioning Alternative", ASHRAE Trans., Vol. 92, pt. 1, No. 4, 1986.
Tube-type infrared medium-intensity radiant heating systems installed in four diverse applications demonstrated substantial energy savings. (40-70%). Converting energy savings to cost savings demonstrates the cost advantage that is-available with radiant heating. Medium-intensity radiant heating provides a cost-effective, energy-efficient alternative for space heating. All of the comparisons were made with steam radiators which can be "energy inefficient" if not properly maintained.
* * *
12. , "Conserving Energy with Infrared Heating", Plant Engineering. Vol. 31, No. 8, Pg. 155 +, Technical Pub. Co., April, 1977.
A general discussion of infrared heating* systems and how they affect comfort.
* * *
13. Cohn, Lisa, "Radiant Heating Units Net Big Savings in Special Cases", Energy User News, Vol. 7, No. 47, November 1982.
A news article concerning the advantages and disadvantages of gas infrared heaters. Focuses most on the economics of the systems and where they are most suited.
* * *
14. DeWerth, D. W. , "A Study of Infra-Red Energy Generated by Radiant Gas Burners", Research Bulletin No. 92. American Gas Association, Nov. 1962.
Radiation characteristics of atmospheric, powered and catalytic gas-fired infra-red generators, and those of different types of electric generators are developed and tabulated. These data show how much total normal energy is emitted by the different types of radiant heat sources at different temperatures and the spectral distribution of this energy in the infra-red spectrum from 1.4 to 16.0 microns. A Gas Infra-Red Radiation (GIR) factor was developed for gas burners, which is essentially a burner emissivity factor. This GIR factor takes flue gas radiation into account since it adds considerably to incandescent gas burner surface radiation.
Gas-fired burner total normal radiation values as high as 62,800 Btu per hour, square foot of burner surface were measured. The only electric infra-red generator which had a value this high was the quartz lamp. However, most of this lamp's energy was short wavelength energy (about '65 percent below 2 microns).
B-83
The only electric infra-red generator which emitted energy qualitatively equal to that of the gas-fired generators was the quartz tube, but its total normal radiation was somewhat lower than that measured for comparable gas equipment.
Brief data are presented to show the effect of special burner coatings on the burner total normal radiation and red brightness temperature.
* * *
15. DeWerth, D. W., "Gas-Fired Radiant Heat", Mechanical Engineering. Vol. 86, Nov., 1964.
Tha author describes the various types of gas infrared heaters, which have been developed since 1950. These include the radiant tube, the porous refractory burner, the direct fired refractory burner and the catalytic burner. The burner characteristics are presented including the spectral radiation curves and their typical applications.
* * *
16. DeWerth, D. W. , "Literature Review of Infrared Energy Produced with Gas Burners", Research Bulletin 83, American Gas Association Laboratories, Cleveland, Ohio, 1960.
The abstract of this work done in 1960 is reproduced here.
"An evaluation of the available literature pertaining to I-R energy production by both gas and competitive means is presented. Gas-fired I-R burners are described and evaluated, and compared to I-R energy generated by competitive means.
The literature search has indicated that the use of gas-fired I-R burners should become a field of ever increasing gas usage. Many current applications of I-R are described and an extensive list of possible future applications is presented.
The necessity for future research is indicated. Four subjects for future work are recommended. They are: (1) measure the emittance spectra of gas-fired I-R burners, (2) measure the absorption spectra of possible loads not covered by the literature, (3) work on problems of current applications and develop recommended new applications of gas-fired I-R burners, and (4) develop information on low temperature (below 750 F) and high temperature (above 2,900 F) I-R burners".
* * *
17. Diamant, R. M. E., "Radiant Heating", The Heating and Air Conditioning Journal. ASHRAE Journal, Vol. 50, No. 577, pg. 26-28, Troop Publ., February, 1980.
This article compares the radiant and convective heat output of various English infrared radiant heaters. They indicate that the ratio of radiant to convective heat output varied from 0.63 to 1.76 as the tube or surface temperature varied from 150oC to 500oC. Additional details on the performance
B-84
of these heaters is also given.
* * *
18. Faucett, J. W. , "Field Evaluation of High Intensity Infrared Space Heating Systems - Research Project RP-98", Final Report, ASHRAE, 1972.
This study program was designed to validate the conclusions of RP-41 by measuring in the field the performance of high intensity, electric and gas-fired infrared space heating systems in order to determine thermal comfort level, energy consumption, and installation and operating details. Data for the study came largely from actual field study and analysis. Data is presented but in the form in which it exists; it is difficult to evaluate. There were no general summaries presented.
* * *
19. Field, A. A. - "Direct Fired Radiant Heating Systems", Heating and Ventilating Engineer. Vol. 49, No. 571, pg. 6-9, Technitrade Journals Ltd., February 1975.
The author discusses three types of direct-fired radiant heating systems: (1) gas-fired infra-red panels, (2) gas-fired radiant pipes, and (3) air-heated radiant tube. Factors such as efficiency, energy consumption and cost are presented.
* • * *
20. Heath, George, "Basic Infrared Heating Applications", ASHRAE Journal, Dec. 1967.
A general discussion on the use of gas-fired infrared heater is given. The author discusses heat loss (reduced by 80-85%) energy requirements (reduced by 80-85%), operation, controls, applications, ventilation requirements, unit placement, and the need for evaluating condensation. A good discussion of the placement of radiant heaters is given.
* * *
21. Hunter, Robert K., "Controllable and Efficient Infrared Radiant Heating", Automation. October 1972.
The author presents descriptions of electric infra-red heating systems and their applications. The information is just general in nature.
B-85
22. Janssen, John E., "Field Evaluation High Temperature Infrared Space Heating Systems", ASHRAE Trans., Vol. 82, Pt. 1, 1976.
This paper is an analysis of the data collected in RP-98 by J. W. Faucett of York Research. The data were widely scattered due to conditions beyond the control of this test. With exception of the swimming pool, buildings were poorly insulated and subject to drafty conditions when there was a wind. Even so, people tended to indicate general comfort.
The operative temperature was an average of 5.2 deg F (2.9 deg C) higher than the dry-bulb temperature. It is estimated that this resulted in at least 20% reduction in fuel over convective heating systems.
It appeared that the estimates of physical activity were too high. Although certain tasks may have required 3.5 to 4.5 met, this probably was not sustained for any length of time. It is doubtful that the steady state activity exceeded about 2.5 met.
Radiant heat appears to offer distinct advantages, especially in industrial installations. However, care must be exercised in designing the system to assure uniform distribution of radiation to the occupied areas. These areas should be protected from excessive drafts caused by large open access doors if possible. Comfort is compromised when systems are manually operated or operated on an -intermittent basis. Intermittent operation tends to lead to complaints of dampness and cold floors. This result would not be peculiar to radiant systems but could be expected with convective systems also.
23. Morelli, A. J., "An Economic Study of an Electric Infrared Space Heating Installation", ASHRAE Journal, June 1968.
This form of heating for commercial and industrial type buildings with high ceilings has these advantages:
1. The installed heating capacity for electric infrared can be as much as 15% less than calculated heat loss for other types of heating systems. The data obtained actually showed up to one-third less than the calculated heat loss would indicate. However, it seems that an installed capacity less than 80 to 85% of calculated heat loss at the design outdoor temperature is risky at this time.
2. Infrared heating involves not only the air temperature in the space but gains due to the direct heating effect of the radiation as well. Therefore, in the same area there can be in effect two or more comfort temperatures called operative temperatures by installing different densities at different locations in the space.
3. Infrared heating has a lower roof heat loss because there is less stratification. In fact, tests showed temperatures at the ceiling several degrees lower than at waist level. In a conventional system the reverse would have been true. With infrared heating, the energy is used more effectively and efficiently because the heat is more confined to the lower level where it is needed.
B-86
24. , National Fuel Gas Code. ANSI Z 223.1, 1984, American Gas Association, 1984.
This code listed a short section concerning infrared heaters which is given below.
6.18 Infrared Heaters.
6.18.1 Support: Suspended type infrared heaters shall be safely and adequately fixed in position independent of gas and electric supply lines. Hangers and brackets shall be of noncombustible material. Heaters subject to vibration shall be provided with
vibration isolating hangers.
6.18.2 Clearance:
a. Listed heaters shall be installed with clearances from combustible material in accordance with their listing and the manufacturer's instructions.
b. Unlisted heaters shall be installed in accordance with clearances from combustible material acceptable to the authority having jurisdiction.
c. In locations used for the storage of combustible materials, signs shall be posted to specify the maximum permissible stacking height to maintain required clearances from the heater to the combustibles.
6.18.3 Combustion and Ventilation Air:
a. Where unvented infrared heaters are used, natural or mechanical means shall be provided to supply and exhaust at least 4 cfm per 1,000 Btu per hour input of installed heaters.
b. Exhaust openings for removing flue products shall be above the level of the heaters.
6.18.4 Installation in Commercial Garages and Aircraft Hangers: Overhead heaters installed in garages for more than 3 motor vehicles or in aircraft hangars shall be of a listed type and shall be installed in accordance with 5.1.10 and 5.1.11.
25. NEMA, Infrared Application Manual. Standard Publication No. HE 3-1983, National Electrical Manufacturers Association, Washington, DC, 1983.
This manual is the third edition of NEMA Standards Publication HE 3, first issued in 1971 and updated in 1976. It reflects much of the current technology involved in infrared heating for improvement of comfort level, and its purpose is to help potential users determine how infrared electric heating can be a viable method of meeting their space conditioning needs. As NEMA Authorized Engineering Information, HE 3-1983 addresses the infrared heater both as a unit and as part of an overall system involving other factors affecting comfort
B-87
level such as draft and humidity. The various applications of infrared heaters are described as well.
* * *
26. Pam, R. L. and Kesselring, J. P., Burner Survey for a High Efficiency Gas-Fired Heating Unit. Alzeta Report No. 84-706-104, Teledyne Laars,. January 1984.
A survey was undertaken to evaluate the suitability of a variety of burner types for use in a new concept, high efficiency residential hydronic heating unit. Emphasis was placed on radiant systems that are currently commercially available. Nine different burner types were reviewed and evaluated against criteria established by the hydronic heating unit manufacturer. Results of the survey show that a porous fiber matrix burner can most easily be incorporated into the heating unit design and meet the operational criteria.
* * *
27. Sanford, Len, "No Problem with Radiation", The Heating and Air Conditioning Journal. Vol. 54, No. 630, Pg. 14-18, 18, 21, Troup Publ., July-August 1984.
A discussion about the types of radiant heating systems, which are currently available, and how they compare with each other. A good discussion of radiant vs warm air as well as good points brought out about various types of radiant systems.
*• * *
28. Simmons, R. C., "Five Years Operation of an Industrial Infrared Heating System", ASHRAE Journal, June 1968.
Five years of operating experience with this installation has demonstrated the adequacy of the infrared system design for maintaining the design operational temperature at 50 F. Comparison of the electric infrared energy consumption with the calculated consumption of a conventional system, compensating for both the ambient air and radiant temperature effect upon occupant comfort, demonstrates substantial operating economies for the infrared system, particularly when the structure U value is high.
* * *
29. Taylor, F. M. H., "Radiant Space Heating", Building Materials. May/June, 1973.
The authors' conclusions were stated as follows.
1. There is a large potential field of expansion in the use of gas for commercial and industrial space heating.
B-88
2. By the use of radiant heating and localised burners, gas for heating at a premium price can compete effectively with oil.
Overhead radiant systems must have effective insulation or reflectors such that the radiant heat component is paramount, since convected heat rises to the roof space where a substantial proportion is wasted.
The concept of environmental temperature is accepted as a satisfactory measure of comfort temperature taking due regard to both air temperature and mean radiant temperature.
The utilisation of gas in the most efficient types of equipment will play a significant part in the reduction of atmospheric pollution as well as making a contribution to the conservation of national energy supplies.
* * *
Trewin, R. , Pate, M. , and Nelson, R. , "An Experimental Study of an Installed High Temperature Radiant Heater and Enclosure", ASHRAE Trans., Vol. 92, Pt. 1, No. 1, 1986.
The first conclusion to be drawn from this study is that steady state for the tube is reached in approximately 12 minutes. This can be important for comfort after a large decrease in the temperature of the inside air and inside surfaces. Second, the reflector assembly causes most of the energy to radiate in a downward direction and, therefore, the variation in wall and floor temperatures depends on the orientation to the tube. For example, during the on-off cycle, the wall temperatures vary at levels below 3 m (10 ft) and the floor temperatures are highest directly below the radiant tube. Significant changes in the floor and wall temperatures may affect comfort. Third, the air is heated by the radiant tube, the reflector, and the floor beneath the radiant tube and not by the walls since they are cooler. Fourth, the temperature of the air is lower in the garage area heated with the radiant heater than in the warehouse area heated with a forced-air heating system. This difference is largest in the lower 4.5 m (15 ft). Finally, savings come from reduced infiltration losses, less air movement on the inside walls, and lower inside wall temperatures. Heat loss through the ceiling was not reduced significantly.
* >v *
31. Zawacki, T. S., Huang, V. and Macriss, R. A., "Development of a Standard Test Method for Measurement of the Radiant Heat Output of Gas-Fired Infrared Heaters", ASHRAE Trans.. V. 93, Pt. 1, 1987.
"The current American National Standards Institute (ANSI) standard (Z83.6-1982) method for determining total infrared heat output from commercial gas-fired infrared-tube space heaters is not sufficiently broad to encompass all of the equipment currently in the marketplace. For example, long straight or U-shaped tubular heaters cannot be evaluated under the spherical grid approach specified in the standard. The grid size would be simply too large for practical use.
4.
5.
B-89
An approach to extending the procedure to other heaters has been developed through a literature survey and a laboratory evaluation. Current test procedures were evaluated along with instrumentation used to measure infrared heat rates. A procedure was developed for measuring total radiant output based on the use of a 180° view angle radiometer and a cylindrical grid surrounding the heater. The procedure was tested in the laboratory and found to perform satisfactorily. A total radiant output of 40,500 Btu/h was measured emanating from a straight tube heater of 104,000 Btu/h gross input. The measuring procedure developed includes several parameters that were carefully assessed and defined, specifically, radiometer window material, the need for radiometer water cooling, calibration techniques, and measurement of background radiation."
B-90
I. DESIGN PROCEDURES
1. Adlam, T. N. , Radiant Heating. The Industial Press, New York, NY, 1947.
One of the original design texts for radiant heating systems. Discusses types of systems (ceiling, floor, wall, baseboard, and electrical) panels. Presents layout details for heating and cooling as well as snow melting. Much of the information is based on practical experience and some measurements that are given are limited to specific cases. Specific design steps are given but need to be updated with data accumulated over the last forty years.
* * *
2. ASHRAE, "Panel Heating and Cooling Systems - Chap. 8", 1984 Systems Handbook. American Society of Heating, Refrigerating and Air Conditioning Engineers, Atlanta, GA, 1984.
The current chapter in the ASHRAE Handbook describes panel heating and cooling systems.
* * *
•3. Buckley, N. A., "Engineering Principles Support an Adjustment Factor When Sizing Gas-Fired Low-Intensity Infrared Equipment", ASHRAE Trans. . V. 93, Pt. 1, 1987.
The conclusions from this study as given by the author are as follow.
"The demonstrated performance characteristics of low-intensity radia,nt heating equipment support an overall reduction in the input energy requirement for radiant heating system installations. Factors that contribute to this conclusion are:
+ Floor temperatures are elevated above ambient by low-intensity gas-fired radiant heating systems without heating the air.
+ The floor heat reservoir reradiates and convects heat into the space, increasing comfort and improving temperature recovery capability.
+ Improved mean radiant temperature is evidenced by the positive responses of a globe thermometer to the radiant field and, therefore, equal comfort conditions are maintained by radiant systems with lowered thermostat settings.
+ Reduced air temperature stratification with low-intensity gas-fired radiant heating systems reduced structural heat loss significantly.
+ Demonstrated operational efficiencies of low-intensity radiant systems established that space comfort control can be maintained with a lower energy input.
B-91
These factors support the application of an adjustment factor to standard heat loss calculations. Manufacturers of radiant equipment use adjustment factors of .80 to .85 in design calculations. These values are based on 20 years' experience and have been applied successfully in thousand of installations ."
* * *
4. Chapot, Jean Robert, Manual Techniques "Avadis". Utilization Chauffage Par Le Sol, Pont-A-Mousson, S. A.
This is a design manual for selecting and installing radiant floor heating systems using using plastic pipe. It is in French and appears to be published by technical society in France. It would be useful to have translated if one is going to be interested In installing embedded pipe floor radiant systems.
* * *
5. Correa, Edward L., Design Guidelines for Heating Aircraft Hangars with Radiant Heaters. Naval Civil Engineering Lab, December, 1983.
The results of this investigation indicate that radiant heaters are practical heaters for use in large open bay buildings. Generally, radiant heaters surpass convective forced-air counterparts in heating large open bays in the following ways: (1) by providing increased thermal comfort at the floor level while substantially reducing heating costs and heat stagnation, (2) by being able to heat objects to just above the dew point temperature to prevent condensation and corrosion, (3) by allowing heating flexibility with zone or whole building heating.
Gas fired, high-intensity, porous, refractory, IR radiant heaters are recommended for use in aircraft hangars if natural gas is available because of the following: (1) the porous, refractory, IR burners emit heat energy primarily in the longer wavelengths (2 to 6 microns) which is within the optimum absorptance range for personnel and concrete floors and has no adverse physiological effects, (2) porous, refractory heaters are safe for use in aircraft hangars, (3) when the burners glow a dull red, a malfunctioning burner would be visually apparent by intermittent burner incandescence.
The following disadvantages should be noted:
1. Due to the inherent nature of suspending these heaters near the ceiling of an aircraft hangar, access to these heating units is restricted for inspections and maintenance.
2. Natural gas may not be available at the hangar location.
3. The gas-fired heaters require ventilation for elimination of the flue gases. ,
B-92
6. , Gas Engineers Handbook. Industrial Press, NY, 1977.
This handbook of fuel gas engineering practices discusses the application of gas fired infrared radiant heaters. They make several recommendations concerning the application and installation of these types of heaters. They recommend a minimum roof vent opening of 50 sq. in. per 100,000 Btuh for unvented overhead heaters. They also discuss total plant heating, area heating and spot heating. This source indicates that the design heat loss can be reduced by 15 percent if the perimeter heating method is used. If this is not done or if the products of combustion are immediately exhausted this percentage reduction does not hold true. The reasons given for this 15 percent reduction are: (1) very little temperature stratification, (2) no air movement required for heat distribution, (3) the gas unit efficiency with unvented combustion is about 90% rather than 80% as with vented convection heaters, and (4) lower air temperatures for comfort conditions can be maintained with radiant heaters.
* * *
7. Gluck, B., Strahlungsheizung Theorie und Praxis. Verlag C. F. Muller, Karlsruhe, Germany, 1982.
Radiant Heating - Theory and Practice is a German design book on radiant heating systems. It contains the following chapters: 1) Fundamentals of Heat Transfer, 2) Emission of Heat from Large Panels with Imbedded Tubes, 3) Heat Emission from Segments of Surfaces and Radiant Panels, 4) Heat Emission from Infrared Rays, 5) Multiple Irradiation, 6) Mechanisms of physical Heating in Radiant Heated Rooms, 7) Heating Requirements and Location of Heated Surfaces, 8) Design and Construction of Heating Elements, 9) Control Procedures, 10) System Design, 11) Special Radiant Heaters, and 12) Comprehensive Examples. It is not available in English.
* * *
8. Hutchinson, F. W., A Graphical Design Procedure for Radiant Panel Heating. Revere Copper & Brass, Inc., New York, NY, 1948.
This work is based on the rational heat balance concept and procedure developed in a series of technical papers written by Hutchinson. The method is simple and is stated to give results with accuracy equal to the analytical procedure. The method can be applied to ceiling, floor or wall panels. It also allows calculation of design load for the building.
* * *
9. Hutchinson, F. W., Design of Heating and Ventilating Systems. The Industrial Press, New York, 1955.
This is a textbook for HVAC System design. It does have a chapter on panel heating design procedure. The method presented here is similar to that given in a book published by Revere Copper and Brass Co.
B-93
10. IHVE, Estimation of Plant Capacity. A9, IHVE Guide, 1975.
This British guide for designing heating ventilating and air conditioning systems contains some design information for radiant types of heating systems. An environmental temperature is defined, which is based on a combination of the mean radiant temperature and the air temperature. This combination is based on the resultant temperature being in the range for human comfort. Therefore, as adjustments are made in the air temperature, this will also affect the design heat loss calculations. This is particularly true with factories, where air and mean radiant temperature may differ appreciably.
Factors which cause this adjustment are: improved insulation, radiant heating, reduction in the infiltration rate, and acceptance of a lower temperature for comfort.
In this heat loss calculation, a uniform temperature throughout the height of the heated space is assumed. Certain modes of heating cause vertical temperature gradients which lead to larger losses, particularly through the roof.
Percentages to be added to the calculated heat loss to allow for these temperature gradients are given below.
Method and type of heating "
Radiant, warm floor Radiant, warm ceiling
Medium and high temperature radiant units from high levels
Forced warm air convective system with cross flow at low level
Forced warm air convective system with downward flow from high level
Medium and high temperature cross radiant from intermediate level
Percent to be added for the following heights of heated space
16 ft
0 0
0
0-5
0-5
0
16 to 32 ft.
0 0-5
0
5-15
5-10
0-5
>32ft.
0
0-5
15-30
10-20
5-10
* * *
11. Prince, Fred J., "Infrared Heating for Overall Comfort", ASHRAE Journal, Dec. 1968.
A design procedure is presented and some useful information is given for infrared heaters. Useful design charts are presented for heat delivery from heaters at specific locations, building heat loss correction factors, CO2 dilution air requirements, surface temperature charts, etc.
B-94
12. Prince, F. J., "Selection and Application of Overhead Gas-Fired Infrared Heating Devices", ASHRAE Journal, October, 1962.
A general description of the selection and application procedures for infrared heaters. His conclusions are as follow: "With the growing popularity of gas infrared heat, engineers, architects and the heating industry need standards for evaluating units and for system design.
In the interim, it is suggested the heating engineer evaluate carefully and demand proof of all performance claims regarding infrared units before he makes his selection. Having assured himself of the validity of such claims, he should apply sound engineering principles in the design of any infrared system.
Since applications have become so varied, a paper of this length cannot cover all or even a small part of present practices. Reputable manufacturers will have engineering and application manuals which should be studied. With the many thousands of successful installations, previous history on almost any type of application is available and the experience gained should be utilized".
* * *
13. Raber, B. F. and Hutchinson, F. W., Panel Heating and Cooling Analyses. John Wiley & Sons, NY, 1947.
A textbook describing the analysis and design of panel heating systems. It gives descriptions, advantages and disadvantages of panel systems. Chapters included are: radiation equations, comfort relationship, mean radiant temperature, evaluation of shape factors and heat balance equations.
* * *
14. Subcommittee of TAC, "Thermal Design of Warm Water Concrete Floor Panels", ASHRAE Research Laboratory, Trans., ASHRAE, V. 63, 1957.
This paper presents a simplified procedure for the thermal design of water heated floor panels for use in residences and commercial buildings. It complements the previously published paper.
The procedures in both papers are based primarily on the experimental data obtained at the ASHRAE Research Laboratory under the guidance of the ASHRAE Technical Advisory Committee on Panel Heating and Cooling. This work has been reported in a series of research papers which are listed in the references, together with other papers which contain supplementary data.
A panel designed by this procedure will maintain the desired room air temperature for the selected outdoor conditions. Room air temperature is the selected criterion of comfort, and the design procedure is restricted to situations in which the area-weighted average temperature of the walls, the ceiling, and glass does not differ greatly from room air temperature. The room-scale tests, which simulated various conditions of construction and outdoor temperature, showed that this near-equality of the 2 temperatures normally prevails.
B-95
15. , Thermo Lutz. Planungsunterlage Fur Ingenieure, Thermolutz GMBH and Co. Heizungstechnik Kg.
This is a design procedure presented by a German manufacturer for hydronic floor heating systems. The pamphlet is in German and presents design and installation details.
B-96
J. ENERGY CONSUMPTION
1. Ashley, J. L., Correa, E. and Canfield, K., "Energy Conservation: Heating Navy Hangars", Technical Report R-910, Naval Civil Engineering Laboratory, July 1984.
How energy is used for hangar heating and what methods are used to reduce hangar thermal energy consumption were investigated. The results of measurements of hangar air infiltration and stratification, two major causes of heating related energy consumption, are reported. Methods to reduce this type of energy consumption (reduction of air infiltration and installation of destratifiers, vehicle access doors, door seals, vinyl strip doors, and radiant heating) were evaluated and are discussed. Design criteria providing hangar air infiltration rates versus hangar size and climatic conditions and design criteria for hangar destratifiers were developed and are presented.
* * *
2. Bailey, H. R., "An Experimental Comparison of Energy Requirements for Space Heating with Radiant and Convective Systems", ASHRAE Trans.. V. 86, Pt. 1, 1980.
The purpose of this project was to compare energy consumption for radiant and convective heating units. Two buildings on the Rose-Hulman campus were selected -- one is heated with gas and the other with electricity. They were originally both heated convective systems. In the fall of 1976 radiant systems were installed in both buildings. Energy consumption and temperature data were recorded daily for two heating seasons with each building heated about a 1/2 yr. with a radiant heat and 1/2 yr. with convective heat. One reason for a 2-yr. test was to be able to have each system operate for both halfs of the heating season. For example, with the electric system radiant heat was used during the first part of the first year and the last part of the second year.
A measure of heating effectiveness is obtained by comparing daily energy consumption with AT where At is the inside temperature minus the average outside temperature for the day. An overall comparison for the 2-yr period was made by dividing the total energy consumption by the sum of all the AT's.
The convective systems for both gas and electricity used about 15% more energy/AT than the corresponding radiant system.
* * *
3. , "Fuel Bills Halfed After Switch to Infrared Heating System", Energy Management Technology. Vol. 7, No. 8, pg. 44-45, Walker-Davis Publications, November 1983.
Discusses the conversion from a forced-air oil-fired heating system to an infrared heating system in an auto dealership. They indicate about a 40% savings in fuel costs and a payback of 1 1/2 to 2 years.
B-97
4. __, "Gas Radiant Heating Installation Achieves 90% Overall Efficiency", The Heating and Air Conditioning Journal. Vol. 48, No. S60, pg. 60-61, Troup Publ., September 1978.
Discusses the installation and operation of a gas radiant heating unit which achieved 90% operating efficiency and saved approximately 50% on energy requirements for the structure.
* * *
5. Grum, R. E., "Building Heat with Natural Gas Infrared", ASHRAE Journal, June 1968.
This article presents some heating cost data for particular industrial buildings for various years of operation. This authors analysis shows that 15% reduction in unit size as well as energy consumption were present. Another building experienced an 10% reduction in fuel usage. It was also claimed that the heating systems provided greater comfort for the workers.
* * *
6. Guillaume, M. , "Integration of Different Energy Saving Possibilities in Dwellings", Energy Savings in Buildings. Commission of the European Communities, Edited by Ehringer H. and Zito V., November 1983, D. Reidel Publ. Co.
The overall efficiency of two heating plants was measured in two well insulated and identical houses. The differences of these heating installations are: for the first one radiant floor heating using a high efficiency and low water temperature boiler was used; and for the second one direct electrical heating with convectors was used.
In the first house (heated with a radiant floor) a reduction of 9% of the energy consumption of the boiler was made by using a closed circuit instead of an open one, and a reduction of 11% of the energy consumption was obtained with a night set back from 10 pm to 6 am. The results also show that the values obtained for the electrical heating system are higher than those for the classical system even when this one used low water temperatures.
* * *
7. , "In Old Building - New Infrared Heating System Solves Old Problems", Air Conditioning Heating and Refrigeration News. April, 1985.
Discussion of a specific building which went through a retrofit and used a gas infra-red heating system to replace a hot water heating system. They reported a payback of one year and are expecting up to 80% savings in energy costs.
B-98
8. Trewin, R. R., Langdon, F. M., Nelson, R. M. and Pate, M. B., "An Experimental Study of a Multipurpose Commercial Building with Three Different Heating Systems", ASHRAE Trans.. V. 93, Pt. 1, 1987.
"A commercial building with three distinct zones, each having different heating equipment, was monitored for two heating seasons using a computerized data acquisition system. An analysis of the thermal performance of both the building envelope and the heating equipment was performed. The office area was heated by a heat pump with an auxiliary backup furnace; the warehouse area used off-peak thermal-electric storage units; the garage area heating load was met by a gas-fired radiant system. The microcomputer-based data acquisition system obtained and stored hourly temperatures, humidities, and energy flows. The zone loads and equipment energy consumption and performance were calculated. Variable-base degree-days were computed for each zone, and it was observed that standard-based (65 F) degree-days could be in error by a factor of two."
"It was observed that each of the three spaces had a different base temperature to be used in the calculation of degree-days and that the use of standard degree-days based on 65 F could lead to erros in heating load estimates. The results for the office space were in good agreement with the65 F based degree-days as expected. The warehouse had few internal gains, resulting in the degree-day base temperature being close to the thermostat setting. The base temperature in the garage was significantly lower than the thermostat setting because a radiant-type heating system was used.
A data base has been established that will be used in future studies to verify the loads obtained from computer programs. The verified programs will then be used to assess the economic impact of various energy conservation measurements".
B-99
K. TRANSIENT EFFECTS
1. Aiulfi, D., Fort K., Ottin, T., "Modelization of Floor Heating and Oil Furnaces for the Unilization of Microprocessors in DDC", Clima 2000 - Heating. Ventilating and Air-Conditioning Systems. Vol. 6, W S Kongres - W S Messe, 1985.
Transient Models of floor heating, oil furnaces and rooms are developed for the application of microprocessors for online control strategies. The discussed models of floor heating and oil furnace as well as the overall model of a heated space have proven to be sufficiently accurate and simple. The importance of the radiation heat transfer for the accuracy of transient models is recognized. Finite differences are used for solving the conduction equation. Only a small quantity of results are given and no equations are presented.
* * *
2. Algren, A. B., and Ciscel, Ben, "Heating Panel Time Response Study", Heating. Piping & Air Conditioning. March, 1949.
A floor panel heating system was tested for transient response. A continuous record was made of water, ground, panel, air and surface temperatures resulting from a sudden and large change in the heat input to the panel system. Use a 4" concrete slab laid on 9 in. of crushed rock. Room had well insulated walls, no windows and a single door. The floor had 1" OD pipe, 4" in below the surface laid on 12" centers. The air temperature reached 50% of its change in 4 hours and 90% in 9 hours. The MRT was approximately 3oF higher than the air temperature for the majority of the time. Four hours were required for the panel to reach 63% of its change. There was an hour and a quarter lag between the air temperature and the panel surface temperature at the 63 percent value. The maximum air temperature rise was 8oF per hour and occurred about an hour after the heat input. Floor coverings would delay the air change rate. The heat input rates were high (5 to 6 times steady state loss) in order to account for thermal storage.
* * *
3. Berglund, L. , Rascati, R., Markel, M., "Radiant Assisted Comfort Heating for Energy Conservation in Intermittently Occupied Spaces", Energy conservation in the Built Environment: Proceedings of the CIB W67 Third International Symposium, Dublin, Ireland, 1982, pg. 98.
In winter, energy consumption can be reduced if intermittently occupied spaces are kept at a low, energy conserving temperature when unoccupied and raised to a comfortable level when occupied. The occupant's response and acceptance of such a plan were investigated. Subjects from a comfortable area at 22oC entered a space at 15oC and occupied it for 2 hours. Fast acting radiant heaters were activated when the subjects entered the cold space bringing the mean radiant temperature to a high level and the operative
B-100
temperature or the temperature that the environment feels like to 22oC. The air temperature increased at 3oC/h during the exposure to simulate the response of a conventional convective heating system. The radiant heaters were regulated by operative and air temperature controllers. Subjective responses of thermal sensation, degree of comfort and thermal acceptability were gathered periodically during the tests. The 16 subjects judged the environment of the radiantly heated system for intermittent occupancy to be thermally acceptable. The radiant system controlled by operative temperature was more acceptable and more energy efficient than the air temperature controlled radiant system because it produced less overheating. Spot and heated ceiling type radiant systems were tested.
* * *
4. Berglund, L., Rascati, R. and Markel, M. L., "Radiant Heating and Control for Comfort during Transient Conditions", ASHRAE Trans.. V. 88, Pt. 2, 1982.
Comparison tests are described that show people will accept spaces being cool upon entry if the spaces can be brought quickly to a comfortable level with radiant heat. Subjects from a comfortable area at 22oC (72oF) entered a space at 15oF) and occupied it for two hours. After the subjects' entry, spot radiant or fast-acting radiant ceiling panels rapidly raised the operative temperature of the space to 22oC. A sensor that averaged air and mean radiant temperatures was found to be superior to an air temperature sensor as input to the radiant heat controller. It produced less operative temperature overshoot and greater occupant thermal acceptability and reduced power consumption. There are numerous applications for energy savings with fast radiant systems, particularly where there is intermittent occupancy. The technique and control are also applicable to steady state situations. The savings depend on the application and can be predicted by calculation from the response characteristics of these tests.
* * *
5. Borresen, B. A., "Thermal Room Models for Control Analysis", ASHRAE Trans.. V. 87, Pt. 2, 1981.
The analysis of a dynamic control loop often requires the use of a room model. This paper discusses four simplified dynamic room models which in different ways take into account the thermal interaction between room air and surrounding walls. The room air is assumed to be fully mixed.
It is shown that the choice of the simplification level employed depends on how closely the long-term responses and steady-state values are to fit the actual room response. For modeling short-term dynamic responses, a simple time constant corresponding to the air change rate of the room is usually adequate and will lead to choosing conservative control parameters.
An experimental procedure for determining typical parameter values is discussed.
B-1GL
6. Boyar, R. E., "Room Temperature Dynamics of Radiant Ceiling and Air Conditioning Comfort Systems", Trans. ASHRAE, V. 69, 1963.
This article disusses the response time for water carrying radiant panel systems. The author compared the response time (time to reach 90% of its terminal value) of this system to that of a forced air system, and found that they were comparable. The dynamic response is more than adequate for perimeter areas of buildings, which have a relatively high percentage of glass.
* * *
7. Mclntyre, D. A., "Eight Hour Floor Warming: A Feasibility Study", Electricity Council Research Centre, Capenhurst, England, March, 1977 (NTIS-PB 277 115).
This brief note looks at the effect of improved building insulation on the temperature variation of a building heated by off peak underfloor heating. It seems that with good insulation an eight hour charge period can produce an acceptable temperature variation throughout the day. Increasing the thickness of screed over the heating cable is very beneficial. A carpet increases the downward loss substantially and its use reduces the effectiveness of floor warming. Eight-hour floor warming appears to merit further study. The practical problems of increasing speed thickness, cost effectiveness of underfloor insulation and the specification of control systems need attention, as well as extending the design rpocedures to cover eight hour operation.
* * *
8. Pfafflin, J. R., "Space Heating Dynamics", IEEE Trans, on Indust. Appls.. Vol. IA-19, No. 5, Sept/Oct, 1983.
Verification of models previously advanced for description of the active and passive modes of baseboard heating and forced air convection is given. Results of tests conducted under rigidly controlled conditions at the former Electric Space conditioning Institute are shown to support the proposed models. Sequences of heating and cooling reponses for the two means of energy input are evaluated by means of incremental forms of the fundamental equations. It is found that the dynamic responses are functions of the means of energy input.
* * *
9. Zhang, Z. and Pate, M. B. , "An Experimental Study of the Transient Response of a Radiant Panel Ceiling and Enclosure", ASHRAE Trans.. V. 92, Pt. 2, 1986.
The transient response of a radiant heating system and enclosure was investigated for a range of hot-water supply temperatures and flow rates. The radiant heating system consisted of copper tubes embedded in a standard plaster ceiling at 6-in (152-mm) intervals. Transient experiments were
B-102
performed by heating the radiant ceiling and enclosure from a cooled-down condition by using a step change in the hot-water supply temperature. Temperature transients in the water supply and return lines on the ceiling and wall surfaces and in the room air were then monitored for a period of several hours. Results were as follows: the ceiling temperature was uniform; the thermal response of the ceiling and enclosure was slow because of the large thermal mass in the ceiling; the air temperature did not lag the wall and floor temperature; and the room walls were heated by a combination of radiant heat transfer from the ceiling and convection heat transfer from the air. In addition, the transient response of the radiant system was found to be a function of water supply temperatures but not of water flow rate.
B-103
L. INSTRUMENTS
1. Benzinger, Theodore H., Maglum, B. W., and Hill, James, "The Design Construction and Operation of a Scanning Radiometer for Measurement of Plane Radiant Temperature in Buildings", ASHRAE Trans., Vol. 82, Pt. 2, 1976.
A description of an instrument is given along with its design details and its initial performance evaluation.
* * *
2. Braun, D. L. and McNall, P. E., Jr., "A Radiometer for Environmental Applications", ASHRAE Trans.. V. 75, Pt. 1, 1969.
A convection-nulling radiometer which involves the use of a thermoelectric module is described. It appears to have some advantages over other known instruments in convenience, ease of use, and ability to record the results. Its accuracy is felt to be "at least as good as other devices in use. It is hoped that others will further develop the device and that its use will further the research on the effects of radiation on the occupants of controlled space.
The radiometer also has advantages in the determination of the Effective Radiant Field (ERF) proposed by Gagge.
* * *
3. Korsgaard, V., "A New Radiometer Measuring Directional Mean Radiant Temperatures", Heating. Piping and Air Conditioning, -July, 1949.
A special radiometer described was developed to measure directional mean radiant temperatures. The instrument has been successfully vised in a test room with highly reflective walls. The basic principles of the radiometer and its construction are very simple, and by further investigations it should be possible to develop a complete theory for the thermal sensitivity of the radiometer and hence make its construction still more sensitive. The calibration of the radiometer is simple. Experience up to this time has shown that the radiometer is very handy and durable and further requires only a few minutes to reach equilibrium.
* * *
4. Madsen, T. L., "Thermal Comfort Measurements", ASHRAE Trans.. Vol. 82, Part I, 1976, pp. 67-70.
The author describes a comfort meter which he has developed. His conclusions were as follows:
1. The new comfort meter provides a quick, direct measurement of the predicted mean vote in a given space.
B-104
2. Comparison with calculated PMV values based on separate measurements of the thermal parameters in typical environments shows good agreement.
3. In cases where man is exposed to asymmetric radiation, the comfort meter gives a better approximation to the PMV value than can be calculated from traditional measurements of the thermal parameters.
4. Inaccuracy of a certain PMV value is due mainly to the fact that in practice it is difficult to state activity level and clothing with great accuracy. In order to compare different thermal environments, to measure the thermal effect produced by changes in the heating and ventilating system, as well as for the reproducibility of thermal environmental measurements in common, it is still important that the thermal parameters are measured accurately.
5. The comfort meter measures the thermal effect of ta, MRT, and v, simultaneously and at the same position; this gives a good reproduction especially under non-steady-state conditions.
* * *
5. Madsen, T. L. , "Definition and Measurement of Local Thermal Discomfort Parameters", ASHRAE Trans.. V. 86, Pt. 1, 1980.
The author indicates that there are no standard measuring methods for determination of the degree of thermal asymmetry or draft. Neither are there any measuring instruments on the market which are particularly adapted for these measurements. The measuring instrument described here is developed as an aid to fulfilling this need. Just as in the case of the comfort meter, the instrument has been constructed with the aim of simulating a person's heat exchange with the surroundings. While the comfort meter takes into consideration the heat loss of a person as a whole, and hence of the central temperature perception, the discomfort analyzer aims at simulating a person's peripheral temperature perception.
* * *
6. Tenney, A. S., Ill, "Red Hot and Hotter - Industrial Radiation Thermometry", Mechanical Engineering. Oct., 1986.
A discussion of the types of radiation thermometers which are available and their range or field of applications is given. Various industrial applications are discussed. A useful table is given for choosing the correct thermometer.
B-105
M. CONTROLS
1. Algren, A. B., Snyder, E. F. , Jr., Locke, J. S., "Field Studies of Floor Panel Control Systems", Heating. Piping and Air Conditioning. February, 1953.
The significant results of this two-year study on floor panels in two factory buildings and a residence indicate the importance of proper design and installation of the basic heating system and the importance of the selection and location of various control elements. It was pointed out that returns should not be combined prior to the three-way mixing valves used for recirculation. Proper demand heat rates for the boiler were also illustrated. The outdoor thermostat used for resetting the water temperature should be located such that it is exposed to the same climate conditions (wind and solar effects) as the structure. It was also determined that continuous circulation of the water was important in order to avoid surges in panel temperature. Best performance was obtained from systems with outdoor thermostats used for circulating water temperature reset. On-off and modulating valve action controls provided system stability, when large masses in the panel were available.
* * *
2. Algren, A. B. , Snyder, E. F., Jr., Head, R. R., "Field Studies of Floor Panel Control Systems - Part II", Heating. Piping and Air Conditioning. April, 1954.
This paper discusses results of field tests covering various control systems in three different types of construction. It recommends specific practices in both the operation and type of control system. The temperature difference across carpeting was lloF and floor tile was 2oF. The field studies indicate that the water temperature should be reset with the outside temperature. Large glass areas and solar loads increase the transient response and control problems. More of the heating load should be delivered to the perimeter of the system below windows. The vertical thermostat location did not appear to affect the results. These are field studies and it is difficult to extrapolate design data from these results.
* * *
3. Hazard, W. G. , "Radiant Heat Control in Industrial Plants", ASHRAE Journal, November, 1959.
Discusses the use of spot cooling, radiation shields and ventilation for controlling radiant loads in industrial applications. The main objective was to maintain comfort condition, for the workers. Discussions are also presented on instruments to be used for measuring radiant heat.
B-106
4. McNall, Preston E., Jr., "A Manufacturer's View of Radiant Heater Control", ASHRAE Trans., Vol. 81, Pt. 1, 1973.
The author's conclusions are as follows:
1. An ideal thermostat, which responds thermally in a manner similar to people, does not exist, not withstanding considerably R&D effort by the control industry.
2. Even if an ideal thermostat did exist, most spaces are subject to variations of comfort conditions with position in the spaces and load conditions, etc., so that no single ideal location can be found.
3. Most commonly used thermostats are sensitive enough to db and MRT to be suitable for use with comfort radiant heat systems, provided that the cycling rate is adequate and the droop is not too great.
4. For radiant heating sources with time constants greater than 5 minutes, a thermostat cycling rate of 6-10 cycles per hour is satisfactory to meet" the proposed ASHRAE unsteady-state comfort criteria. This includes all of the systems described here.
5. The electric ceiling cable system has a time constant of 40-50 minutes so that a cycling rate of 2-3 cycles per hour is adequate.
6. True proportional control (modulating thermostats) can be advantageous in reducing the effects of cycling and may have advantages in reducing demand charges, increasing efficiency, etc. However, the comfort standard can be met properly designed to position control.
7. When radiant sources are used in factories, loading docks, waiting platforms, etc., ordinary thermostats are not usually recommended. Here comfort is not as important as the alleviation of severe cold. Special techniques, such as "% - on" cyclers and large-droop thermostats or temperature transmitters are often warranted, but they are not within the scope of this paper.
* * *
5. Nakanishi, E. , Pereira, N. C , Fan, L. T. and Hwang, C. L. , "Simultaneous Control of Temperature and Humidity in a Confined Space - Part I, Building Science. Vol. 8, 1973, p. 39-49.
A pair of nonlinear differential equations which describe the transient behavior of temperature and humditiy in a confined space have been derived from simultaneous material balances of dry air and water along with the enthalpy balance of moist air. The equations are sufficiently general to take into account external heat loads, and internal heat and moisture loads within the confined space. Since temperature-humidity control systems allow usually only small deviations of temperature and humidity from a desired operating point, a linearization of the above equations within the bounds of small deviations is justified. Hence, the nonlinear equations are linearized around
B-107
a desired steady state operating point. On using available relations for the specific volume and enthalpy of moist air, the linearized equations further result in a pair of linear uncoupled differential equations. The response from the linear equations is found to compare very favorably with that from the original nonlinear equations.
* * *
6. Walker, C. A., "Control of High Intensity Infrared Heating", ASHRAE Journal, November, 1962.
This article presents a short description of various means of controlling infrared heating systems. He discusses the use of automatic controls vs manual and the use of electronic, and the thermostat control systems.
B-108
N. SPOT HEATING & COOLING
1. Best, W. H., "Spot Heating and Condensation Control Using Gas Infrared Systems", ASHRAE Journal, June, 1968.
This article describes the design procedure that might be used for spot radiant heters for comfort control and condensation control. It uses gas infra-red heaters for the specific units. It contains a useful design figure for the energy per hour per ft^ to be applied and a few other general rules of thumb.
* * *
2. Olesen, B. W. and Nielsen, R. , "Radiant Spot Cooling of Hot Working Places", ASHRAE Trans.. V. 87, Pt. 1, 1981.
Radiant spot cooling can improve the thermal conditions in warm working environments.
Radiant spot cooling decreases discomfort caused by warmth, but may create discomfort caused by radiant asymmetry. It is important to optimize panel positions according to the angle factor between worker and panels. Water condensation on the cooling panels and positioning of the panels may in practice limit the use of radiant spot cooling. The efficiency of radiant spot cooling is rather poor (10-15%). A good section on definitions is presented.
* * *
3. Sofrata, H. M., and Al-Hukail, Y., "Spot Cooling System Design", ASHRAE Jnl.. Jan., 1987.
An interactive computer program for spot cooling system design is discussed. Input includes ambient conditions for the industrial environment, the metabollic heat production and clothing value of workers in the target area, the jet and target area geometry, and the maximum and minimum of the conditioned air at the target area. Their results have been compared with available data in the literature and a good agreement has been achieved.
B-109
APPENDIX C
LISTING OF COMPUTER PROGRAM
C-l
LIST OF INPUT VARIABLES
CP - Presence (True) or absence (False) of ceiling panels HF - Presence (True) or absence (False) of a heated floor COOL - Presence (True) or absence (False) of cooling panels
TOUT ALTH BTH HT U HI
HIUP HIDOWN EPSI FP
P CFM ACH HREF SLOPE Q2L
RATIO AMI EFF CLO FCL V RH
Outside ambient temperature Length, breadth and height of the room U-factors for each room surface Standard convection coefficients (incl. radiation effect) for each room surface ASHRAE standard coefficient for upward and downward convection Emissivity of each room surface Person-to-room surface shape factors
Number of persons Supply air cfm/sq. ft of floor area Infiltration air changes per hour Reference height and gradient for the air temperature gradient Sensible heating load due to lights
Ratio of radiative to Dubois area for Metabolic rate and mechanical efficiency Clothing level and clothing factor Relative air velocity Relative humidity
a person
XPREF - Required temperature of panels NCP - Number of panels NPAL - Number of panels in a lengthwise row XCP,YCP- Coordinates of the center of the panel ALCP - Length and BCP - width of the panel EPSIP - Emissivity of panel surface XMULT - Multiplier for convection off the panels
XIN - Initial temperature for calculation PTOL - Tolerance and MAXIT - maximum number of iterations for panel temperature convergence
Note: A sample input data file is shown at the end of the computer program listing.
C-2
C******** FORTRAN CODE irk************* C TRIAL'3 , DEC 4 C C C C
IMPLICIT REAL*8(A-H,0-Z) DIMENSION WK(5000),X(10),PAR(10),F(10),X0(10),VAR(25) DIMENSION HIIN(6),OUT(25,25), OUTl(25,25) DIMENSION XCPIN(25),YCPIN(25),ALCPIN(25),BCPIN(25)
C CAN GO UPTO 15 PANELS WITH THIS EXTERNAL FCN
C C
LOGICAL CP,COOL,HF C
CHARACTER*21 TEMP(6) CHARACTER*50 TITLE CHARACTER*45 POUT(50), POUT1(50)
C COMMON /OUT/ TOUT COMMON /CEL/ XCEL(ll) COMMON /COMF/ AM1,EFF,AICL,FCL,HC,V,PA,RH COMMON /AIR1/ P,CFM,AICFM,ACH,UQ2P,Q2P,UQ2L,Q2L COMMON /Q/ Q1,Q2,QSTD3,QACT3,Q3,Q4,QP5,Q5,Q6,Q7,Q8 COMMON /QNET/ QNET1,QNETP2,QNET2,QNET3,OUA COMMON /QP/ QCVP,QRP COMMON /QI/ QR(6),QCV(6),QCD(6) COMMON /U/ U(6) ,HI(6),CI(6),EPSI(6) COMMON /UP/ UP,HIP,CIP,EPSIP,XMULT COMMON /CONV/ HIUP,HIDOWN COMMON /FSURF/ FS(6,6) COMMON /FPEOP/ FP(6) COMMON /GRAD/ HREF,SLOPE COMMON /TERM/ TERM6,TERM7,ALHS COMMON /DIMEN/ ALTH,BTH,HT COMMON /AREAS/ RAREA(6)
C /AREAS/ APPEARS IN MAIN,FCN AND SHAPE PROGRAMS. COMMON /PAN1/ CP,COOL,HF COMMON /PAN2/ PRAREA,PAREA COMMON /PAN3/ NCP COMMON /PAN4/ FSCP(25,25) COMMON /DIMP/ XCP(25),YCP(25),ALCP(25),BCP(25)
C C
DATA PAR/10*0.DO/ C C C
TEMP(1)='TEMP OF FLOOR ' TEMP(2)='TEMP OF CEILING * TEMP(3)='CLOTHING SURFACE TEMP1
TEMP(4)='ROOM AIR TEMP* TEMP(5)='MEAN RADIANT TEMP* TEMP(6)='SUPPLY AIR TEMP'
C TEMP(6) IS SUBSTITUTED BY PANEL TEMPERATURE FOR (CP) C C C
READ (5,*) XIN READ (5,*) TOUT
C-3
c c
READ READ READ READ READ READ READ READ READ READ READ READ READ READ READ
READ READ READ READ
(5,*) (5,*) (5,*) (5,*) (5,*) (5,*) (5,*) (5,*) (5,*) (5,*) (5,*) (5,*) (5,*) (5,*) (5,*)
(5,*) (5,*) (5,*) (5,*)
RATIO AM1.EFF AICL.FCL V RH P CFM,ACH HREF, SLOPE Q2L (U(K3),K3=1,6) (HIIN(K3),K3=1,6) HIUP.HIDOWN (EPSI(K3),K3=1,6) ALTH,BTH,HT (FP(K3),K3=1,6)
N.NSIG, ITMAX TITLE NVAR (VAR(I),I=1,NVAR)
C CP IS TRUE MEANS THAT CEILING PANELS ARE BEING USED C IF CP IS FALSE, THE PROGRAM ASSUMES CONVECTIVE HEATING/ COOLING
READ (5,*) CP READ (5,*) HF
IF (CP) THEN READ(5,*) COOL
C COOL IS TRUE MEANS THAT COOLING AND NOT HEATING IS BEING PERFORMED.
READ(5,*) XPREF READ(5,*) XMULT READ(5,*) PTOL ,MAXITP READ (5,*) HREF, SLOPE
C THESE WILL OVERRIDE THE VALUES READ-IN PREVIOUSLY . READ(5,*) NCP READ(5,*) NPAL READ(5,*) EPSIP DO 1002 J = 11,NCP+10
READ(5,*)XCPIN(J),YCPIN(J),ALCPIN(J),BCPIN(J) 1002 CONTINUE
ENDIF C ASDF
C C C CCCCCCCCCCCC CCCCCCCCCCCCC 3525 C C C C C C C C C C
DO 1550 IVAR = 1,NVAR ACH = VAR(IVAR)
READ (5,*) ALTH,BTH,HT READ (5,*) (FP(K3),K3=1,6)
DO 3525 J = 11,NCP+10 READ(5,*)XCPIN(J) ,YCPIN(J) ,ALCPIN(J) ,BCPIN(J)
CONTINUE READ(5,*) XPREF READ (5,*) (U(K3),K3=1,6)
READ(5,*) NCP READ(5,*) NPAL
C-4
!
C C C WRITING THE INPUT DATA... C WRITE (6,123) C WRITE (6,123) C WRITE (6,123) C WRITE (6,123) C WRITE (6,234) 123 FORMAT( //) 234 FORMAT(1X,T15,100('*')) C WRITE (6,77) 77 FORMAT (1X/,T5,'INPUT DATA',///) C WRITE(6,78)ALTH,BTH,HT 78 F0RMAT(1X/,T5,'R00M DIMENSIONS:',T25,'LENGTH = '.F6.2,
& T45,'BREADTH = *,F6.2,T65,'HEIGHT = *,F6.2) C WRITE (6,66)TOUT - 460.DO 66 FORMAT (1X/,T5,'OUTSIDE TEMPERATURE = ',F6.2,' DEG. F1) C WRITE (6,1)AM1,EFF 1 FORMAT (IX,/,T5,'METABOLIC RATE PER UNIT DUBOIS AREA = ',
&F5.1,' KCAL/HR.(SQ.M)*,10X,'EFFICIENCY= ',F4.2) C WRITE (6,2)AICL,FCL 2 FORMAT (IX,/,T5,'AICL = ',F4.1,' CLO',12X,'FCL = ',F5.2) C WRITE (6,3)V 3 FORMAT (1X,/,T5,'V = ',F6.2,' M/S') C WRITE (6,41)RH 41 FORMAT (1X,/,T5,'R.H. = \F4.2) C C WRITE (6,5)P 5 FORMAT (IX,/,T5,'NUMBER OF PERSONS =',F4.1) C WRITE(6,921) RATIO 921 FORMAT(1X/,T5,'RATIO, OF RADIATION AREA TO DUBOIS AREA = ',E15.8) C WRITE (6,8)UQ2L 8 FORMAT (IX,/,T5,'SENSIBLE HEAT (LIGHTS) BTU/HR =',E15.8) C WRITE (6,6)CFM,ACH 6 FORMAT (IX,/,T5,'SUPPLY AIR CFM PER SQ.FT = ',E15.8,10X,
&'TOTAL INFILTRATION AIR CHANGES PER HOUR = *,E15.8) C WRITE (6,1003)HREF,SL0PE 1003 FORMAT(IX,/,T5,'USING A GRADIENT FOR THE TEMPERATURE OF AIR',
&' AT DIFFERENT HEIGHTS :*,/,T5,'REFERENCE HEIGHT IN FT = ',F6.2, & 5X,'SLOPE (DEG. F PER FT) = ',F6.2 )
C WRITE(6,211)(K3,K3=1,6) C WRITE(6,21) (U(K3),K3=1,6) C WRITE(6,22) (HIIN(K3),K3=1,6) C WRITE(6,235) (EPSI(K3),K3=1,6) C WRITE(6,609) (FP(K3),K3=1,6) 211 FORMAT(IX,//,T5,'SURFACE (I) ',2X,6(I2,14X) ) 21 F0RMAT(1X,//,T5,'U(I) ',2X,6(E15.8,2X) ) 22 F0RMAT(1X,//,T5,'HI(I) ',2X,6(E15.8,2X) ) 23 F0RMAT(1X,//,T5,'CI(I). ' ,2X,6(E15.8,2X) ) 235 F0RMAT(1X,//,T5,'EPSI(I)',2X,6(E15.8,2X) ) 609 F0RMAT(1X,//,T5,'FP(I) *,2X,6(E15.8,2X) ) C C C C C C WRITE (6,31)N,NSIG,ITMAX C C C WRITE(6,123) C WRITE(6,123)
C-5
C WRITE(6,1022) 1022 F0RMAT(1H1,/)
WRITE(6,*) TITLE WRITE(6,123) WRITE (6,*)* CEILING PANELS ? = ', CP WRITE(6,123) WRITE (6,*)' HEATED FLOOR ? = *, HF WRITE(6,123) WRITE (6,*)' COOLING ? = ', COOL WRITE(6,123)
C CALCULATING CI FROM GIVEN U AND STANDARD HI C
DO 51 Kl=l,6 CI(K1)= l/( 1/U(K1) -1/HIIN(K1) )
51 CONTINUE C C C C INITIALISING THE UNKNOWNS ..
DO 11 Jl=l,10 11 X(J1)=0.0
X(l)= XIN C TEMPERATURE DISTRIBUTION INITIALLY GIVEN TO THE IMSL SUBROUTINE IS:
DO 20 11=2,10 X(I1)=X(I1-1)+1.D0
20 CONTINUE C C C C C FOR THE PANELS C WRITE(6,123)
C TEMP(6)='CEILING PANEL TEMP' TEMP(2)='REST OF CEILING TEMP '
C TEMP(6) IS SUBSTITUTED BY SUPPLY AIR TEMP. FOR .NOT.(CP) IF(HF) THEN
TEMP(1)='CEILING TEMP ' TEMP(2)='REST OF FLOOR TEMP ' TEMP(6)='HEATED FLOOR TEMP '
ENDIF
C ITERP = 1
C IF (CP) THEN
X(10) = XPREF ENDIF
C C THIS IS NOT USED RIGHT NOW, PANEL CONDUCTION BEING EXCLUDED FROM THE C ANALYSIS . C IF(CP) THEN, C CIP = l/( 1/UP -1/HIP ) C ENDIF C
WRITE(6,781) XPREF,XPREF-460.DO
C-6
781 FORMAT( 1X/.T5,'PANEL TEMPERATURE DESIRED = ', F7.2, &' BEG. R OR ',F7.2,* DEC F*) WRITE(6,782)PTOL,MAXITP
782 FORMAT(1X/.T5,'PARAMETERS FOR PANEL TEMPERATURE', &' ITERATION :',/,T45,'TOLERANCE = ',E15.8, &//,T45,'MAXIMUM NUMBER OF ITERATIONS = ',15 )
G WRITE(6,174) NCP
174 FORMAT(1X/,T5,'TOTAL NUMBER OF CEILING PANELS = ',13 ) WRITE(6,1023) NPAL
1023 FORMAT(1X/,T5,'NUMBER OF PANELS IN A LENGTHWISE ROW = ',13 ) WRITE(6,811)EPSIP
C605 FORMAT(1X/,T5,'OVERALL U FACTOR (UP) = *,E15.8,//, C & T5, 'STD. CONVECTION COEFF.(HIP) = ',E15.8,//, C & T5, 'EMISSIVITY (EPSIP) = ',F6.3,/) 811 FORMAT (//,T5,'EMISSIVITY OF PANELS (EPSIP) = ',F6.3,/)
WRITE(6,1511)XMULT 1511 FORMAT (//,T5,'MULTIPLIER FOR CONVECTION ,XMULT = ',F6.3,/) C C
WRITE(6,175) 175 FORMAT(1X/,T5,'PANEL CONFIGURATION :',//,T15,'PANEL NUMBER'
&T30,'COORDINATES OF CENTER',T75,'PANEL DIMENSIONS (INITIAL)f, &/.T30 'X (ALONG LENGTH) ', & T50, Y (ALONG BREADTH) ',T75,'LENGTH',T90,'BREADTH')
C C C INITIALISING THE PANEL GEOMETRY AND DIMENSIONS C
DO 3501 IPAN = 11, NCP+10 XCP(IPAN) = XCPIN(IPAN) YCP(IPAN) = YCPIN(IPAN) ALCP(IPAN) = ALCPIN(IPAN) BCP(IPAN) = BCPIN(IPAN)
3501 CONTINUE C C C C
DO 165 J =11,NCP+10 WRITE(6,176)J,XCP(J),YCP(J),ALCP(J),BCP(J)
176 FORMAT(1X/,T20,I3,T30,F6.2,T50,F6.2,T75,F6.2,T90,F6.2,//) 165 CONTINUE
C CALL SHAPE(ALTH,BTH,HT,FS)
C C C THE GOTO STATEMENT LEADS TO THE FOLLOWING 777 777 CONTINUE
TALCP = 0.D0 DO 1019 J = 11, NPAL + 10
1019 TALCP = TALCP + ALCP(J) UALTH = ALTH/ NPAL
C CALL SHCP(ALTH,BTH,HT,FS,NCP,XCP,YCP,ALCP,BCP,FSCP)
C C RESULTS ARE PRINTED AFTER THE XREF TEMP IS CONVERGED TO.L C
ELSE
C-7
I CALL SHAPE(ALTH,BTH,HT.FS)
C SHOULD LATER TRY TO PASS FS THROU1 THE SUBROUTINE..RATHER THAN C THROU* THE COMMON
WRITE (6,403) (K ,K=1,6) 403 FORMAT (1H1,/,T35,'WALL-TO-WALL SHAPE FACTORS',5X,
& 'WITHOUT CEILING PANELS ',//,T15,6(I3,9X) ) C
DO 404 1=1,6 WRITE (6,405)1,(FS(I,K) ,K=1,6)
405 FORMAT (1X/,T5,I3,T15,6(E10.4,2X) ) 404 CONTINUE C
ENDIF ,__1_^^__1_^1_I_1__J_1_1^^
C c c c c c c c 606 CONTINUE - • C INITIALISING THE UNKNOWNS .. C DO 11 Jl=l,10 Cll X(J1)=0.0 C X(l)= XIN C ABOVE IS NEEDED BECAUSE THIS LOOP IS EXECUTED MANY TIMES. C TEMPERATURE DISTRIBUTION INITIALLY GIVEN TO THE IMSL SUBROUTINE IS: C DO 20 11=2,10 C X(I1)=X(I1-1)+1.D0 C20 CONTINUE C IF (CP) THEN C X(10) = XPREF C ENDIF C C C C C C WRITE(6,123) C C***A*****A*S0LOTI0N OF THE NON- LINEAR EQUATIONS*********^^^ C
CALL ZSPOW (FCN,NSIG,N,ITMAX,PAR,X,FNORM,WK,IER)
C TO WRITE OUT THE F AT CONVERGENCE OF EACH EQUATION CALL FCN(X,F,N,PAR)
C IER IS WRITTEN LATER ON C C WRITE(6,123) C WRITE(6,123) C IF(ITERP.EQ.l) WRITE(6,*) 'TEMP. DISN- AFTER FIRST ITERATION', C & (X(JK),JK=1,10) C C WRITE(6,123) C WRITE(6,123)
C-8
/ /
C WRITE(6,123) C C
IF(IER.EQ.130 .OR. IER.EQ. 131) THEN WRITE(6,*)'IER = 129 : NUMBER OF CALLS TO FCN HAS EXCEEDED*,
& ' ITMAX*(N+1);MAY TRY NEW GUESS* WRITE(6,*)'IER = 130 : CANNOT GET ACCURACY (NSIG) REQUIRED * WRITE(6,*)*IER = 131 : MAY TRY NEW GUESS,ITERATION IS NOT',
& ' MAKING GOOD PROGRESS'
c #mmmmm# AN ABRUPT STOP mmmmmmm GO TO 999
ENDIF C
IF (CP) THEN IF ( DABS(X(10)-XPREF) .GT. PTOL .AND. ITERP.LE. MAXITP) THEN
C WRITE(6,*)* X(10) AT ',ITERP,' = *,X(10) DO 778 J = ll.NCP +10
IF(COOL) THEN C THIS IS FOR PANEL COOLING .INCREASE AREA IF COOLER THAN REFERENCE
ALCP(J) = ALCP(J)* XPREF/X(10) C WRITE(6,*),ALCP(',J,' ) AT ',ITERP,' = \ALCP(J)
ELSE C THIS IS FOR PANEL HEATING, INCREASE AREA IF HOTTER THAN REFERENCE
ALCP(J) = ALCP(J)* X(10)/XPREF . C WRITE(6,*)'ALCP(,,J,' ) AT ',ITERP,' = ',ALCP(J)
ENDIF 778 CONTINUE
ITERP = ITERP + 1 GO TO 777
ELSE C
IF(ITERP.GT.MAXITP) THEN WRITE(6,*)' NO CONVERGENCE IN',MAXITP,' ITERATIONS'
ELSE C WRITE(6,*)' CONVERGENCE IS OBTAINED AFTER ',ITERP, C &' ITERATIONS'
ENDIF C
ENDIF C C C =========================^^ c C WRITE(6,123) C WRITE (6,234)
WRITE (6,88) 88 FORMAT (1H1,T35,'OUTPUT DATA ',/)
WRITE (6,234) WRITE(6,123)
C***** PRINTING OUT RESULTS OF PANEL CASE, WRITE(6,*) ' ERROR PARAMETER OF IMSL ROUTINE,IER = ',IER WRITE(6,123) WRITE(6,1021)
1021 FORMAT(1X/,T5,'PANEL CONFIGURATION :',//,T15,'PANEL NUMBER', &T30,'COORDINATES OF CENTER',T75,'PANEL DIMENSIONS (FINAL) ', &/.T30 'X (ALONG LENGTH) *, & T50,fY (ALONG BREADTH) ',T75,'LENGTH',T90,'BREADTH')
DO 786 J =11,NCP+10 WRITE(6,176)J,XCP(J),YCP(J),ALCP(J),BCP(J)
786 CONTINUE
C-9
\
WRITE (6,785) PAREA 785 F0RMAT(1X///,T5,'TOTAL PANEL AREA =*,F10.2)
IF (ALCP(ll) .GE. UALTH .OR. TALCP.GE.ALTH) THEN WRITE(6,234) WRITE(6,1018)
1018 F0RMAT(1X///,T5,' COMPUTED LENGTH OF PANELS EXCEEDS ', &'THE LENGTH OF THE ROOM*,/,T5,'CHOOSE A GREATER VALUE OF', &' PANEL WIDTH ,(BCP), INITIALLY')
WRITE(6,234) C$$$$$$$$$$$ AN ABRUPT STOP $$$$$$$$$$$$$$$$$$$ C GO TO 999 C?$$$$$$$$$$$$$$$$$$$$$$$$$$$$$$$$$$$$$$$$$$$$$$$$
ENDIF WRITE(6,811)EPSIP WRITE(6,1511)XMULT WRITE(6.812)HIP
812 FORMAT (T5,'CONVECTION COEFF. USED FOR PANELS = ',E15.8) C C C C PRAREA IS NOW AVAILABLE THROUGH THE CALL SUB.
WRITE(6,207) PRAREA 207 FORMAT(1X/,T5,'REMAINING CEILING AREA.PRAREA, SQFT =',F10.2)
WRITE (6,192) (K ,K=1,6) 192 FORMAT (1H1,/,T35,'WALL-TO-WALL SHAPE FACTORS (FSCP) ',
&//,T15,6(I3,9X) )
DO 193 1=1,6 WRITE (6,194)I,(FSCP(I,K) ,K=1,6)
194 FORMAT (1X/,T5,I3,T15,6(E10.4,2X) ) 193 CONTINUE
WRITE(6,195) 195 FORMAT (1X//,T5,'PANEL-TO-WALL SHAPE FACTORS (FSCP) ',/)
DO 198 J = 11,NCP+10 WRITE(6,199) (J,K,FSCP(J,K), K=l,6)
199 F0RMAT(1X/,(T5,'FSCP(P,,I3,I3,') = ',E10.4,/) ) 198 CONTINUE
WRITE(6,413) 413 FORMAT (1X//,T5,'WALL-TO-PANEL SHAPE FACTORS (FSCP) ',/)
DO 201 K = 1,6 WRITE(6,412) (K,J,FSCP(K,J), J=11,NCP+10 )
412 FORMAT(1X/,(T5,*FSCP(',13,' ,P',I3,') = \E10.4,/) ) 201 CONTINUE C
WRITE (6,411) (K ,K=1,6) 411 FORMAT (1X/,T35,'WALL-TO-WALL SHAPE FACTORS *,5X,
& 'WITHOUT CEILING PANELS ',//,T15,6(I3,9X) ) C
DO 212 1=1,6 WRITE (6,213)1,(FS(I,K) ,K=1,6)
213 FORMAT (1X/,T5,I3,T15,6(E10.4,2X) ) 212 CONTINUE C C
ENDIF C THIS IS THE END OF THE IF(CP) C C CALCULATING OPERATIVE TEMPERATURE AND ERF C
C-10
C RATIO IS READ IN. C REFER ASHRAE(F)1985 , P8.4 (THE FOLLOWING ARE IN SI UNITS.
SIGSI = 5.67D-8 XCOMF = ( XCEL(8) + XCEL(9) )/2.D0 HRSI = 4.DO* SIGSI*RATIO*(273.D0 + XC0MF)**3.D0 HCSI = 8.5D0* (V**0.5D0) TOCEL = (HRSI*XCEL(9) + HCSI*XCEL(8) )/(HRSI+HCSI) TOF = TOCEL*1.8D0 + 32.DO ERFSI = HRSI*( XGEL(9) - XCEL(8) )
C 3.15 W/SQ.M = 1 BTU/HR.SQ.FT ERFFPS = ERFSI/3.15D0
C C C CALCULATING AVERAGE UNHEATED SURFACE TEMPERATURE, AUST
TOTAT = O.DO TRAREA = O.DO DO 1004 I = 1,5
TOTAT = TOTAT + RAREA(I)* X(I) TRAREA= TRAREA + RAREA(I)
1004 CONTINUE C
IF(CP) THEN TRAREA = TRAREA +.PRAREA TOTAT = TOTAT + PRAREA*X(6)
ELSE TRAREA = TRAREA + RAREA(6) TOTAT = TOTAT + RAREA(6)*X(6)
ENDIF C THIS WILL GIVE AUST IN RANKINE , AUSTC IN DEG. CELSIUS.
AUST = TOTAT / TRAREA AUSTC = (AUST-492.D0)/1.8D0
C C C C WRITE (6,234) C WRITE(6,123) C WRITE(6,123) C WRITE(6,123) C WRITE(6,123) C WRITE (6,234) C C C C WRITE(6,123) C C WRITING THE INITIAL TEMPERATURE CHOICE
WRITE(6,101)XIN -460.D0,FN0RM 101 F0RMAT(1H1,///,T5,'X(1) INITIALLY =',F6.2,T75,'FNORM = \E15.8)
WRITE(6,123) WRITE(6,*)' (NOTE: TEMPERATURES ARE CALCULATED IN THE PROGRAM IN &DEG. RANKINE )' WRITE(6,123)
C THE TEMPERATURES ON OUTPUT C TEMPS. 1 TO 4
DO 10 J=l,4 WRITE(6,100)J,J, X(J)-460.D0,XCEL(J),J,F(J)
100 FORMAT (IX,T5,'X(',12. ')=*,T15,'TEMP OF WALL1,I2,T40,'=', & E15.8,3X,',(,,F6.2,1X, C ) * ,5X, 'F(' ,12, *) =', & T88,E15.8,/)
10 CONTINUE C AND TEMPS. 5 TO 10
C-ll
DO 13 J5=5,10 WRITE(6,206)J5.TEMP(J5-4), X(J5)-460.D0,XCEL(J5),J5,F(J5)
206 FORMAT (1X,T5, X(',12. ') =',T15,A21,T40,'=', & E15.8,3X,*(',F6.2,1X, C )',5X,'F(*,12,') =', & T88,E15.8,/)
13 CONTINUE C
WRITE(6,123) WRITE(6,922) TOCEL.TOF WRITE(6,923) ERFSI3ERFFPS
922 FORMAT(1X/,T5,'OPERATIVE TEMPERATURE : TOCEL = ',E15.8, & ' DEG. C',15X,'T0F= ',£15.8,' DEG. F')
923 FORMATQX/,T5,'EFFECTIVE RADIANT FIELD .ERFSI = ',E15.8, & * W/SQ.M *,15X,*ERFFPS = ',E15.8,' BTU/HR.SQFT.*)
C WRITE(6,1005) AUSTC,AUST-460.D0
1005 FORMAT(1X//,T5,'AVERAGE UNHEATED SURFACE TEMPERATURE, AUST : & * = \E15.8,' DEG. C.'.ISX.EIS.S,' DEGF.')
C C C C C C
WRITE(6,902) QNET1 902 FORMAT(1H1,/,T5,'QNET1 = Ql +Q2 -Q3 +Q4 = ',E15.8)
WRITE(6,903) QNETP2 903 FORMAT(1X//.T5,'QNETP2 = Ql +Q2 -Q3 -QP5 = \E15.8)
WRITE(6,904) QNET3 904 FORMAT(1X//,T5,'QNET3 = Ql +Q2 -Q3 -Q6 = ',E15.8)
WRITE(6,907) Q7 + QSTD3 907 F0RMAT(1X//,T5,'STD.Q-DESIGN = Q7 +QSTD3 = '.E15.8)
WRITE(6,905) Q6 +QACT3 905 F0RMAT(1X//,T5,*Q-DESIGN-0VERALL = Q6 +QACT3 = \E15.8)
PDIFF1 = -((Q7 + QSTD3) - (Q6 + QACT3))*10O.D0/(Q7+QSTD3) WRITE(6,1006) PDIFF1
1006 FORMAT(1X//,T5,*PERCENTAGE DIFFERENCE BETWEEN STD.Q-DESIGN*, & ' AND Q-DESIGN-OVERALL = ',/,T5, & * PDIFF1 = -((Q7 + QSTD3) - (Q6 + QACT3) )*100.D0/(Q7+QSTD3)', & ' = ',F7.2,' %',/)
WRITE(6,1513) Q5 + QACT3 1513 F0RMAT(1X//,T5,'Q-DESIGN-C0ND = Q5 + QACT3 = *,E15.8) C CALCULATING % LOAD OF DESIGN
PDIFF2 = -((Q7 + QSTD3) - (Q5 + QACT3) )*100.D0/(Q7+QSTD3) WRITE(6,1514) PDIFF2
1514 FORMAT(1X//,T5,*PERCENTAGE DIFFERENCE BETWEEN STD.Q-DESIGN', & ' AND Q-DESIGN-COND1 = ',/,T5, & ' PDIFF2 = -((Q7 + QSTD3) - (Q5 +QACT3) )*100.D0/(Q7+QSTD3)', & ' = ',F7.2,' %',/)
WRITE(6,906) Q5 + Q3 906 F0RMAT(1X//,T5,'Q-DESIGN-C0ND = Q5 + Q3 = ',E15.8) C CALCULATING % LOAD OF DESIGN
PDIFF3 = -((Q7 + QSTD3) - (Q5 + Q3) )*100.D0/(Q7+QSTD3) WRITE(6,1031) PDIFF3
1031 F0RMAT(1X//,T5,'PERCENTAGE DIFFERENCE BETWEEN STD.Q-DESIGN', & ' AND Q-DESIGN-C0ND2 = \/,T5, & ' PDIFF3 =, -((Q7 + QSTD3) - (Q5 + Q3) )*100.D0/(Q7+QSTD3)', & ' = ',F7.2,' %',/)
WRITE(6,908) Ql + Q2 908 FORMAT(1X//,T5,'QINPUT = Ql +Q2 = '.E1S.8)
PDIFF4 = -((Q7 + QSTD3) - (Q1+Q2 ) )*100.D0/(Q7+QSTD3)
C-12
WRITE(6,1032) PDIFF4 1032 FORMAT(1X//,T5,'PERCENTAGE DIFFERENCE BETWEEN STD.Q-BESIGN *,
& ' AND Q-INPUT= Q1+ Q2 = ',/,T5, & ' PDIFF4 = -((Q7 + QSTD3) - (Ql + Q2) )*100.D0/(Q7+QSTD3)', & ' = '.F7.2,' %',/)
IF(CP) THEN WRITE(6,806)QCVP,QRP
806 FORMAT(1X//.T5,'CONVECTION FROM PANELS TO ROOM ,QCVP = ', & E15.8,//,T5,'RADIATION FROM PANELS TO ROOM.QRP = ',E15.8)
WRITE(6,1512)100.DO*QRP/(Q1+Q2) 1512 F0RMAT(1X//,T5,'PERCENTAGE RADIATION = 100* QRP/(Q1+Q2) = ',
& F7.2,' % ' ) WRITE(6,1033) PAREA*100.D0/(RAREA(6))
1033 FORMAT(1X//.T5,'PERCENTAGE OF CEILING COVERED WITH PANELS*, & ' = ' T**7 0 ' *Y '
WRITE(6,809)Q1/PAREA , X(10)-460.D0 809 F0RMAT(1X//,T5,'HEAT OUTPUT BY PANELS DOWNWARDS, PER UNIT',
& ' AREA OF PANELS,Ql/PAREA = ',F10.2,' BTU/HR.SQFT.' ,/,T5, & ' FOR PANELS AT ',F7.2,' DEG.F ')
WRITE(6,1034)Q1/RAREA(6) 1034 FORMAT(1X//.T5,'HEAT OUTPUT BY PANELS DOWNWARDS, PER UNIT',
& ' AREA OF WALL.6,Q1/RAREA(6) = '.F10.2,' BTU/HR.SQFT.') PARM1 = Q1/(PAREA*(X(10)-X(8) )) WRITE(6,1035) PARM1
1035 FORMAT(lX//.T5,'QPANEL/( PAREA*(X(10)-X(8) )) = ', & T55,'PARM1 = *,E15.8)
PARM2 = Q1/(RAREA(6)*(X(10)-X(8) )) WRITE(6,1036) PARM2
1036 FORMAT(1X//.T5,'QPANEL/(RAREA(6)*(X(10)-X(8) )) = *, & T55,'PARM2 = f,E15.8)
SIGMAR = 0.1714D-8 PARM3 = Ql/( PAREA*(X(10)-X(8) )*(XC10)**4-AUST**4)*SIGMAR) WRITE(6,1037) PARM3
1037 FORMAT(1X//.T5, & ' QPANEL/( PAREA*(X(10)-X(8) )*(X(10)**4-AUST**4)*SIGMAR )', & T65,'PARM3 = ',E15.8)
PARM4 = Q1/(RAREA(6)*(X(10)-X(8) )*(X(10)**4-AUST**4)*SIGMAR) WRITE(6,1038) PARM4
1038 FORMAT(1X//,T5, & ' QPANEL/(RAREA(6)*(X(10)-X(8) )*(X(10)**4-AUST**4)*SIGMAR )', & T65,'PARM4 = ',E15.8)
C C C-—
ENDIF C
WRITE (6,202) Q1,Q2,Q3 202 FORMAT (IX ,/,T5,
&'Q1= NET HEAT INPUT TO THE ROOM BY SUPPLY AIR OR PANELS', & T65,'=',E15.8,//, &T5,'Q2= HEAT INPUT BY PEOPLE AND LIGHTS',T65,'=',E15.8,//,T5, &'Q3= ACTUAL INFILTRATION LOSS = 1.08*CFM*(XINF-TOUT)', &T75,'=',E15.8) WRITE (6,1515) QACT3
1515 FORMAT (1X,/,T5, &'QACT3 = INFILTRATION LOSS (NO GRAD)= 1.08*CFM*(TA-TOUT)', & T75,'=',E15.8) WRITE (6,911) QSTD3
911 FORMAT (1X,/,T5, &'QSTD3 = STANDARD INFILTRATION LOSS = 1.08*CFM*(75-TOUT)',
C-13
& T75,'=\E15.8) WRITE (6,801) Q4
801 FORMAT(1X//.T5,'BASED ON TOTAL AREAS,',/,T5, &'Q4= TOTAL HEAT LOST FROM SURFACES TO AIR BY CONVECTION1, &T65,'=',E15.8) WRITE (6,912) Q5
912 F0RMAT(//,T5,'Q5 = ', &'HEAT LOST THROUGH SURFACES TO THE OUTSIDE BY CONDUCTION', & T65,'=',E15.8) WRITE (6,913) Q6
913 F0RMAT(//,T5,'Q6 = ', &'OVERALL ROOM HEAT LOSS = SUM OF U*A*(X(8)-T0UT)', & T65,'=',E15.8) WRITE (6,914) Q7
914 FORMAT(//,T5,'Q7 = ', &*STD.OVERALL ROOM HEAT LOSS = SUM OF U*A*(75-T0UT)', & T65,'=',E15.8) WRITE (6,915) Q8
915 FORMAT(//,T5,'Q8 = ', &'SUM OF NET OUTWARD RADIATION FROM THE SURFACES', & T65,'=*,E15.8)
C WRITE(6,123) " '
C WRITE(6,123) WRITE(6,25) (K4,QR(K4),QCV(K4),QCD(K4),K4=1,6)
25 FORMAT(1H1,/,T20,'PER UNIT AREA HEAT FLOW ', & //,T18 'A POSITIVE VALUE DENOTES A LOSS FROM THE SURFACE', &//,T15,fQR(I)',T35,*QCV(I)',T55,,QCD(I)',/,(/,T2,Il,T8,E15.8, & T28,E15.8,T48,E15.8) ) WRITE(6,802)
802 FORMAT(IX,//,T20,'HEAT FLOW THROUGH THE ROOM SURFACES(BTU/HR)', & //,T18.'A POSITIVE VALUE DENOTES A LOSS FROM THE SURFACE', &//.T15,fQRT',T35,'QCVT*,T55,'QCDT')
DO 804 IR = 1,6 QRT = QR(IR)*RAREA(IR) QCVT = QCV(IR)*RAREA(IR) QCDT = QCD(IR)*RAREA(IR)
IF(CP .AND. IR.EQ.6) THEN QRT = QR(IR)*PRAREA QCVT = QCV(IR)*PRAREA QCDT = QCD(IR)*PRAREA
ENDIF C
WRITE(6,803) IR,QRT,QCVT,QCDT 803 F0RMAT(1X//,T2,I1,T8,3(E15.8,5X) ) 804 CONTINUE
WRITE(6,125) TERM6,TERM7,ALHS 125 F0RMAT(1X///,T5,'RADIATION EXCHANGE BY PERSON,TERM6',T45,
&'=*,E19.12,//,T5,,CONVECTIVE EXCHANGE,TERM7',T45,'=',E19.12,//,T5, &'ALHS (SHOULD BE = TERM6+TERM7)',T45,'=',E19.12,T70, &'= CONDUCTION THROUGH THE CLOTHING')
C NEW PAGE WRITE (6,123)
C PRINTING THE INPUT DATA AS A CHECK AT THE END OF THE PROGRAM WRITE (6,123)
C WRITE (6,234) WRITE(6,100?)
1007 FORMAT(1H1.T35,'ECHO OF INPUT DATA (AS A CHECK)',/) WRITE (6,234)
C-14
WRITE(6,123) C C WRITE (6,31)N,NSIG,ITMAX 31 FORMAT(IX,//,T5,'NUMBER OF EQUATIONS* , T35 ,'N = ', 12,
&//T5,'NUMBER OF SIGNIFICANT DIGITS' T35,'NSIG = ',12, &//T5,'MAX. NO. OF ITERATIONS', T35, ITMAX = ',14) WRITE(6,78)ALTH,BTH,HT WRITE (6,66)T0UT - 460.DO WRITE (6,1)AM1,EFF WRITE (6,24)AICL.FCL,HC
24 FORMAT (1X,/,T5 'AICL =',F5.2,' CLO',12X,'FCL =',F5.2,10X, &'HC =',E15.8, <KCAL/HR.(SQ.M).C') WRITE (6,3)V WRITE (6,4)RH,PA
4 FORMAT(1X/,T5,'RELATIVE HUMIDITY, RH = ',F5.3,10X, &'PA (MM HG) = ',E15.8) WRITE (6,5)P WRITE(6,921) RATIO IF( P.EQ.O.DO) THEN
WRITE(6,7)Q2P ELSE
WRITE (6,7)Q2P/P ENDIF
7 FORMAT (IX,/,T5,'SENSIBLE HEAT (PEOPLE) BTU/HR/PERSON =',E15.8) WRITE (6,8)Q2L WRITE (6,6)CFM,ACH WRITE (6,1003)HREF,SLOPE
C WRITE(6,211)(K3,K3=1,6) WRITE(6,21) (U(K3),K3=1,6) WRITE(6,22) (HI(K3),K3=1,6)
C THE CI.S HAVE BEEN CALCULATED BY THE PROGRAM.. WRITE(6,23) (CI(K3),K3=1,6) WRITE(6,235) (EPSI(K3),K3=1,6) WRITE(6,609) (FP(K3),K3=1,6)
C C FOR THIS 1 CASE C
WRITE(6,1520) 1520 FORMAT( 1H1,/) C C C C C C C
IF(CP) THEN C C WRITE(6,2501)XPREF-460.D0 2501 FORMAT(T25,*PANEL TEMPERATURE',T50 ,'= ',F6.1) C WRITE (6,1521) PAREA 1521 F0RMAT(1X//,T5,'T0TAL PANEL AREA ',T50,'= ', F10.1,' SQ.FT ') C
ENDIF C C C WRITE(6,1522) Q7 + QSTD3 1522 FORMAT(1X//,T5,'ASHRAE DESIGN HEAT LOSS',T50,'= ',F10.1) C WRITE(6,1523) Q6 +QACT3 1523 F0RMAT(1X//,T5,'ACTUAL DESIGN HEAT LOSS',T50,'= '.F10.1)
C-15
C WRITE(6,1524) PDIFF1 1524 FORMAT(lX//,T5,'PDIFFl',T60,,= *,F7.2,' %'J) C WRITE(6,1525) Q5 +QACT3 1525 FORMAT(1X//,T5,'C0NDUC.DESIGN HEAT LOSS l',T50,'= ',F10.1) C WRITE(6,1526) PDIFF2 1526 FORMAT(1X//,T5,'PDIFF2',T60,,= ',F7.2,' %*,/) C WRITE(6,1527) Q5 +Q3 1527 FORMAT(1X//,T5,'C0NDUC.DESIGN HEAT LOSS 2',T50,'= *,F10.1) C WRITE(6,1528) PDIFF3 1528 FORMAT(1X//.T5,'PDIFF3',T60, * = *,F7.2,' %',/) C WRITE(6,1529) Ql +Q2 1529 F0RMAT(1X//,T5,'ACTUAL HEAT INPUT ',T50,*= ',F10.1) C WRITE(6,1530) PDIFF4 1530 FORMAT(1X//,T5,'PDIFF4',T60,'= ',F7.2,' %',/) C
IF(CP) THEN C C WRITE(6,1531)100.D0*QRP/(Q1+Q2) 1531 FORMAT(1X//,T5,*PERCENTAGE RADIATION ',T50,'= \F7.2,' %') C WRITE(6,1532) PAREA*100.D0/(RAREA(6)) 1532 FORMAT(1X//.T5,'PERCENT CEILING COVERED BY PANELS',
& T60,'= ',F7.2,' % ') C WRITE(6,1533)Q1/PAREA 1533 FORMAT(1X//,T5,'HEAT OUTPUT PER UNIT AREA',T50,'= ',F10.2 ) C WRITE(6,1534) PARM1 1534 FORMAT(1X//,T5,'PARAMETER l',T50,'= ',F9.5) C WRITE(6,1535) PARM3 1535 FORMAT(1X//,T5,*PARAMETER 3',T50,'= ',F10.6) C
ENDIF C
IF(HF) THEN CONTINUE
ELSE C WRITE(6,1536) X(5)-460.D0 1536 FORMATQX//,T5,'FLOOR TEMPERATURE',T50,' = ',F5.1 )
ENDIF C C WRITE(6,1537) X(8)-460.D0 1537 F0RMAT(1X//,T5,'ROOM AIR TEMPERATURE',T50,*= ',F5.1 ) C WRITE(6,1538) X(9)-460.D0 1538 FORMAT(1X//,T5,*MEAN RADIANT TEMPERATURE',T50,'= ',F5.1 ) C WRITE(6,1539) TOF 1539 FORMATQX//,T5,'OPERATIVE TEMPERATURE' ,T50, '= ' ,F5.1 ) C WRITE(6,1540) ERFFPS 1540 FORMAT(1X//,T5, EFF. RADIANT FIELD',T50,'= \F5.1 ) C WRITE(6,1541) AUST-460.D0 1541 F0RMAT(1X//,T5,'A.U.S.T.',T50,'= *,F5.1 ) C C
IF(CP) THEN CONTINUE
ELSE C WRITE(6,1542) X(10)-460.D0 1542 F0RMATC1X//,T5,* SUPPLY AIR TEMPERATURE',T50,'= ',F5.1 )
ENDIF C C TRYING TO WRITE IN THE DESIRED FORMAT C 0UT(25,25) C ASDF
0UT(IVAR,1) = ACH
C-16
PAREA Q7 + QSTD3 Q6 +QACT3 PDIFF1 Q5 +QACT3 PDIFF2 Q5 +Q3 PDIFF3
0UT(IVAR,2) 0UT(IVAR,3) 0UT(IVAR,4) 0UT(IVAR,5) 0UT(IVAR,6) 0UTCIVAR,7) OUT(IVAR,8) OUT(IVAR,9) OUT(IVAR,10) = Ql +Q2 OUT(IVAR,ll)= PDIFF4 OUT(IVAR,12)= 100.DO*QRP/(Q1+Q2) OUT(IVAR113)= PAREA*100.DO/(RAREA(6))
IF(CP) THEN 0UT(IVAR,14)= Ql/PAREA
ENDIF OUT(IVAR,15)= X(5)-460.D0 OUT(IVAR,16)= X(8)-460.D0 OUTCIVAR,17)= X(9)-460.D0 OUT(IVAR,18)= TOF OUT(IVAR,19)= ERFFPS OUT(IVAR,20)= AUST-460.D0
0UT(IVAR,21)= X(10)-460.D0
C C
C C C
OUTl(IVAR,l)= PARM1 OUTl(IVAR,2)= PARM3 OUTl(IVAR,3)= ALTH 0UT1(IVAR,4)= BTH
C C C THIS IS THE END OF THE IVAR=1,NVAR LOOP C 1550 CONTINUE C C OUT(l,23) ,0UT(2,23) ETC HAVE CONSECUTIVE VALUES OF X(10)-460. C C ASDF FIRST LINE
POUT( POUTC POUT( POUT( POUT( POUTC POUTC POUTC POUTC POUTCIO POUTC11 POUTC12 POUTC13 POUTC14 POUTC15 POUTC16 POUTC17 POUTC18 POUTC19 POUTC20
'INFILTRATION AC/H 'PANEL AREA REQUIRED , SQ FT 'ASHRAE DESIGN HEAT LOSS, BTU/HR 'ACTUAL DESIGN HEAT LOSS, BTU/HR 'PERCENTAGE DIFFERENCE 1 'CONDUCTION DESIGN HEAT LOSS 1, BTU/HR 'PERCENTAGE DIFFERENCE 2 'CONDUCTION DESIGN HEAT LOSS 2, BTU/HR 'PERCENTAGE DIFFERENCE 3 'ACTUAL HEAT INPUT, BTU/HR 'PERCENTAGE DIFFERENCE 4 'PERCENTAGE RADIATION 'PERCENT CEILING COVERED BY PANELS 'HEAT OUTPUT PER UNIT PANEL AREA, BTU/HR.SQ FT 'FLOOR TEMPERATURE, DEG. F 'ROOM AIR TEMPERATURE, DEG. F 'MEAN RADIANT TEMPERATURE, DEG. F 'OPERATIVE TEMPERATURE, DEG. F 'EFFECTIVE RADIANT FIELD, BTU/HR. SQ FT 'A.U.S.T, DEG. F
C-17
c c c
P0UT(21) = 'SUPPLY AIR TEMPERATURE, DEG. F
POUTl(l) = 'PARAMETER 1, BTU/HR.SQ FT.F P0UT1(2) = 'PARAMETER 3, DIMENSIONLESS P0UT1(3) = 'NO GLASS IN ANY WALL P0UT1(4) = 'ONE WALL, HALF GLASS POUTl(5) = 'ONE WALL, ALL GLASS POUTl(6) = 'ONE WALL, ALL GLASS-SECOND WALL, HALF GLASS POUTl(7) = 'TWO WALLS, ALL GLASS POUTl(8) = 'PANEL TEMPERATURE, DEG. F POUTl(9) = "U-FACTOR CASE
C C C REWRITING ONLY THE REQUIRED OUTPUT
WRITE(6,1520) C ABOVE GIVES NEW PAGE C
WRITE(6,2500) TITLE 2500 FORMAT(T1,'.CE *,A50,/) C C WRITE (6,3523) P0UT1(9), (J,J=1,NVAR) 3523 FORMAT(/,T2,A45,T50,7(I8,1X) ) C c c
WRITE (6,3511) POUT(l), (0UT(J,1),J=1,NVAR) 3511 FORMAT(/,T2,A45,T50,7(F8.1,1X) ) C C THE FOLLOWING IS FOR PANEL HEATING C
IF (CP) THEN DO 3111 I = 2,20
WRITE (6,3120) POUT(I),(OUT(J,I),J=1,NVAR) 3120 FORMAT(/,T2,A45,T50,7(F8.1,1X) ) 3111 CONTINUE C
DO 3503 I = 1,2 WRITE(6,3502) POUTl(I), (0UT1(J,I),J=1,NVAR)
3502 F0RMAT(/,T2,A45,T50,7(F8.4,1X) ) 3503 CONTINUE
ELSE C C C THE FOLLOWING IS TO BE USED FOR CONVECTIVE HEATING C
DO 5000 I = 3,11 WRITE (6,5001) POUT(I),(OUT(J,I),J=l,NVAR)
5001 F0RMAT(/,T2,A45,T50,7(F8.1,1X) ) 5000 CONTINUE
DO 3550 I = 15,21 WRITE (6,3551) POUT(I),(OUT(J,I),J=1,NVAR)
3551 FORMAT(/,T2,A45,T50,7(F8.1,1X) ) C 3550 CONTINUE
ENDIF C C THE FOLLOWING IS FOR THE GLASS CASE C DO 3521 I = 1,5 C WRITE(6,3522) I, POUTl(I+2)
C-18
3522 F0RMAT(/,T2,'CASE NUMBER ',12,': *, A45) 3521 CONTINUE C C C WRITING WITH DEVICE TYPE 8 SO AS TO GET AN OUTPUT IN A DIFFERENT FIL C C
WRITE(8,4014) WRITE(8,4010) WRITE(8,4011) WRITE(8,4012) WRITE(8,4013)
4014 FORMAT(T1,'.RF CANCEL') 4010 FORMAT(T1,'.LL 120') 4011 FORMAT(Tl.'.PN OFF') 4012 FORMAT(Tl.'.FO OFF') 4013 FORMAT(Tl,'.US l')
WRITE(8,2500) TITLE C C C
WRITE (8,3511) POUT(l), (OUT(J.l),J=1,NVAR) C C
IF(CP) THEN C C THE FOLLOWING IS FOR PANEL HEATING
DO 4000 I = 2,20 WRITE (8,3120) POUT(I),(OUT(J,I),J=1,NVAR)
4000 CONTINUE C C THE FOLLOWING IS FOR CONVECTIVE HEATING
DO 4002 I = 1,2 WRITE(8,3502) POUTl(I), (0UT1(J,I),J=1,NVAR)
4002 CONTINUE C
ELSE C C
DO 5002 I = 3,11 WRITE (8,5001) POUT(I),(OUT(J,I),J=l,NVAR)
5002 CONTINUE DO 5003 I = 15,21
WRITE (8,3551) POUT(I),(0UT(J,I),J=1,NVAR) 5003 CONTINUE
ENDIF C C 999 STOP
END C C C C C******* SUBROUTINE STARTS *****fr&****ftifo ^ C
SUBROUTINE FCN(X,F,N,PAR) IMPLICIT REAL*8 (A-H,0-Z) LOGICAL CP,COOL,HF DIMENSION X(10),F(10),PAR(10) COMMON /OUT/ TOUT
C-19
COMMON /CEL/ XCEL(ll) COMMON /COMF/ AM1,EFF,AICL,FCL,HC,V,?A,RH COMMON /AIR1/ P,CFM,AICFM,ACH,UQ2P,Q2P,UQ2L,Q2L COMMON /Q/ Q1,Q2,QSTD3,QACT3,Q3,Q4,QP5,Q5,Q6,Q7,Q8 COMMON /QNET/ QNETl,QNETP2,QNET2,QNET3,OUA COMMON /QP/ QCVP.QRP COMMON /QI/ QR(6),QCV(6),QCD(6) COMMON /U/ U(6),HI(6),CI(6),EPSI(6) COMMON /UP/ UP,HIP,CIP,EPSIP,XMULT COMMON /CONV/ HIUP.HIDOWN COMMON /FSURF/ FS(6,6) COMMON /FPEOP/ FP(6) COMMON /GRAD/ HREF,SLOPE COMMON /TERM/ TERM6,TERM7,ALHS COMMON /DIMEN/ ALTH,BTH,HT COMMON /AREAS/ RAREA(6) COMMON /PAN1/ CP,COOL,HF COMMON /PAN2/ PRAREA,PAREA COMMON /PAN3/ NCP COMMON /PAN4/ FSCP(25,25) COMMON /DIMP/ XCP(25),YCP(25),ALCP(25),BCP(25)
C FSURF APPEARS IN MPROG. AND FCN C PAN1 APPEARS IN MPROG. AND FCN' C PAN2 APPEARS IN MPROG. AND FCN C PAN3 APPEARS IN MPROG. AND FCN C PAN4 APPEARS IN MPROG. AND FCN C C C C C CONVERTING RANKINE TEMP.S TO CELSIUS VALUES
DO 114 11= 1,N 114 XCEL(Il) = (X(I1) -492D0 )/1.8D0 C I I I I I I I I I I l I I I 1 1 I I I I I
IF(CP) THEN XPANEL = X(10)
ENDIF C I M I I I I I I I I H I I I H I C TOTAL AREA OF PANELS,PAREA C AND REST OF CEILING AREA ,PRAREA ARE OBTAINED FROM C THE MAIN PROGRAM THROUGH THE COMMON STATEMENT. C C C C CALCULATING THE RADIATIVE HEAT TRANSFER NET-GOING-OUT ... C FOR EG. C FS(1,5) IS USED AND NOT FS(5,1) BECAUSE C FS(1,5)* AREA(l) = FS(5,1)* AREA(5) AND EVERYTHING HERE IS PRORATED C TO THE AREA OF THE SURFACE EG. AREA(l)
SIGMA = 0.1714D-8 C C
IF(CP) THEN C FOR SURFACES 1 TO 6 (CEILING IS NOW OF AREA PRAREA) ,FSCP MATRIX C IS USED INSTEAD OF FS.
DO 111 J6 = 1,6 QROUT = EPSI(J6) *SIGMA* (X(J6)**4) QRIN =O.DO
C FROM OTHER WALLS, DO 113 J7 = 1,6 QRIN = QRIN + EPSI(J7)* SIGMA* FSCP(J6,J7)* ( X(J7)**4)
C-20
I
113 CONTINUE C FROM THE PANELS, ^ ^ ^ ^ ^ ^ ^ ^ ^
C FOR A SPECIAL CASE , TO MAKE THE PANELS RADIATE ONLY TO THE C FLOOR, RADIATION TO THE WALLS IS MADE EQUAL TO ZERO C C IF THIS IS NOT DESIRED, MAKE STATEMENTS FROM "INFRARED STARTS" TO C "INFRARED ENDS" AS COMMENT LINES. C C
C DO 808 K = 11,NCP+10
C******** INFRARED STARTS********* DO 6002 JJ =1,4
FSCP(JJ,K)= O.DO 6002 CONTINUE
FSCP(5,K) = ( ALCP(K)*BCP(K) )/(ALTH*BTH) C******** INFRARED ENDS **********
QRIN = QRIN + EPSIP*SIGMA* FSCP(J6,K)* (XPANEL**4) 808 CONTINUE
QR(J6) = QROUT -QRIN. Ill CONTINUE
C C NOTE: THE HEAT BALANCE OF THE PORTION COVERED BY PANELS IS NOT C CONSIDERED HERE . IT SHOULD BE TAKEN INTO ACCOUNT WHEN THE ULTIMATE C SOURCE OF HEAT IS EXAMINED. C C
ELSE C I.E, IF NO PANELS
DO 603 J6 = 1,6 QROUT = EPSI(J6) *SIGMA* (X(J6)**4) QRIN =0.D0
DO 604 31 = 1,6 QRIN = QRIN + EPSI(J7)* SIGMA* FS(J6,J7)* ( X(J7)**4 )
604 CONTINUE QR(J6) = QROUT -QRIN
603 CONTINUE C
ENDIF C C NOW COMPUTING THE CONVECTIVE AND CONDUCTIVE PARTS OF HEAT TRANSFER C NET-GOING-OUT AND HENCE MAKING A HEAT BALANCE ..' C C CONVECTION COEFFICIENTS C C FOR USE INSIDE THE PROGRAM, HI(5) & HI(6) MUST BE CHANGED TO C VALUES NOT INCLUDING RADIATION C HIDOWN = 0.162D0 , STANDARD ASHRAE VALUE C HIDOWN IS THE CONVECTION COEFF. WHEN THE HEAT FLOW IS VERTICALLY C DOWNWARD . C HIUP = 0.712D0 .STANDARD ASHRAE VALUE C HIUP IS FOR HEAT FLOW VERTICALLY UPWARD. C C FOR THE CONVECTIVE CASE C — — -C FOR THE FLOOR
IF( X(5).LT.X(8) ) THEN HI(5) = HIDOWN
C-21
ELSE HI(5) = HIUP
ENDIF C FOR THE CEILING...
IF( X(6).LT.X(8) ) THEN HI(6) = HIUP
ELSE HI(6) = HIDOWN
ENDIF C C FOR THE WALLS C C HI BASED ON DELTA T ,BAUMAN*S CORRELATION 1 FT = 0.3048 M (EXACT) C FOR WALLS 1,2,3 AND 4. C NOTE : THIS CORRELATION NEEDS TEMP. DIFF. IN CELSIUS, H IN METRES, CAND THE RESULTING HI IS IN W/M**2,C ( DIV BY 5.68 TO GET BTU/HR.SQ FT.F C
DO 52 K2=l,4 HEI = HT*0.3048DO HIBAU= 2.03D0*( DABS((XCEL(8)- XCEL(K2))/HEI) **0.22D0 ) HI(K2) = HIBAU/5.68D0
52 CONTINUE C C NOW TRYING THE CORRELATIONS GIVEN BY MIN AND SCHUTRUM ... C FOR HEATED CEILING PANEL OR HEATED FLOOR HIP IS GIVEN JUST BEFORE USE. C C THE EQUIVALENT DIAMETER DE FOR THE CEILING OR THE FLOOR IS,
DE = 4.DO* RAREA(5)/ (2.D0*(ALTH+BTH) ) C C HEATED CEILING CASE C -
IF(CP) THEN C C THAT IS, IF THE PANELS ARE IN THE CEILING.. C X(5)= FLOOR AND X(6) = CEILING TEMPS. C C UNHEATED CEILING PORTION ...
IF( X(6).LT.X(8) ) THEN HI(6) = HIUP
ELSE HI(6) = HIDOWN
ENDIF C FLOOR C HI(5) = 0.041D0*( DABS( X(8)-X(5) )**0.25D0)/(DE**0.25D0) C C TRYING ASH STD FOR THE FLOOR C
IF( X(5).LT.X(8) ) THEN HI(5) = HIDOWN
ELSE HI(5) = HIUP
ENDIF C C C WALLS
DO 950 IW = 1,4 CI I IMIIHI(IW) = 0.26D0* (DABS(X(IW)-X(8) )**0.32D0 )
HI(IW) = 0.29D0*( DABS(X(IW)-X(8) )**0.32D0 )/(HT**0.05D0) 950 CONTINUE C C THE ENDIF FOR THE IF (CP) FOLLOWS.
C-22
ENDIF C C HEATED FLOOR C C THAT IS, IF A HEATED FLOOR IS USED , ABOVE ARE TO BE CHANGED C X(5) IS NOW THE CEILING TEMP & X(6) IS THAT OF THE FLOOR.. C
IF(HF) THEN C C CEILING C***********HI(5) = 0.39* (DABS(X(8)-X(5) )**0.31DO )/(DE**0.08)
IF( X(5).LT.X(8) ) THEN HI(5) = HIUP
ELSE HI(5) = HIDOWN
ENDIF C UNHEATED FLOOR PORTION
IF( X(6).LT.X(8) ) THEN HI(6) = HIDOWN
ELSE HI(6) = HIUP
ENDIF C C WALLS "
DO 1200 IW = 1,4 CI I I I I IIHI(IW) = 0.26D0* (DABS(X(IW)-X(8) )**0.32D0 )
HI(IW) = 0.29D0*( DABS(X(IW)-X(8) )**0.32D0 )/(HT**0.05D0) 1200 CONTINUE C C THE ENDIF FOR THE IF (HF) FOLLOWS.
ENDIF C C
QCV(1)= HI(1)*( X(l)-X(8) ) QCD(1)= CI(1)*(X(1)-T0UT) F(l)= QR(1) + QCV(l) +QCD(1)
C QCV(2)= HI(2)*(X(2)-X(8)) QCD(2)= CI(2)*(X(2)-T0UT) F(2)= QR(2) + QCV(2) +QCD(2)
C QCV(3)= HI(3)* (X(3)-X(8)) QCD(3)= CI(3)*(X(3)-T0UT) F(3)= QR(3) + QCV(3) +QCD(3)
C QCV(4)= HI(4)* (X(4)-X(8)) QCD(4)= CI(4)*(X(4)-T0UT) F(4)= QR(4) + QCV(4) +QCD(4)
C C
QCV(5)= HI(5)* ( X(5) -X(8))
C QCV(5)= HI(5)* ( X(5) - (X(8)-H*G) )
QCD(5)= CI(5)*(X(5)-T0UT) F(5)= QR(5) + QCV(5) +QCD(5)
C IF(CP) THEN
C I.E, WITH CEILING PANELS, C QCV(6) IS NOW PER UNIT PRAREA
C-23
\
QCV(6) = HI(6)* ( X(6)-X(8) ) C** . =+ C QCV(6)= HI(6)*( X(6) -( X(8)+(8.D0-H)*G ) ) C** + C
QCD6 = CI(6)*(X(6)-TOUT) IF ( PRAREA .LT. 1.0D-4) THEN
QCD(6) = O.DO X(6) = XPANEL F(6) = O.DO
ELSE QCD(6) = QCD6 F(6)= QR(6) + QCV(6) +QCD(6)
C NOTE:THIS HEAT BALANCE IS PER UNIT OF REDUCED CEILING, PRAREA END IF
ELSE C I.E, WITHOUT ANY CEILING PANELS,
QCV(6)= HI(6)*( X(6) -X(8) )
C QCV(6)= HI(6)*( X(6) -( X(8)+(8.D0-H)*G ) )
QCD(6)= CI(6)*(X(6)-TOUT) F(6)= QR(6) + QCV(6) +QCD(6)
ENDIF C C C C C C THE COMFORT EQUATION ( TEMPS. ARE CONVERTED TO CELSIUS) C THE FOLLOWING IS TO EVALUATE PA AT THE CURRENT AIR TEMP(R): X(8) C 51.715 CONVERTS PSI TO MM HG
C8=-10440.4 C9=-ll.29466692 C10=-0.02700133 Cll=0.1289706D-4 C12=-0.2478068D-8 C13=6.5459673
C TO TAKE CARE OF NEGATIVE X(8)!M (DURING ITERATIONS), IF(X(8).LT.1.E-50)THEN
ALNPWS = O.DO ELSE ALNPWS=C8/X(8) +C9 +C10*X(8) +C11*(X(8)**2) +C12*(X(8)**3)+
& C13*DL0G(X(8)) ENDIF PA=DEXP(ALNPWS) * 51.715D0* RH
C C
TERM1= AM1*(1.D0-EFF) TERM2= 0.35D0* ( 43.D0-0.061D0*AMl*(l.DO-EFF)-PA ) TERM3= 0.42D0*( AM1*(1.D0-EFF) -50.DO) TERM4= 0.0023* AMI*(44.DO-PA) TERM5= 0.0014DO*AM1*(34.DO- XCEL(8) )
C ALHS REPRESENTS THE NET CONDUCTION THROUGH THE CLOTHING ALHS = TERM1 - TERM2 -TERM3 -TERM4 -TERM5 TERM6= 3.4D-8*FCL*( (XCEL(7) +273.D0)**4 - (XCEL(9)+273.D0)**4 )
C C HCV IS THE CONV. COEFF. OF CLOTHING BASED ON VELOCITY, (FORCED CONV) C HCTD, BASED ON TEMP. DIFF. FOR FREE -CONV. C OBSERVE THE IF LOOP.., THE GREATER OF THE TWO VALUES IS USED
HCV = 10.4D0* DSQRT(V)
C-24
HCTD= 2.05*( (DABS(XCEL(7)-XCEL(8)) )**0.25D0 ) HC= HCTD
IF (HCV.GT.HC) THEN HC= HCV
ENDIF CC CC CC CC
TERM7= FCL*HC*( XCEL(7)-XCEL(8) ) RHS = TERM6 +TERM7 TERM8= 0.032DO*AM1*(1.DO-EFF)
C F(7)= 35.7D0- TERM8 -0.18*AICL*ALHS - XCEL(7) F(8)= RHS- ALHS
C CALCULATION OF MEAN RADIANT TEMPERATURE (NEGLECTING REFLECTIONS), IF (CP) THEN
C THIS DIVISION OF FP(6) IS QUITE APPROXIMATE. FP6 = FP(6)*PRAREA/ (ALTH*BTH) FPP = FP(6)*PAREA / (ALTH*BTH)
FPTOT = O.DO DO 607 IP = 1,5 FPTOT = FPTOT + FP(IP)* (X(IP)**4)
607 CONTINUE FPTOT = FPTOT + FP6*(X(6)**4) + FPP*(XPANEL**4) F(9) = FPTOT - ( X(9)**4 )
ELSE FPTOT = O.DO DO 608 IP = 1,6 FPTOT = FPTOT + FP(IP)* (X(IP)**4)
608 CONTINUE F(9) = FPTOT - ( X(9)**4 )
ENDIF C C C TO GET THE HEAT INPUT TO THE ROOM BY THE PANELS BY C CONVECTION AND RADIATION, TO INCLUDE IN THE HEAT BALANCE OF THE ROOM
IF(CP) THEN C FOR THE PANELS.IN THE CEILING C THE EQUIVALENT DIAMETER DE FOR THE CEILING OR THE FLOOR IS,
DE = 4.DO* RAREA(5)/ (2.D0*(ALTH+BTH) ) C
C
C IF(HF) THEN
C FOR THE PANELS.IN THE FLOOR C
HIP = 0.39* (DABS(X(8)-XPANEL)**0.31D0 )/(DE**0.08D0) C C*****HIP = 0.31D0* ( DABS(X(8)-X(6) )**0.31D0 ) C •: HIP = HIUP
ENDIF C
IF(CP) THEN UQPOUT = EPSIP *SIGMA * (XPANEL**4) QPOUT = UQPOUT*PAREA
C RADIATION OUTWARDS FROM ALL THE PANELS QPIN =0.D0
HIP = O.041D0*(DABS(X(8)-XPANEL)**0.25D0 )/(DE**0.25D0)
ENDIF
C-25
HCTD= 2.05*( (DABS(XCEL(7)-XCEL(8)) )**0.25D0 ) HC= HCTD
IF (HCV.GT.HC) THEN HC= HCV
ENDIF CC CC CC CC
TERM7= FCL*HC*( XCEL(7)-XCEL(8) ) RHS = TERM6 +TERM7 TERM8= 0.032D0*AM1*(1.D0-EFF)
C F(7)= 35.7D0- TERM8 -0.18*AICL*ALHS - XCEL(7) F(8)= RHS- ALHS ___tJ_1_^_1_tJ_ J_1_1_UJ_^
C CALCULATION OF MEAN RADIANT TEMPERATURE (NEGLECTING REFLECTIONS), IF (CP) THEN
C THIS DIVISION OF FP(6) IS QUITE APPROXIMATE. FP6 = FP(6)*PRAREA/ (ALTH*BTH) FPP = FP(6)*PAREA / (ALTH*BTH)
FPTOT = 0.D0 DO 607 IP = 1,5 FPTOT = FPTOT + FP(IP)* (X(IP)**4)
607 CONTINUE FPTOT = FPTOT + FP6*(X(6)**4) + FPP*(XPANEL**4) F(9) = FPTOT - ( X(9)**4 )
ELSE FPTOT = 0.D0 DO 608 IP = 1,6 FPTOT = FPTOT + FP(IP)* (X(IP)**4)
608 CONTINUE F(9) = FPTOT - ( X(9)**4 )
ENDIF C C C TO GET THE HEAT INPUT TO THE ROOM BY THE PANELS BY C CONVECTION AND RADIATION, TO INCLUDE IN THE HEAT BALANCE OF THE ROOM
IF(CP) THEN C FOR THE PANELS.IN THE CEILING C THE EQUIVALENT DIAMETER DE FOR THE CEILING OR THE FLOOR IS,
DE = 4.DO* RAREA(5)/ (2.D0*(ALTH+BTH) ) C
C
C
HIP = 0.041D0*(DABS(X(8)-XPANEL)**0.25D0 )/(DE**0.25D0)
ENDIF
IF(HF) THEN C FOR THE PANELS.IN THE FLOOR C
HIP = 0.39* (DABS(X(8)-XPANEL)**0.31D0 )/(DE**0.08D0) C C*****HIP = 0.31D0* ( DABS(X(8)-X(6) )**0.31D0 ) C = = HIP = HIUP
ENDIF C
IF(CP) THEN UQPOUT = EPSIP *SIGMA * (XPANEL**4) QPOUT = UQPOUT*PAREA
C RADIATION OUTWARDS FROM ALL THE PANELS QPIN =0.D0
C-25
DO 602 JP = ll.NCP+10 UQPIN =0.D0 DO 601 J7 = 1,6
INFRARED STARTS********* DO 6001 JI = 1,4 FSCP(JP,JI) = O.DO
6001 CONTINUE FSCP(JP,5) = l.DO
C******** INFRARED ENDS ************ 601 UQPIN = UQPIN + EPSI(J7)* SIGMA* FSCP(JP,J7)* ( X(J7)**4 ) C ABOVE : RADIATION INWARDS FOR THIS ONE PANEL
QPIN = QPIN + UQPIN* ALCP(JP)*BCP(JP) 602 CONTINUE
QRP = QPOUT - QPIN C ABOVE : NET OUTWARD RADIATION FROM ALL PANELS (BTU/HR)
QCVP = ( XMULT*HIP*(XPANEL -X(8)) )* PAREA C ABOVE : CONVECTION FROM TOTAL PANEL AREA TO AIR
Ql = QCVP + QRP C Q1=NET HEAT INPUT BY PANELS (CONDUCTION NOT CONSIDERED)
ELSE Ql= 1.08D0*CFM*ALTH*BTH*( X(10)-X(8) )
C Q1=NET HEAT INPUT BY AIR ENDIF „ .
C C c C***** Q2p = p*UQ2P ,USING A READ-IN VALUE C TO USE THE CONDUCTION THROUGH THE CLOTHING FOR UQ2P C BTU/HR = (#)* KCAL/HR.SQM * (DUBOIS AREA = 1.8)SQM* 3.973D0 BTU/KCAL C FOR AM1= 50,I.E. MET =1, THIS IS 252.8 BTU/HR/PERSON (STD.=250)
Q2P = P*ALHS*1.8D0*3.973D0 C CCCCC Q2L= P*UQ2L
Q2= Q2P + Q2L C Q2=HEAT INPUT BY PEOPLE AND LIGHTS C ACH= NO. OF AIR CHANGES PER HOUR, AICFM = INF. CFM
AICFM = ACH* (ALTH*BTH*HT)/60.D0 QSTD3= 1.08D0*AICFM *( 460.D0+75.D0 -TOUT ) QACT3= 1.08D0*AICFM *( X(8) -TOUT )
C C C
XINF = X(8) + (HT-HREF)*SLOPE Q3= 1.08D0* AICFM* (XINF-TOUT )
Q3=HEAT LOSS DUE TO INFILTRATION AIR
QP4= ALTH*HT*(QCV(1) +QCV(3)) + BTH*HT*(QCV(2) +QCV(4)) & + ALTH*BTH*QCV(5) + PRAREA*QCV(6) QP5= ALTH*HT*( QCD(l) +QCD(3) ) + BTH*HT*( QCD(2) +QCD(4) ) & + ALTH*BTH*( QCD(5) ) + PRAREA*QCD(6) QP8= ALTH*HT*( QR(1) +QR(3") ) + BTH*HT*( QR(2) +QR(4) ) & + ALTH*BTH* QR(5) +PRAREA*QR(6) POUA = ALTH*HT*( U(l)+U(3) ) +BTH*HT*( U(2)+U(4) ) & + ALTH*BTH* U(5)+ PRAREA*U(6)
C C C AND WHEN THERE ARE NO CEILING PANELS,
Q4= ALTH*HT*(QCV(1) +QCV(3)) + BTH*HT*(QCV(2) +QCV(4)) & + ALTH*BTH*(QCV(5) +QCV(6))
Q5= ALTH*HT*( QCD(l) +QCD(3) ) + BTH*HT*( QCD(2) +QCD(4) ) & + ALTH*BTH*( QCD(5) +QCD(6) )
Q8= ALTH*HT*( QR(1) +QR(3) ) + BTH*HT*( QR(2) +QR(4) )
C-26
& + ALTH*BTH*( QR(5) +QR(6) ) OUA = ALTH*HT*( U(l)+U(3j ) +BTH*HT*( U(2)+U(4) ) & + ALTH*BTH*( U(5)+U(6) )
C NEGATIVE OF Q4 = HEAT LOST FROM AIR TO THE SURFACES BY CONVECTION C NOTE: QCVS ARE SET UP AS POSITIVE,WHEN FLOW IS FROM SURFACE TO AIR.
QNET1= Ql +Q2 -Q3 +Q4 C NOTE: QNET1 = 0 = F(10) IS USED FOR CONVECTIVE HEATING C Q5= HEAT LOST THROUGH THE SURFACES BY CONDUCTION TO THE OUTSIDE
QNET2= Ql +Q2 -Q3 -Q5 QNETP2= Ql +Q2 -Q3 -QP5
C NOTE: QNETP2 = 0 = F(10) IS USED FOR PANEL HEATING. Q6= (X(8)-TOUT)* OUA
C Q6=OVERALL ROOM HEAT LOSS Q7= (460.D0+75.D0-TOUT)* OUA
C Q7=STD.OVERALL ROOM HEAT LOSS QNET3 = Ql +Q2 -Q3 -Q6
C IF(CP) THEN
F(10)= QNETP2 ELSE
F(10)= QNET1 ENDIF
C RETURN END
C C C*************************** C* IF(CP) THEN C* ELSE C* ENDIF
SUBROUTINE SHAPE(ALTH,BTH,HT,FS) C INPUT NEEDED : ALTH, BTH, HT C
IMPLICIT REAL*8 (A-H.O-Z) DIMENSION A(4),B(4),FS(6,6) COMMON /AREAS/ RAREA(6) DATA A/4*0.D0/ , B/4*0.D0/
C THIS IS TO CALCULATE THE SHAPE FACTORS OF THE SIX SURFACES, C CEILING, 4 VERTICAL WALLS,PERPENDICULAR TO EACH OTHER & FLOOR. C ******** SEE HANDOUT FOR BELOW CODING EXPLANATION C CALL ERRSET(263,256,1,1,1) C CALL ERRSET(209,256,1,1,1) C *** 263 ERR NO. (209 ALSO) C *** 256 UNLIMITED NO. OF ERROR OCCURRENCES C *** 1, NO OF MESSAGES TO BE PRINTED C *** 1, NO TRACEBACK IS TO BE PRINTED C *** i, STANDARD CORRECTIVE ACTION TO BE APPLIED C *** RANGE OF ERROR NUMBERS UPTO THIS NUMBER<OMITTED HERE) C502 FORMAT(1X/,T5,?ROOM DIMENSIONS : ',3X,'LENGTH = ',F6.2, C & SX.'BREADTH = ',F6.2,5X, 'HEIGHT = ',F6.2) C =====
RAREA(l) = ALTH * HT RAREA(3) = RAREA(l) RAREA(2) = BTH * HT RAREA(4) = RAREA(2) RAREA(5) = ALTH * BTH
C-27
RAREA(6) = RAREA(5) C INITIALISING.TO ZERO
DO 503 IJ = 1,6 FS(IJ,IJ) = O.DO
503 CONTINUE C FOR FS(1,2)...
DO 504 1=1,4 A(I) = O.DO
504 B(I) = O.DO C
A(2) = HT B(2) = HT A(4) = ALTH B(4) = BTH CALL SHPRP(A,B,F12) FS(1,2) = F12
C C C C C FOR FS (1,5)
DO 505 1=1,4 A(I) = O.DO
505 B(I) = O.DO A(2) = ALTH B(2) = ALTH A(4) = HT B(4) = BTH CALL SHPRP(A,B,F12) FS(1,5) = F12
C C C FOR FS(1,3)...
DO 506 1=1,4 A(I) = O.DO
506 B(I) = O.DO A(2) = ALTH B(2) = ALTH A(4) = HT B(4) = HT G = BTH
C CALL SHPRL (A,B,G,F12) FS(1,3) = F12
C C C SINCE 2 AND 4 ARE BOTH OF THE SAME AREA AND ORIENTATION TO 1.
FS(1,4) = FS(1,2) FS(1,6) = FS(1,5)
C C USING THE PRINCIPLE OF RECIPROCITY
DO 507 IJ2 = 2,6 507 FS (IJ2,1) = FS(1,IJ2)* RAREA(1)/RAREA(IJ2) C FOR (SURFACE.3)
FS(2,3) = FS(2,1) DO 508 IJ4 = 4,6
508 FS(IJ4,3) = FS(IJ4,1) C FOR FS (2,5)....
DO 509 1=1,4 A(I) = O.DO
509 B(I) = O.DO
C-28
c c C FOR
510
C
A(2) = BTH B(2j = BTH A(4) = HT B(4) = ALTH CALL SHPRP(A,B,F12) FS(2,5) = F12
FS(2,4)... DO 510 1=1,4 A(I) = O.DO B(I) = O.DO A(2) = BTH B(2) = BTH A(4) = HT B(4) = HT G = ALTH
CALL SHPRL (A,B,G,F12) FS(2,4) = F12
FS(2,6) = FS(2,5) C USING THE PRINCIPLE OF RECIPROCITY
DO 511 IJ2 = 3,6 511 FS (IJ2,2) = FS(2,IJ2)* RAREA(2)/RAREA(IJ2) C FOR (SURFACE,4)
FS(3,4) = FS(3,2) DO 512 IJ4 =5,6
512 FS(IJ4,4) = FS(IJ4,2) C
FS(3,5) = FS(1,5) FS(3,6) = FS(1,6) FS(4,5) = FS(2,5) FS(4S6) = FS(2,6)
C C FOR FS(5,6)...
DO 513 1=1,4 A(I) = O.DO
513 B(I) = O.DO A(2) = BTH B(2) = BTH A(4) = ALTH B(4) = ALTH G = HT
C CALL SHPRL (A,B,G,F12) FS(5,6) = F12
C FS(6,5) = FS(5,6)
C C C THESE ARE WRITTEN OUT AT THE END C ******** FOR PANELS IN THE CEILING, NCP IN NUMBER, C***** PRINTING OUT RESULTS C TO COMPARE WITH THE EARLIER CASE C WRITE (6,514) (K ,K=1,6) C514 FORMAT (1X/.T35,'WALL-TO-WALL SHAPE FACTORS',5X, C & 'WITHOUT THE CEILING PANELS', //,T15,6(I3,9X) ) C C DO 516 1=1,6 C WRITE (6,515)1,(FS(I,K) ,K=1,6)
C-29
A(2) = BTH B(2) = BTH A(4) = HT B(4) = ALTH CALL SHPRP(A,B,F12) FS(2,5) = F12
C C C FOR FS(2,4)...
DO 510 1=1,4 A(I) = 0.D0
510 B(I) = O.DO A(2) = BTH B(2) = BTH A(4) = HT B(4) = HT G = ALTH
C CALL SHPRL (A,B,G,F12) FS(2,4) = F12
C FS(2,6) = FS(2,5)
C USING THE PRINCIPLE OF RECIPROCITY DO 511 IJ2 = 3,6
511 FS (IJ2,2) = FS(2,IJ2)* RAREA(2)/RAREA(IJ2) C FOR (SURFACE,4)
FS(3,4) = FS(3,2) DO 512 IJ4 =5,6
512 FS(IJ4,4) = FS(IJ4,2) C
FS(3,5) = FS(1,5) FS(3,6) = FS(1,6) FS(4,5) = FS(2,5) FS(4,6) = FS(2,6)
C C FOR FS(5,6)...
DO 513 1=1,4 A(I) = O.DO
513 B(I) = O.DO A(2) = BTH B(2) = BTH A(4) = ALTH B(4) = ALTH G = HT
C CALL SHPRL (A,B,G,F12) FS(5,6) = F12
C FS(6,5) = FS(5,6)
C C C THESE ARE WRITTEN OUT AT THE END C ******** FOR PANELS IN THE CEILING, NCP IN NUMBER, C***** PRINTING OUT RESULTS C TO COMPARE WITH THE EARLIER CASE C WRITE (6,514) (K ,K=1,6) C514 FORMAT (1X/.T35,'WALL-TO-WALL SHAPE FACTORS',5X, C & 'WITHOUT THE CEILING PANELS', //,T15,6(I3,9X) ) C C DO 516 1=1,6 C WRITE (6,515)1,(FS(I,K) ,K=1,6)
C-29
C515 FORMAT (1X/,T5,I3,T15,6(E10.4,2X) ) C516 CONTINUE
RETURN END
C C C INPUT NEEDED: DIMENSIONS OF ROOM :ALTH,BTH,HT & FS(6,6,) C NO. OF PANELS, C THEIR LOCATION : XCP,YCP OF CENTRE & C ALCP,BCP (LENGTH AND BREADTH)
C IMPLICIT REAL*8 (A-H,0-Z) DIMENSION A(4),B(4),FS(6,6) DIMENSION XCP(25),YCP(25),ALCP(25),BCP(25) ,FSCP(25,25) COMMON /AREAS/ RAREA(6) COMMON /PAN2/ PRAREA.PAREA
C /PAN2/ IS COMMON TO M/PROG.,FCN AND SHCP C C C ALL MATRICES FOR PANELS MUST BE DIMENSIONED FOR NCP+ 10 AT THE C LEAST FS = FSCP FOR THE FIRST 6,6 PART. C PANELS ARE NAMED FROM 11 ONWARDS
DATA A/4*0.D0/ , B/4*0.D0/ C ******** SEE HANDOUT FOR BELOW CODING EXPLANATION C CALL ERRSET(263,256,1,1,1) C CALL ERRSET(209,256,1,1,1) C *** 263 ERR NO. (209 ALSO) C *** 256 UNLIMITED NO. OF ERROR OCCURRENCES C *** 1, NO OF MESSAGES TO BE PRINTED C *** 1, NO TRACEBACK IS TO BE PRINTED C *** 1, STANDARD CORRECTIVE ACTION TO BE APPLIED C *** RANGE OF ERROR NUMBERS UPTO THIS NUMBER(OMITTED HERE) C C C C WRITE(6,173) ALTHjBTH.HT C173 FORMAT(1X/,T5,'ROOM DIMENSIONS : ',3X,'LENGTH = ',F6.2, C & 5X,'BREADTH = ',F6.2,5X, 'HEIGHT = ',F6.2) C READ(5,*) NCP C WRITE(6,174) NCP C174 FORMAT(1X/,T5,'TOTAL NUMBER OF CEILING PANELS = ',13) C WRITE(6,175) C175 F0RMAT(1X/,T5,'PANEL CENTER LOCATION :',T30 'X (ALONG LENGTH) ', C & T50,*Y (ALONG BREADTH) ',T70,'LENGTH*,T85,'BREADTH') C C C DO 165 J =11,NCP+10 C READ(5,*)XCP(J),YCP(J),ALCP(J),BCP(J) C WRITE(6,176)J,XCP(J),YCP(J),ALCP(J),BCP(J) C176 F0RMAT(1X/,T20,I3,T30,F6.2,T50,F6.2,T70,F6.2,T85,F6.2,//) C165 CONTINUE C ===== C =================== C c c
RAREA(l) = ALTH * HT RAREA(3) = RAREA(l)
C-30
RAREA(2) = BTH * HT RAREA(4) = RAREA(2) RAREA(5) = ALTH * BTH
WITHOUT CEILING PANELS RAREA(6) = RAREA(5)
WITH CEILING PANELS IS GIVEN JUST BEFORE THE LOOP DO 171
******** FOR PANELS IN THE CEILING, NCP IN NUMBER, ** EACH PANEL HAS ITS CENTER LOCATED AT XCP(J),YCP(J) AND
HAS DIMENSIONS ALCP (LENGTH) AND BCP (BREADTH) THE ORIGIN OF XCP AND YCP IS AT THE INTERSECTION OF WALL1, WALL2 AND THE CEILING.
DO 205 11= 1,25 DO 205 12= 1,25
205 FSCP(I1,I2) = O.DO DO 208 11= 1,6 DO 208 12= 1,6
208 FSCP(I1,I2) = FS(I1,I2)
C WITH CEILING PANELS THE REST OF THE CEILING AREA PRAREA, BECOMES PAREA = O.DO DO 178 K = 11,NCP+10
178 PAREA = PAREA + ALCP(K)"* BCP(K) C NOW PRAREA IS THE AREA OF THE PORTION (OF THE CEILING )
C C207 C
WITHOUT PANELS PRAREA = (ALTH* BTH) - PAREA WRITE(6,207) PRAREA FORMAT(1X/,T5,'REMAINING CEILING AREA,PRAREA, SQFT = \F10.2)
DO 171 J = 11,NCP+10 FOR PANEL NUMBER J,
FOR FSCP(5,J) DO 166 1 = 1 , A(I) = O.DO
166 B(I) = O.DO
B(l) B(2) B(3) B(4)
A(2) A(4)
XCP(J) XCP(J) YCP(J) YCP(J)
ALTH BTH
ALCP(J)/2.DO ALCP(J)/2.D0 BCP(J)/2.DO BCP(J)/2.D0
G = HT
C C C FOR
167
CALL SHPRL (A,B,G,F12) FSCP(5,J) = F12
FSCP(1,J) DO 167 1 = 1 , A(I) = O.DO B(I) = O.DO
B(l) = XCP(J) - ALCP(J)/2.D0 B(2) = XCP(J) + ALCP(J)/2.D0 B(3) = YCP(J) - BCP(J)/2.DO
C-31
B(4) = YCP(J) + BCP(J)/2.D0 C
A(2) = ALTH A(4) = HT
C CALL SHPRP (A,B,F12) FSCP(1,J) = F12
C C FOR FSCP(3,J)
DO 168 I = 1,4 A(I) = 0.D0
168 B(I) = O.DO C
B(l) = XCP(J) - ALCP(J)/2.D0 B(2) = XCP(J) + ALCP(J)/2.D0 B(3) = BTH - ( YCP(J) + BCP(J)/2.D0 ) B(4) = BTH - ( YCP(J) - BCP(J)/2.D0 )
C A(2) = ALTH A(4) = HT
C CALL SHPRP (A,B,F12) FSCP(3,J) = F12
C C C FOR FSCP(2,J)
DO 169 I = 1,4 A(I) = O.DO
169 B(I) = O.DO C
B(l) = YCP(J) - BCP(J)/2.D0 B(2) = YCP(J) + BCP(J)/2.D0 B(3) = XCP(J) - ALCP(J)/2.D0 B(4) = XCP(J) + ALCP(J)/2.D0
C A(2) = BTH A(4) = HT
C CALL SHPRP (A,B,F12) FSCP(2,J) = F12
C C FOR FSCP(4,J)
DO 170 I = 1,4 . A(I) = O.DO
170 B(I) = O.DO C
B(l) = YCP(J) - BCP(J)/2.D0 B(2) = YCP(J) + BCP(J)/2.D0 B(3) = ALTH - ( XCP(J) + ALCP(J)/2.D0) B(4) = ALTH - ( XCP(J) - ALCP(J)/2.D0)
C A(2) = BTH A(4) = HT
C CALL SHPRP (A,B,F12) FSCP(4,J) = F12
C C FOR FSCP(6,J)
FSCP(6,J) = O.DO C C AT THIS STAGE (ALL SURFACES,J) ARE CALCULATED
C-32
C TO CALCULATE, (J,SURFACES) C
DO 179 K = 1,6 FSCP(J.K) = FSCP(K,J)* RAREA(K)/( ALCP(J)*BCP(J) )
179 CONTINUE C C 171 CONTINUE
C WITH THIS ALL PANEL-SHAPE FACTORS ARE CALCULATED. C C IF REMAINING CEILING AREA IS GREATER THAN ZERO
IF ( PRAREA.GT. O.DO ) THEN C TOTAL OF FSCP(WALL OR FLOOR,PANELS IN CEILING) = TFSCP C TO CHANGE THE SHAPE FACTORS(SURFACES,REST OF CEILING)..
DO 181 Kl = 1,5 TFSCP = O.DO DO 180 LI = ll.NCP+10
180 TFSCP = TFSCP + FSCP(K1,L1) FSCP(K1,6) = FSCP(K1,6) - TFSCP
C ANGLE OR SHAPE FACTOR ALGEBRA ..SEE ALGO.FOR BHT.SUBROUTINES FSCP(6,K1) = FSCP(K1,6)* RAREA(K1)/PRAREA
181 CONTINUE C C
ELSE DO 182 Kl = 1,5 FSCP(6,K1) = O.DO
182 FSCP(K1,6) = O.DO ENDIF
C***** PRINTING OUT RESULTS C WRITE (6,192) (K ,K=1,6) C192 FORMAT (1X/,T35,'WALL-TO-WALL SHAPE FACTORS', C &//,T15,6(I3,9X) ) C C DO 193 1=1,6 C WRITE (6,194)I,(FSCP(I,K) ,K=1,6) C194 FORMAT (1X/,T5,I3,T15,6(E10.4,2X) ) C193 CONTINUE C C WRITE(6,195) C195 FORMAT (1X//.T5,'PANEL-TO-WALL SHAPE FACTORS',/) C DO 198 J = ll.NCP+10 C WRITE(6,199) (J,K,FSCP(J,K), K=l,6) C199 F0RMAT(1X/,(T5,'FSCP(P*,I3,I3,') = ',E10.4,/) ) C198 CONTINUE C WRITE(6,413) C413 FORMAT (1X//,T5,'WALL-TO-PANEL SHAPE FACTORS',/) C DO 201 K = 1,6 C WRITE(6,412) (K,J,FSCP(K,J), J=11,NCP+10 ) C412 F0RMAT(1X/,(T5,'FSCP(',I3,' ,P',I3,') = \E10.4,/) ) C201 CONTINUE C C TO COMPARE WITH THE EARLIER CASE C WRITE (6,211) (K ,K=1,6) C211 FORMAT (1X/,T35,* WALL-TO-WALL SHAPE FACTORS',5X, C & 'WITHOUT THE CEILING PANELS', //,T15,6(I3,9X) ) C C DO 212 1=1,6 C WRITE (6,213)1,(FS(I,K) ,K=1,6)
C-33
C213 FORMAT (1X/,T5,I3,T15,6(E10.4,2X) ) C212 CONTINUE C C C
RETURN END
C C C C ** RADIATION SHAPE FACTOR F12 BETWEEN TWO PARALLEL SURFACES C **** INPUT : A(4),B(4) AND G (DISTANCE IN-BETWEEN ) C **** OUTPUT: F12 (SHAPE FACTOR FROM 1 TO 2)
SUBROUTINE SHPRL (A,B,G,F12) IMPLICIT REAL*8 (A-H.O-Z) DIMENSION A(4),B(4) COMMON /COORDS/ Al,Bl,Cl,DltA2,B2,C2,D2
C Al = A(l) Bl = A(2) CI = A(3) Dl = A(4) A2 = B(l) B2 = B(2) C2 = B(3) D2 = B(4)
C G IS OBTAINED AS AN INPUT TO THE SUBROUTINE CALL PARA(G,F12) RETURN END
C C C
SUBROUTINE PARA(G,F12) IMPLICIT REAL*8 (A-H,0-Z) COMMON /COORDS/ A1,B1,C1,D1,A2,B2,C2,D2 COMMON PDIST PDIST = G
RHS1 = PQ(B2-B1,C2-C1) + PQ(B2-B1,D2-D1) & - PQ(B2-B1,C2-D1) - PQ(B2-B1,D2-C1)
RHS2 = PQ(A2-A1,C2-C1) + PQ(A2-A1,D2-D1) & - PQ(A2-A1,C2-D1) - PQ(A2-A1,D2-C1)
RHS3 = PQ(B2-A1,C2-D1) + PQ(B2-A1,D2-C1) & - PQ(B2-A1,C2-C1) - PQ(B2-A1,D2-D1)
RHS4 = PQ(A2-B1,C2-D1) + PQ(A2-B1,D2-C1) & - PQ(A2-B1,C2-C1) - PQ(A2-B1,D2-D1)
RTOTAL = RHS1 + RHS2 + RHS3 + RHS4 C
PI = DATAN(l.DO) * 4.DO ATOTAL = 2.DO * PI *(B1-A1)* (Dl-Cl)
C WRITE(6,*)'A1= \A1 C WRITE(6,*)'B1= \B1 C WRITE(6,*)'C1= \C1 C WRITE(6,*)'D1= ',D1 C WRITE(6,*)'A2= ',A2 C WRITE(6,*)*B2= ',B2 C WRITE(6,*)'C2= *,C2 C WRITE(6,*)'D2= ' D2 C WRITE(6,*) 'G = {,G C WRITE(6,*)'RTOTAL = ' ,RTOTAL, ' ATOTAL = ',ATOTAL
C-34
F12 = RTOTAL/ ATOTAL
RETURN END
C C
FUNCTION PQ( Z1,Z2) IMPLICIT REAL*8 (A-H,0-Z) COMMON PDIST
C V = DSQRT( PDIST**2 + Zl**2 ) W = DSQRT( PDIST**2 + Z2**2 )
C PQ1= Z1*W* DATAN(Z1/W)
C PQ2= Z2*V* DATANCZ2/V)
C PQ3= (PDIST**2)/2 * DLOG( (W**2 +Z1**2)/(W**2) )
C C WRITE(6,*)'PQ1 = ',PQ1 C WRITE(6,*)'PQ2 = ',PQ2 C WRITE(6,*)'PQ3 = ',PQ3 C
PQ = PQ1 +PQ2 - PQ3 -C WRITE(6,*)'PQ = ',PQ
RETURN END
C C C C C RADIATION SHAPE FACTOR F12 , BETWEEN TWO PERPENDICULAR SURFACES C INPUT:A(4),B(4) C OUTPUT : F12 (SHAPE FACTOR FROM 1 TO 2 )
SUBROUTINE SHPRP(A,B,F12) IMPLICIT REAL*8 (A-H,0-Z) DIMENSION A(4),B(4) COMMON /COORDS/ Al,B1,C1,D1,A2,B2,C2,D2
C C
C C
C C C
Al = A(l) Bl = A(2) CI = A(3) Dl = A(4) A2 = B(l) B2 = B(2) C2 = B(3) D2 = B(4)
CALL PERP(F12)
RETURN END
SUBROUTINE PERP(F12) IMPLICIT REAL*8 (A-H,0-Z) COMMON /COORDS/ A1,B1,C1,D1,A2,B2,C2,D2
RHS1 = RS(B2-B1,C2,-C1) + RS(B2-B1,D2,-D1)
C-35
& - RS(B2-B1,C2,-D1) - RS(B2-B1,D2,-C1) RHS2 = RS(A2-A1,C2,-C1) -i- RS(A2-A1,D2,-D1)
& - RS(A2-A1,C2,-D1) - RS(A2-A1,D2,-C1) RHS3 = RS(B2-A1,C2,-D1) + RS(B2-A1,D2,-C1)
& - RS(B2-A1,C2,-C1) - RS(B2-A1,D2,-D1) RHS4 = RS(A2-B1,C2,-D1) + RS(A2-B1,D2,-C1)
& - RS(A2-B1,C2,-C1) - RS(A2-B1,D2,-D1) RTOTAL = RHS1 + RHS2 + RHS3 + RHS4
C PI = DATAN(l.DO) * 4.DO
ATOTAL = 2.DO * PI *(B1-A1)* (Dl-Cl) C WRITE(6,*)'A1= ' ,A1 C WRITE(6,*)'B1= ',B1 C WRITE(6,*)'C1= ',C1 C WRITE(6,*)'D1= ',D1 C WRITE(6,*)'A2= ',A2 C WRITE(6,*)'B2= ',B2 C WRITE(6;*)'C2= ',C2 C WRITE(6,*)'D2= ',D2 C WRITE(6,*)'RTOTAL = *,RTOTAL, * ATOTAL = *,ATOTAL
F12 = RTOTAL/ ATOTAL
END C C
FUNCTION RS( Z1,Y2,Y1) IMPLICIT REAL*8 (A-H,0-Z) COMMON /HT/ G
C T'= DSQRT( Y2**2 + Yl**2 )
C SPECIAL CASE: WHEN EITHER T OR Zl IS EQUAL TO ZERO, IF (T .EQ. O.DO .OR. Zl .EQ. O.DO ) THEN
RSI = O.DO ELSE
RS1= T*Z1* DATAN(Z1/T) ENDIF
C SPECIAL CASE : WHEN BOTH Zl AND T ARE EQUAL TO ZERO, IF (Zl .EQ. O.DO .AND. T .EQ. O.DO ) THEN
RS2 = O.DO ELSE
RS2= 0.25D0*( Zl**2 -T**2) * DLOG (T**2 + Zl**2 ) ENDIF
C C WRITE(6,*)'PQ1 = ',PQ1 C WRITE(6,*)'PQ2 = ',PQ2 C WRITE(6,*)'PQ3 = ',PQ3 C
RS = RSI +RS2 C WRITE(6,*)'RS = \RS
RETURN END
C-36
APPENDIX D
[
tea
REPRODUCTION OF CHAPTER 8
FROM 1984 ASHRAE SYSTEMS
HANDBOOK
D-l
CHAPTER 8
PANEL HEATING AND COOLING SYSTEMS
System Types; System Concepts; System Applications; Heat Transfer by Panel Surfaces; Panel Heating and Cooling Systems; General Design Considerations;
Panel Heating System Design; Panel Cooling System Design
RADIANT panel systems combine controlled temperature room surfaces with central station air conditioning. The
controlled temperature surfaces may be in the floor, walls or ceiling, and the temperature is maintained by circulating water, air or electric resistance. The central station air system can be a basic, one-zone, constant temperature, constant volume system; or it can include some or all of the features of dual-duct, reheat, multizone or variable volume systems. A controlled temperature surface is referred to as a radiant panel if 50% or more of the heat transfer is by radiation to other surfaces seen by the panel. This chapter is concerned with surfaces whose temperatures are controlled and are the'primary source of heating and cooling within the conditioned space.
High temperature surface radiant panels [over about 250 F (121°C)] energized by gas, electricity or high temperature water are discussed in Chapter 18.
SYSTEM TYPES
Residential heating applications usually consist of pipe coils embedded in masonry floors or plaster ceilings. This construe^ tion is suitable where loads are stable and solar effects are minimized by building design. However, in buildings where glass areas are large and load changes occur faster, the slow response, lag and override effect of masonry panels are unsatisfactory. Lightweight metal panel ceiling systems respond quickly to load changes and can be used for cooling and heating.
Warm air and electric heating elements are two design concepts used in systems influenced by local factors. The warm air system has a special cavity construction where air is supplied to a cavity behind or under the panel surface. The air leaves the cavity through a normal diffuser arrangement and is supplied to the room. Generally, these systems are used as floor radiant panels in schools and in floors subject to extreme cold, such as in an overhang. Cold outdoor and heating medium temperatures must be analyzed with regard to potential damage to the building construction. Electric heating elements embedded in the floor or ceiling construction and unitized electric ceiling panels are used in various applications for local spot heating as well as for providing full heating requirements for the space.
Radiant panels are usually located in the ceiling because it is exposed to all other surfaces and objects in the room. Because it is not likely to be covered, as are the floors, higher surface temperatures can be used. Also, its smaller mass enables it to respond more quickly to load changes. Radiant cooling can be incorporated, and, in metal ceiling systems, the piping is accessible for servicing.
The ceiling panel systems commonly used today are an out
growth of the perforated metal, suspended, acoustical ceiling. These radiant ceiling systems are usually designed into buildings where the features of the suspended acoustical ceiling can be combined with panel heating and cooling. The panels can be designed as small units to fit the building module and provide extensive flexibility for zoning and control; or the panels can be arranged as large continuous areas for maximum economy.
Three types of metal ceiling systems are available. One consists of lightweight aluminum panels, usually 12 x 24 in. (300 x 600 mm), attached in the field to 0.5-in. (12.7-mm) galvanized pipe coils. The second consists of a copper coil metallurgically bonded to the aluminum face sheet to form a modular panel. Modular panels are available in sizes up to about 36 x 60 in.(900 x 1525 mm) and are held in position by various types of ceiling suspension systems. The third type is an aluminum extrusion face sheet with a copper tube pressed into an oval channel on the back of the face sheet. Extruded panels can be manufactured in almost any shape and size. Practical limitations dictate a maximum size of about 16x4 ft (4.88 x 1.22 m).
Radiant panel systems are similar to other air-water systems in the arrangement of the system components (see Fig. 1). Room thermal conditions are maintained primarily by direct transfer of radiant energy, rather than by convection heating and cooling. The room heating and cooling loads are calculated in the conventional manner. Manufacturers' ratings gen-
The preparation of this chapter is assigned to TC 9.1, Radiant Space Healing & Cooling.
Fig. 1 Primary/Secondary Water Distribution System with Mixing Control
8.1
D-2
8.2 CHAPTER 8 1984 Systems Handbook erally are for total performance and can be applied directly to the calculated room load.
Principal advantages of panel systems are: 1. Comfort levels are better than those of other conditioning sys-
tems-because radiant loads are treated directly and air motion in the space is at normal ventilation levels.
2. Mechanical equipment is not needed at the outside walls, simplifying the wall, floor and structural systems.
3. AH pumps, fans, filters and so forth are centrally located, simplifying maintenance and operation.
4. Cooling and healing can be simultaneous, without central zoning or seasonal changeover, when three- and four-pipe systems are used.
5. Supply air quantities usually do not exceed those required for ventilation and dehumidification.
6. The occupied space has no mechanical equipment requiring maintenance or repair.
7. Draperies and curtains can be installed at the outside wall without interfering with the heating and cooling system.
8. The modular panel concept provides flexibility to meet changes in partitioning.
9. A 100% outdoor air system may be installed with less severe penalties in terms of refrigeration load because of reduced air quantities.
10. No space is required within the air-conditioned room for the mechanical equipment. This feature is especially valuable when compared to other conditioning methods in existing buildings, hospital patient rooms and other applications where space is at a premium, where maximum cleanliness is essential or where dictated by legal requirements.
11. A common central air system can serve both the interior and perimeter zones.
12. Wet surface cooling coils are eliminated from the occupied space, reducing the potential for septic contamination.
13. The panel system can use the automatic sprinkler system piping. (See NFPA 13-1982, Chapter 5, Sections 5-6). The maximum water treatment must not fuse the heads.
Other factors to consider when using panel systems are: 1. Early evaluation is necessary to use the panel system to full ad
vantage in optimizing the physical building design. 2. Recessed lighting fixtures, air diffusers, hung ceilings and other
ceiling devices must be selected on the basis of providing the maximum ceiling area possible for use as radiant panels.
3. The air-side design mast be able to maintain humidity levels at or below design conditions at all times to eliminate any possibility of condensation on the panels. This becomes more critical if space dry- and wet-bulb temperatures are allowed to drift as an energy conservation measure.
4. Cooling, panels should not be used in or adjacent to high humidity areas.
5. As with any hydronic system, the piping system should be designed to avoid noises from entrained air, high velocity or high pressure drop devices or from pump and pipe vibrations.
6. Thermal expansion of the ceiling and other devices in or adjacent to the ceiling should be anticipated.
7. Operable sash should be designed to discourage unauthorized operation.
SYSTEM CONCEPTS
All bodies with a surface temperature above absolute zero emit rays with wavelengths depending on the body surface temperature. Every facet of the surface emits rays in straight lines at right angles to the facet. When examined under a microscope, the surface of concrete or rough plaster is covered with numerous facets, each giving off radiant energy. Polished steel or similar polished surfaces show no such facets. Thus, a rough surface emits heat rays more efficiently than a polished surface.
The invigorating effect of radiant heat is experienced when the body is exposed to the sun's rays on a cool but sunny day in spring. Some of these rays impinging on the body come directly from the sun and include the whole range of ether
waves. Other rays coming from the sun impinge on surrounding objects, where they are increased in wavelength and reflected to the body as low temperature radiation, producing a comfortable feeling of warmth. Should a cloud pass over the sun, instantly there is a sensation of cold; although in such a short interval, the air temperature does not vary at all.
In searching for the correct conditions compatible with the physiological demands of the human body, no system can be rated as completely satisfactory unless it satisfies the three main factors controlling heat loss from the human body: radiation, convection and evaporation. It is sometimes thought that a radiant heat system is desirable only for certain buildings and only in some climates. However, wherever people live, these three factors of heat loss must be considered. It is as important to provide the correct conditions in very cold climates as it is in moderate climates. Maintaining the correct comfort conditions by low temperature radiation is possible for even the most severe weather conditions.
Panel heating and cooling systems function to provide a comfortable environment by controlling surface temperatures and minimizing excessive air motion within the space. Thermal comfort, as defined by ASHRAE Standard 55-198/,' is "that condition of mind which expresses satisfaction with the thermal environment." A person is not aware that his environment is being heated or cooled. Recent study has given us better insight on the human body and its response to the surrounding environment. The mean radiant temperature (MRT) strongly influences the feeling of comfort. When the surface temperature of the outside walls, particularly those with large amounts of glass, begins to deviate excessively from the ambient air temperature of the space, it is increasingly difficult for convective systems to counteract the discomfort resulting from cold or hot walls. Heating and cooling panels neutralize these deficiencies and minimize excessive radiation losses from the body.
Unlike most heat transfer equipment where performance can be measured in specific terms, the performance of the radiant panel is related directly to the structure in which it is located, and an evaluation of this interrelationship is desirable. Research and testing of panel performance have been conducted by various independent researchers and manufacturers. Heat transfer between the radiant panel and the other room surfaces is well established in a boxlike room where the primary heat gains and losses are from the wall, floor or ceiling surfaces. The performance ratings presented in this chapter for radiation and convection can be applied directly to the calculated room heating and cooling loads. Various investigators and manufacturers report increased cooling performance because of solar effects and ceiling-mounted lighting fixtures. This empirical information, which has been developed as a result of field testing, should only be used in consultation with manufacturers experienced in this field.
Fortunately, most building surfaces have high emissivity factors and therefore absorb and reradiate energy from the active panels. This is significant because all surfaces within the room tend to assume an equilibrium temperature resulting in an even thermal comfort condition within the space. In much the same way that light energy from a lighting fixture illuminates the room so that all surfaces can be seen, a warm radiant panel emits energy that is absorbed and reradiated, and all surfaces become warm. Warm ceiling panels are effective for winter heating because they warm the floor and glass surfaces by direct transfer of radiant energy. The surface temperature of well constructed and properly insulated floors will be 2 to 3 deg F (1 to 2°Q above the ambient air temperature, and the inside surface temperature of glass is increased significantly. [Inside single-glass surface temperatures
D-3
Panel Heating and Cooling Systems 8.3
10 to IS deg F (5 to 8°C) above those indicated in Fig. 7 are commonly observed.] As a result, downdrafts are minimized to the point where no discomfort is felt. Installation with ceiling heights of SO ft (IS m) and single glass from floor to ceiling provide satisfactory results.
SYSTEM APPLICATIONS
Office Buildings The panel system is usually applied as a perimeter system
providing heating and cooling. A single-zone central air supply system provides ventilation air, dehumidification and usually some sensible cooling. Often, tempered air is supplied at a constant volume, and the room thermostat modulates the panel output. In some applications, the panels are arranged for zone control, and the air system is designed to provide individual room control. Water distribution systems using the two- or four-pipe concept may be used. Panel systems are readily adaptable to accommodate most changes in partitioning. Installations can be made where complete flexibility is on a modular basis. Electric panels in lay-in ceilings have been used for full perimeter heating.
Schools Panels are usually selected for heating and cooling, or for
heating only, in all areas except gymnasiums and auditoriums. For heating only applications, the system may be used with any type of approved ventilation system. The panel-system is usually sized to offset the transmission loads plus any reheating of the air required. Room control is accomplished by modulating the water flow through the panel. If the school is air conditioned by a central air system and has perimeter heating panels, a single-zone piping system might be used to control panel heating output, and the room thermostat would modulate the supply temperature or supply volume of air delivered to the room. Heating and cooling panel applications are similar to office buildings. Another advantage of panel heating and cooling for classroom areas is that mechanical equipment noise does not interfere with instructional activities.
Hospitals The principal application of radiant panel systems over the
past 30 years has been for hospital patient rooms. This system is well suited because it: (1) provides a draft-free, thermally stable environment, (2) requires no mechanical equipment or bacteria and virus collectors in the space requiring maintenance and (3) does not take up space within the room. Individual room control is usually by throttling the water flow through the panel. The air supply system is often a 100% outdoor air system, and minimum air quantities delivered to the room are those required for ventilation and exhaust of the toilet room and soiled linen closet. The piping system may have a two- or four-pipe design. Water control valves should be in the corridor outside the patient room so that they can be' adjusted or serviced without entering the room. All piping connections above the ceiling should be soldered or welded and thoroughly tested. If cubicle tracks are applied to the ceiling surface, track installation should be coordinated with the radiant ceiling. Panel ceilings are often used in areas of the hospital occupied by mentally disturbed patients since no equipment is accessible to the occupant for destruction or self-inflicted injury.
Swimming Pools Panel heating systems are well suited to swimming pools be
cause the partially clothed body emerging from the water is very sensitive to the thermal environment. Floor panel tem
peratures are restricted so as not to cause foot discomfort. Ceiling panels are generally located around the perimeter of; the pool, not directly over the water. Panel surface temperatures are higher to compensate for the increased ceiling height and to produce a greater radiant effect on the partially clothed body.
Apartment Buildings For heating, pipe coils are embedded in the masonry slab.
The coils must be carefully positioned so as not to overheat one apartment when maintaining desired temperatures in another. The slow response of embedded pipe coils in buildings with large glass areas may prove to be unsatisfactory. Installations for heating and cooling have been made with pipes embedded in a hung plaster ceiling. A separate minimum volume dehumidified air system provides the necessary dehumidification and ventilation for each apartment. In recent years, there has been an increased application of electric resistance elements embedded in the floor or behind a skimcoat of plaster at the ceiling. The electric panels are easy to install and have the advantage of simplified individual room control.
Residences Embedded pipe coil systems, electric resistance panels and
forced warm air panel systems have all been used. The embedded pipe coil system is the most common, using grid coils in the floor slab or copper tubing systems in older plaster ceilings. These systems are well suited to normally constructed residences with normal glass areas. Lightweight metal panel ceiling systems have been applied to residences. Prefabricated electric panels have also proved advantageous, particularly in add-on rooms.
Industrial Applications Panel systems have found wide application in general space
heating for industrial buildings in Europe. However, there has been only a limited application of this type in the Western Hemisphere. With the increasing demand for worker comfort, panel systems should be considered. For example, one special application is an internal combustion engine test cell, where the walls and ceilings are cooled with chilled water. Although the ambient air temperature in the space ranges up to 95 F (35°C), the occupants work in relative comfort when 55 F (13°C) water is circulated through the ceiling and wall panels.
Other Building Types Metal panel ceiling systems can be operated as heating sys
tems at elevated water temperatures, and have been used in airport terminals, convention halls, lobbies, museums and especially where large glass areas are involved. Cooling may also be applied. Because radiant energy travels through the air without wanning it, ceilings can be installed at any height and remain effective. The highest ceiling installed for a comfort application is SO ft (IS m) above the floor with a panel surface temperature of approximately 285 F (141°C) for heating. The ceiling panels offset the heat loss from a single-glazed all-glass wall.
The high lighting levels in television studios make them well suited to panel systems. The panels are installed for cooling only and are placed above the lighting system to absorb the radiation and convection heat from the lights and normal heat gains from the space. Besides absorbing heat from the space, the panel ceiling also improves the acoustical properties of the studio.
Metal panel ceiling systems are also installed in minimum and medium security jail cells and other areas where disturbed occupants are housed. The ceiling construction is made more rugged by increasing the gauge of the ceiling panels and using
D-4
8.4 CHAPTER 8 1984 Systems Handbook security dips so that the ceiling panels cannot be removed. Part of the perforated metal ceiling can be used for air distribution.
HEAT TRANSFER BY PANEL SURFACES A heated or cooled panel transfers heat to or from a room
by convection and radiation. In the following paragraphs, the two transfer mechanisms are first considered separately and then combined to facilitate design calculations.
Radiation Transfer The basic equation for radiation exchange is the Stefan-
Boltzmann equation (see Chapter 2 of the 1981 FUNDAMENTALS VOLUME). This equation may also be written as:
q, =0.ni3FaFe [(7V/100)4 -{T„/\00)A] (1) where
qr = heat transferred by radiation, Btu/h«ft2 (W/m2). Tr = mean radiant temperature of unheated surface, F (°C) abs. Tp = average surface temperature of heated panel, F (°C) abs. Fa = the configuration factor (dimensionless). Fe = the emissivity factor (dimensionless).
0.1713 = Stefan-Boltzmann radiation constant, Btu/h«ft2«F (5.6697 x in-8 W/m2 - °Q absolute temperature to the fourth power.
Where several surfaces exposed to the panel have widely differing temperatures, it may be necessary to-compute the area-weighted Average Unheated (or Uncooled) Surface Temperature (AUST) exposed to the panels. In confined situations or special applications, such as shipboard berthing spaces with an adjacent hot gas stack, or in situations where the emissivity of the surfaces is significantly different, it is necessary to evaluate each surface using the geometrical factors from the charts in Chapter 2 of the 1981 FUNDAMENTALS VOLUME. Room related angle and shape factors can also be found in Fanger's book on thermal comfort2 or from the algorithms in ASHRAE Energy Calculations 1,1976.
Similarly, when considering spot cooling it may be necessary to consider the influence of gaseous radiation.3 For normal application, these refinements are generally insignificant. The Hohel equation [Eq. (2)] is used frequently. This equation assumes a simple, boxlike room in which there is a uniformly heated ceiling, floor or wall; all other surfaces are at another temperature; and all surfaces are perfectly diffusing.
F =F F = (2) c ' r ' l/F1.2 + [( l /e I)- l )+yi l />l2[( l /e2)- l ]
where Fc = combined configuration and emissivity factor.
F|.2 = view factor = 1.0. et ande2
= emissivities of the surfaces. A i and .42
= areas of the surfaces. In practice, the emissivity of nonmetallic or painted metal
nonreflecting surfaces is about 0.9. When this emissivity is used in Eq. (2), the combined factor is about 0.87 for most rooms. Substituting this value in Eq. (1), the constant becomes about 0.1S, and the equation for heating can be rewritten:
or for cooling:
0 O 20 JO 40 SO 60 TO 60 90 100 MO 120 ISO WO ISO HEM OUTPUT. BTUHPERSO'T
Fig. 2 Heat Transferred by Radiation from a Heated Ceiling, Floor or Wall Panel
where qr =heat transferred by the panel to or from the room surfaces by
radiation, Btu/h«ft2 (W/m2). tp =the average panel surface temperature, F (°C).
AUST =area-weighted average temperature of the unheated surfaces in the room, F(°C).
0=460(273). The actual radiation transfer in a room may be somewhat
different from that given by Eq. (3) or (3a) because of nonuniform temperatures, irregular room surfaces, variations in emissivity of materials and so forth. It is generally agreed, however, that the equation is accurate to within 10% when used in conventional heating and cooling calculations. Tests4
show that the value of the constant of Eq. (3) and (3a) was 0.1S2 in the test room. The design information in this chapter is based on that constant value.
Radiation exchange calculated from Eq. (3) is given in Fig. 2. The values apply to ceiling, floor or wall panel output.
Radiation removed by a cooling panel for a range of normally encountered temperatures and as calculated from Eq. (3a), which is a variation of Eq. (3), is given in Fig. 3. In many specific instances where normal multistory commercial construction and fluorescent lighting are used, the room temperature at the 5-ft (1.5-m) level will closely approach the AUST (Average Uncooled Surface Temperatures). In structures where the main heat gain is through the walls or where incandescent lighting is used, the wall surface temperatures tend to rise considerably above the room air temperature.
Convection Transfer The convection coefficient qe is defined as the heat trans
ferred by convection in Btu/h«ft2-F (W/m2-°C) difference between air and panel temperatures. Heat transfer convection values are not easily established. Convection in panel systems is usually considered to be natural; that is, air motion is generated by the warming or cooling of the boundary layer of air which starts moving as soon as its temperature rises above or drops below the surrounding air temperature. In practice, however, there are many factors that interfere with or affect natural convection. The configuration of the room and the spaces determines the natural convection. Infiltration, the movement of persons and mechanical ventilating systems can
D-5
Panel Heating and Cooling Systems 8.5
V'EW FAC SUM OF At
' O R . I 0 L SURFACE
\«/ .
N>. X '
XV. X
S-090
fc
CONVf RSION FACTORS:
SO 60 70 60 90 PANEL SURFACE TEMPERATURE , F
Fig. 3 Heat Removed by Radiation to a Cooled Ceiling or Wall Panel
introduce some forced convection that will disturb the natural process.
The effect of forced convection on heat transfer from panels has been reported3 as an increment to be added to the natural convection coefficient. However, increased heat transfer from forced convection should not be used, because -the increments are unpredictable in pattern and performance and do not significantly increase the total capacity of the panel system.
The convection in a panel system is a function of the panel surface temperature and the temperature of the airstream layer directly below the panel. The most consistent results are obtained when the air layer temperature is measured close to the region where the fully developed stream begins, usually 2 to 2.5 in. (SI to 64 mm) below the panels. Very little heat transfer literature describes experiences pertinent to this application, although some of the pioneer work has been done.6
Research4 has determined natural convection coefficients referred to the center of the space 5 ft (1.5 m) above the floor in a 12 x 24 ft (300 x 600 mm) room. Equations (4) to (9), derived from this research, can be used to calculate heat transfer from panels by natural convection.
Natural convection from a heated ceiling
qc = 0.041 (f„-/,)'•"/£>,"•* (4) Natural convection from a heated floor or cooled ceiling
9 c=0.39(/ p - / f l ), :"/£>, 0 0 8 (5)
Natural convection from a heated or cooled wall panel qc = 0.29{tp - tay
M/H00i (6) "where qe = heat transfer by natural convection, Btu/h» ft2 (W/m2). tp = temperature of panel surface, F (°C). ta = temperature of the air, F (°Q.
De = equivalent diameter of panel (area X4T perimeter), ft(m). H «* height of wall panel, ft (m).. Measurements of panel performance in furnished test
rooms that did not have uniform temperature surfaces showed variations that are not large enough to be significant in heating practice.7 Other tests8 established that the effect of room size was also usually insignificant. The convection equations can therefore be simplified to:
eo
7S
6S
b "
s5 0
= 4» I I I a in 3 1-
! «
IS
s
1
1 •
1 i i
1 1 COMVIRSIOKI FACIIM3.
~ c . * , F / i . a Wta1 > BWft I r X l B
-
—
—
...
...
—
.....
^
-
.....
—
...
---
—
—
_~
._.
—
.-.*-
...
f* /
- '
•
1
1
1 1 f i y
/ /
/ / /
.,r. • i *a
j
V / /
/ 1
/
* - * & •
rTLT criiMiB WMtia-V
1 i 1
/
/ >
iao»"o»j
/.
_.
—
—
0 S 10 IS 20 25 30 31 40 45 SO SS 60 65 70 TEMP. OF PANEL SURFACE MINUS TEMP. OF AIR (Ip - lo) - F
Fig. 4 Heat Output by Natural Convection from Floor and Ceiling Heating Panels
Natural convection from a heated ceiling qc = 0.021 (tp - ta)
125
Natural convection from a heated floor or cooled ceiling qe = 0 .32 ( / p - / J , : "
Natural convection from a heated or cooled wall panel qr = 0.26 (t„-tay»
Figure 4 shows heat output by natural convection from floor and ceiling heating panels as calculated from Eq. (7) and (8).
Figure 5 shows heat removed by natural convection by
(7)
(8)
(9)
TEMPERATURE DIFFERENCE TO PANEL SURFACE-F DES
Fig. 5 Heat Removed by Natural Convection to Ceiling Cooling Panels .
D-6
8.6 CHAPTER 8 1984 Systems Handbook
40 SO 20 IO O -10 - 2 0 OUTDOOR AIR TCMPCRATURC -OCCRCC FAHR.
[Indoor air Hmptraturo —70 f\
Fig. 6 Relation of Inside Surface Temperature to Overall Coefficient of Heat Transfer
cooled ceiling panels as calculated by Eq. (8) and data6 for specific panel sizes. An additional curve is shown to illustrate the effect of forced convection on the latter data. Similar adjustment of the ASHRAE data is not exactly appropriate, but the effects would be of the same order of magnitude. As a preliminary basis for design in Fig. 5, use 1 Btu/h«ft2«F (5.68 W/m2«°C) of temperature difference between room design and panel temperature.
Combined Heat Transfer (Radiation and Convection) The combined heat transfer from a panel to a room can be
determined by adding the radiant heat transfer from Fig. 2 or 3 to the convective heat transfer from Fig. 4 or 5, respectively. Use of Fig. 2 and 3 requires calculating the AUST in the room. In calculating the AUST, the surface temperature of
O 0.1 0.2 0.3 O * 0.5 0.6 0.7 OB Ofl 1.0 IJ 1.2 OVEKAU. COEFFICIENT OF HEAT TRANSFER
U - B T U PER ( H « ) ( S Q F f X f DEC)
Fig. 7 Inside Wall Surface Temperature Correction for Air Temperatures Other Than 70 F
the inside walls is assumed to be the same as the room air temperature. The surface temperatures of outside walls and exposed floors or ceilings for heating panel calculations can be obtained from Fig. 6 for a 70 F (21°C) room air temperature. Corrections for other temperatures may be obtained from Rg.7.
The combined heat transfer for ceiling and floor panels when used for heating in rooms in which the air temperature is 70 to 76 F (21 to 24°Q can be read directly from Fig. 8 and 9, respectively. These two diagrams apply to rooms in which the AUST does not differ greatly from room air temperatures. Tests9,10 show that the temperatures are almost equal.
The combined radiation and convection transfer for cooling, as given in Fig. 3 and 5, is shown in Fig. 10. The data in Fig. 10 do not include heat gains from sun, lights, people or equipment. Refer to the manufacturer's data to include these heat gains.
In suspended ceiling panel systems, heat can be transferred from the ceiling panel to the floor slab above (heating) and vice versa (cooling). The ceiling panel surface temperature is affected because of heat transfer to or from the panel and the slab by radiation and, to a much smaller extent, by convection. The radiation component can be approximated using Fig. 1. The convection component can be approximated using Fig. 2 or 3. In this case, the temperature difference used is that between the top of the ceiling panel and the midspace of the ceiling. Theoretically, the temperature of the ceiling space should be determined by testing, since it varies with different types of panel systems. However, much of this heat transfer is nullified with the application of insulation over the ceiling panel, which, for perforated metal panels, also provides acoustical control.
If lighting fixtures are recessed into the suspended ceiling space, radiation from the top of the fixtures will raise the overhead slab temperature and will transfer heat to the ceiling space by convection. This energy will be absorbed at the top of the cooled ceiling panels by radiation, as in Fig. 3, and by convection, generally in accordance with Eq. (4). The amount the top of the panel absorbs depends on the system type. Most system manufacturers have empirical information available. Similarly, panels'installed under a roof will absorb additional heat, again depending on configuration and insulation.
Panel Thermal Resistance The thermal resistance to heat flow may vary considerably
among panel systems, depending on the type of bond between the water tube and the panel material. This bond may change with time, corrosion between lightly touching surfaces, method of maintaining contact and other factors. The actual thermal resistance of any proposed system should be verified by testing whenever practicable. Tables 1 through 4 show some typical values for thermal resistance factors for various types of floor and ceiling panels.
Effect of Floor Coverings Floor coverings can have a pronounced effect on the perfor
mance of a floor heating panel system. The added thermal resistance of the floor covering reduces upward heat flow and increases the heat flow to the underside of the slab. To maintain a given upward heat flow after a floor covering has been added, the temperature of the heating medium must be increased. Data on the thermal resistance of common floor coverings are given in Table 5.
Where covered and bare floor panels exist in the same system, it may be possible to maintain a high enough water temperature to satisfy the covered panels and balance the system by throttling the flow to the bare slabs. In,some instances, however, the increased water temperature required when car-
D-7
Panel Heating and Cooling Systems 8.7
'X
c
1. I
.?
1 O
UIH
I
Cm
I 8
i "
*/
jf-
€%-
_*z
^ ^
L*i
^ ^
UA
i
L ^
^ S
Lftd
^ - ^
L*J
* *
^
L«4
^ j i ^
• ^
LsA
>sf^ P ^ ^ •^s* ***^.
^
r^
/ / /
^ - ^
X . / ' .
^
- — - ^ 1 ,
/
J r o
• I3>
, , /
^~
y
CONV *C *
o/ . -C
y X
ERStON FACTORS ( F - 3 W - B
iW e tt'-F-tVBtu * (t value)
.#/ -*A-?A-#A-$A-&L$A-&L?A-*A. M.M. *l
ttiffi
/ /
/ L^.^Z
S00^
/ /
/
SURFACE OR MEAN WATERTEMPERATURE — F
Fig. 8 Ceiling Panel Design Graph Showing Panel Surface Temperature and Mean Water Temperature vs. Output Downward
^^VV^VVVVv^^^VV^VVsVV^^^J^i^' SURFACE OR MEAN WATER TEMPERATURE — F
Fig. 9 Floor Panel Design Graph Showing Pane) Surface Temperature and Mean Water Temperature vs. Output Upward
peting is applied over floor panels makes it impossible to bal- peting, unless the pipe is arranged to permit zoning using ance floor panel systems in which only some rooms have car- more than one water temperature.
D-8
8.8 CHAPTER 8 1984 Systems Handbook
i 5
, sx\
\ ^
1 1 I CONVERSION FACTORS:
W/m'-BtuA
l\ • V
SA.
•ft'xais
70 75 CEILING PANEL TEMPERATURE,F
Fig. 10 Performance of a Cooled Ceiling Panel (Uniform Environment, No Infiltration, No Internal Heat
Sources)
Panel Heat Losses Heat transferred from the upper surface of ceiling panels,
the back surface of wall panels, the underside of floor panels or the edges of any panel is considered a panel heat loss. Panel
Table 1 Thermal Resistance of Bare Concrete Floor Panels (Heating)*
Panel Coustracuon Spacing, hi.
Thermal Resistance, 11* • F • h/Btu (m2 .°C/W)
Heal Flow Hallo. qu/arf org,,/qa,~
44a. (101.6 mm) Concrete Slab— 24n. ($0.8 mm) Cover
0.5-in(l2.7mm) (nom.) ponferrous tube
(nun)
9 (228.6)
12 (304.8)
up
'us
0.57 (0.10) 0.73 (0.13)
1 3
dowa up down
'ds 'us 'ds
0.52 0.46 0.84 (0.09) (0.08) (0.15)
0.68 0.58 1.16 (0.12) (0.10) (0.20)
5
np down
'us 'ds
0.43 1.17 (0.08) (0.21) 0.54 1.65 (0.10) (0.29)
10
up dowo
'us 'ds
0.42 1.97 (0.07) (0.34) 0.51 2.86 (0.09) (0.50)
<L5-ln. (12.7 ram) (nom.) ferrous pipe or0.7$-in. (19.1 mm) (nom.) non-ferrous tube
9 0.49 0.42 0.41 0.66 0.39 0.90 0.38 1.80 (228.6) (0.09) (0.07) (0.07) (0.12) (0.07) (0.16) (0.07) (0.32)
12 0.63 0.55 0.50 0.93 0.48 1.30 0.46 2.35 (304.8) (0.11) (0.10) (0.09) (0.16) (0.08) (0.23) (0.08) (0.41)
6-tn. (152.4 mm) Concrete Slab— M a . (SO.8 mm) Oner
0.5-ta.(12.7mm) (nom.) nonferrous tube
0.754n(l9.lmm) (nom.) nonferrous tube
0.7S4n.(l9.1mm) (nom.) ferrous pipe
9 (228.6)
12 (304.8)
9 (228.6)
12 (304.8)
9 (228.6)
12 (304.8)
0.59 (0.10) 0.78 (0,14)
0.51 (0.09) 0.68 (0.12)
0.47 (0.08) 0.63 (0.11)
0.70 0.47 (0.12) (0.08)
0.90 0.60 (0.16) (0.11)
0.61 0.43 (0.11) (0.08)
0.78 0.54 (.14) (0.10)
0.55 0.40 (0.10) (0.07)
0.71 0.50 (0.12) (0.09)
1.05 0.45 (0.18) (0.08)
1.40 0.56 (0.25) (0.10)
0.87 0.41 (0.15) (0.07)
1.23 0.51 (0.22) (0.09)
0.77 0.39 (0.14) (0.07)
1.07 0.48 (0.19) (0.08)
1.39 (0.24)
1.97 (0.35)
1.13 (0.20)
1.63 (0.29)
0.98 (0.17)
1.44 (0.20)
0.43 (0.08) 0.54 (0.10
0.40 (0.07) 0.49 (0.09)
0.38 (0.07) 0.46 (0.08)
2.25 (0.40)
3.21 (0.56)
1.78 (0.31)
2.61 (0.46)
1.50 (0.26)
2.36 (0.42)
1-to. (25.4 mm) (nam.) nonferrous tubcorl-in. (nom.) ferrous pipe
12 0.59 0.66 0.48 0.98 0.46 I JO 0.45 2.11 (304.8) (0.10) (Q.12) (0.08) (0.17) (0.08) (0.23) (0.08) (0.37)
15 0.73 0.83 0.57 1.21 0.54 1.73 0.SI 2.74 (381) (0.13) (0.15) (0.10) (0.21) (0.10) (0.J0) (0.09) (0.48)
Table 2 Thermal Resistance of Concrete Ceiling Panels (Heating)
_r Panel Construction
THICKNESS
Thermal Resistance, f |2 .F.b/Bto (m*."C/W
PANEL .SURFACE ,
Spacing. In.
Kent Flow Ratio.1 qulqd 03 TO
• COVER—"
«-ta. (152.4 mm) Coacnrie Slab— 14a. (25.4 ma) Cover
0.5-in.(l2.7mm)(nom.) ROnfcrroustube
0.5-in. (12.7 mm) (nom.) ferrous pipe or 0.75-in. (nom.) non-ferrous tube
up down
'us 'd
9 3.6 0.30 (228.6) (0.63) (0.05)
12 5.1 0.35 (304.8) (0.90) (0.06)
9 2.6 0.25 (228.6) (0.46) (0.04)
12 4.0 0.30 (304.8) (0.70) (0.05)
up down
'us 'd
0.9 0.35 (0.16) (0.06) 1.1 0.45 (0.19) (0.08)
0.7 0.30 (0.12) (0.05) 0.9 0.40 (0.16) (0.07)
np down
'us 'd
0.7 0.45 (0.12) (0.08) 0.9 0 J 5 (0.16) (0.10)
0.6 0.35 ((0.ll)(0.06> 0.8 0.50 (0.14) (0.08)
0.75-in.(l9.l mm)(nom.) ferrous pipe or Inn. (25.4 mm) (nom.) non-ferrous tube
9 2.1 0.20 (228.6) (0.37) (0.04)
12 3J 0.30 (304.8) (0.58) (0.05)
IS 4J 0.35 (381) (0.79) (0.06)
0.6 0.25 (0.11) (0.04) 0.8 0.35 (0.14) (0.06) 1.0 0.45 (0.18) (0.08)
0.6 0.30 (0.11) (0.05) 0.7 0.40 (0.12) (0.07) 0.8 0.55 (0.14) (0.10)
1-in. (25.4 mm) (nom.) ferrous pipe
9 1.6 0.20 (228.6) (0.28) (0.04)
12 2.6 0.25 (304.8) (0.46) (0.04)
15 3.6 0.30 (381) (0.63) (0.05)
0.5 0.25 (0.09) (0.04) 0.7 0.30 (0.12) (0.05) 0.9 0.40 (0.16) (0.07)
0.5 0.25 (0.09) (0.04) 0.9 0.40 (0.16) (0.07) 0.7 0.45 (0.12) (0.08)
S-ln. (203 J mm) Concrete Slab— Wo. (15.4 mm) Cover
0.5-in. (12.7 mm) (nom.) nonferrous tube
9 3.6 0.30 1.0 0.35 0.8 0.40 (228.6) (0.63) (0.05) (0.18) (0.06) (0.14) (0.07)
12 5 J 0.35 1.2 0.45 1.0 0.55 (304.8) (0.92) (0.06) (0.21) (0.08) (0.18) (0.10)
0.5-in. (12.7 mm) (nom.) ferrous pipe or 0.75-in. (19.1 mm) (nom.) non-ferrous tube
9 2.9 0.25 0.9 0J0 0.8 0.35 (228.6) (0JI) (0.04) (0.16) (0.05) (0.14) (0.06)
12 4.0 0.30 I.I 0.40 0.9 0.45 (304.8) (0.70) (0.05) (0.19) (0.07) (0.16) (0.08)
0.75-in. (19.1 mm) (nom.) ferrous pipe or I-in. (25.4 mm) (nom.) non-ferrous tube
9 2.2 0.20 0.8 0.30 0.7 0.30 (228.6) (0.39) (0.04) (0.14) (0.05) (0.12) (0.05)
12 3.3 0.30 1.0 0J5 0.8 0.40 (304.8) (0.58) (0.05) (0.18) (0.06) (0.14) (0.07)
15 4.3 0JS 1.1 0.40 0.9 0J0 (381) (0.76) (0.06) (0.19) (0.07) (0.16) (0.09)
I-in. (25.4 mm) (nom.) ferrous pipe 9 1.7 0.20 0.7 0.25 0.7 0.25 (228.6) (0.30) (0.04) (0.12) (0.04) (0.12) (0.04)
12 2.7 0.25 0.9 0.30 0.8 0.35 (304.8) (0.48) (0.04) (0.16) (0.05) (0.14) (0.06)
15 3.7 0.30 1.0 0.40 0.9 0.45 (381) (0.63) (0.05) (0.18) (0.07) (0.16) (0.08)
aAny ceiling panel also acts as a floor panel to the extent of its upward heat flow. If the upward heat flow is high and the space above is occupied, check floor surface temperature for possible foot discomfort (see Ref 8). Also check effect on heating requirements of the space above, it is not good practice to have the major portion of the upper room's heating requirements supplied by the upward heat flow of a ceiling panel below.
heat losses are part of the building heat loss if the heat is transferred outside of the building. If the heat is transferred to another heated space, the panel loss is a source of heat for the space and is not a part of the building heat loss. In either case, the magnitude of panel loss should be determined.
Panel heat loss to space outside the room should be kept to a reasonable amount by insulation. Panel heat loss to heated spaces may require reduction by insulation if the amount of heat transferred is excessive or if objectionable temperatures develop. For example, a floor panel may overheat the basement below and a ceiling panel may cause the temperature of a floor surface above it to be too high for comfort unless it is properly insulated.
The heat loss from most panels can be calculated by using the coefficients given in Chapter 20 of the 1981 FUNDAMENTALS VOLUME. These coefficients should not be used to determine the downward heat loss from panels-built on grade be-
D-9
Panel Heating and Cooling Systems
Table 3 Thermal Resistance of Piaster Celling Panels (Heating or Cooling)
Thermal Resistance to Downward Heat Flow, rd, ft2.F>h/Bra (m2 ."C/W)
Plaster Pancbb Standard
Gypsum Plaster. Three Coats
Spacing, In.
(mm)
Heat Flow Ratio,* qu/qj
0.2 0.4 0.6 0.8 1.0»
SPACING TUBES OR PIPES
J — * - H
4.5 (114)
0.30 (0.05)
0.34 (0.06)
0.38 (0.07)
0.42 (0.07)
0.46 (0.08)
0.50 (0.09)
METAL LATH
METAL OR /GYPSUM LATH
•miMmm>m^<&i ^TUBES
3/8-in. (9.5 mm) (nom.) nonferrous tube or '/S-in. (12.7 mm) (nom.) 6 ferrous pipe above metal lath tied at (192) 8-in. (203.2 mm) intervals 9 with good tube embedment, (229) or 3/8-in. (9.5 mm) (nom.) nonferrousl6 tube below m clal or gypsum lath (406)
•Any ceiling panel also acts as a floor panel to the extent of its upward heat now. If the upward heat flow Is high and the space above is occupied, check floor surface tempciature for possible foot discomfort (see Ref 10). Also check effect on heating requirements of the space above. It is not good practice to have the major portion of the upper room's heating requirements supplied by the upward heal flow of a ceiling panel below.
bRecommendcd maximum inlet water temperature ('max) - 140F(60°Q.
8.9 rOUNDATK
GRADE
¥ i
0.45 (0.08) 0.75 (0.13) 1.15
(0.20)
0.51 (0.09) 0.85
(0.15) 1.29
(0.23)
0.57 (0.10) 0.95
(0.17) 1.43
(0.25)
0.63 (0.11) 1.05
(0.18) 1.57
(0.28)
0.69 (0.12) 1.15
(0.20) 1.71
(0.30)
0.75 (0-13) 1.25
(0.22) 1.85
(0.33)
< o
4
| i a
ySLAB
•:-. EITHER WAV
INSULATION AT SLAfc/
EDGE ONLY d - 0
1 1 (NO INSULATION
^ C o O j .
com
W/m1
-r— T C3=I.B)
\
1 fERStON FACTOAfl XO3048 - C - B t W I i r i * fXSJBB
(C,»
1 1
1.3
i
i 1.2
I.I
1.0
0.9
0.8
0.7
O.I 0.2 0.3 0.4 INSULATION CONDUCTANCE BTU PER (HR) ISO FT) (F DEC)
Fig. 11 Downward and Edgewise Heat Loss Coefficient for
Concrete Floor Slabs on Grade
cause the heat flow from them is not uniform.""13 The heat loss from panels built on grade can be estimated from Fig. 11.
PANEL HEATING A N D COOLING SYSTEMS
The most common forms of panels applied in panel heating and cooling systems are:
1. Metal ceiling panels. 2. Embedded piping in ceilings, walls or floors. 3. Air-heated floors. 4. Electrically heated ceilings or floors. 5. Electric ceiling panels.
Table 4 Thermal Resistance of Metal Ceiling Panels Thermal Resistance to Heat Flow
fti'F'h/Btu (m»»0C/W)
Type of Panel Spacing—Inches (mm)
STEEL PIPE PAN E06E HELD AGAINST PIPE BY SPRING CUP
3 (76.2)
6
(132)
12 (305)
0.31" (0.05)
0.61" (0.11)
-ALUMINUM PAN 0 .032 IN. THICK
\
TUBE SPACING
COPPER TUBE SOLOEREO TO 0 . 0 4 0 IN. THICK ALUMINUM SHEET
•r 4
(102)
8 (203)
0.071" (0.01)
0.15" (0.03)
Metal Ceiling Panels Metal ceiling panels are usually integrated into a system
that heats and cools. In such a system, a source of dehumidified ventilation air is required in summer, and the system is classed as one of the combination air-water systems. Also, various amounts of forced air are supplied year-round. When metal panels are applied for heating purposes only, a ventilation system may or may not be required, depending on local codes.
Figure 12 illustrates a metal ceiling panel system that uses 0.5-in. (13-mm) pipe laterals, on either 6-, 12- or 24-in. (150-, 300- or 600-mm) centers, hydraulically connected in a sinuous or parallel flow welded system. Aluminum ceiling panels are clipped to these pipe laterals, acting as a heating panel when warm water is flowing, or as a cooling panel when chilled water is flowing. Figure 14 illustrates a metal panel ceiling system using copper tubing, metallurgicaliy bonded to an aluminum panel that can be mounted into various types of ceiling
Table 5 Thermal Resistance of Floor Coverings
•Manufacturer's data.
Description
Bare concrete, no covering Asphalt tile Rubber tile
light carpet Light carpet with rubber pad Light carpet with light pad Light carpet with heavy pad
Heavy carpet Heavy carpet with rubber pad Heavy carpet with light pad Heavy carpet with heavy pad
Resistance ft*«F«h/Bto
0.00 O.OS 0.05
0.6 1.0 1.4 1.7
0.8 1.2 1.6 1.9
fruef <m2»°C/W)
!!!
(0.11) (0.18) (0.25) (0.30)
III!
D-10
8.10 CHAPTER 8 1984 Systems Handbook
, THERMAL «L*N*CT
l £ SQUARE HEADER
SUPFOIIT BRACKET
AND CUPS
COPPER TUBE PRESSED INTO
OVAL CHANNEL .
COftVf NSKMI ACTOR
PANEL CLIP
V , PIPE LATERAL
Fig. 12 Metal Ceiling Panels Attached lo Pipe Laterals
suspension systems. Figure IS illustrates a metal panel using a copper tube pressed into an aluminum extrusion.
Metal ceiling panels can be perforated so that the ceiling becomes sound absorbent when acoustical material is installed, either on the back side of the panels or on the underside of the overhead floor system. The acoustical blanket is also required for thermal reasons, so that the reverse loss or upward flow of heat from the metal ceiling panels is minimized.
The metal panel system shown in Fig. 12 and 13 is designed so that the grid system can expand relative to the suspension system, but the suspension system should not move in relation to the building. The suspension system holding the ceiling grids in place must be braced by cross-furring against the walls of the building in both horizontal directions. The expansion and contraction of the ceiling grids are compensated for by allowing the metal panels to move or adjust in the wall molding. The maximum design value of movement is 0.56 in. (14 mm). Many large grids have been constructed in the field by butt welding the pipe laterals together to produce 100 ft. (30 m) lengths. A steel pipe grid will expand 0.8 in. per 100 ft per 100 deg F (0.67 mm/m per 56°C). Such large grids should have metal panel ceiling expansion joints that line up with the expansion joints of the building.
Metal ceiling panels can be used with two- and four-pipe distribution systems. It is common to design for a 20 deg F (11°C) temperature drop for heating across a given grid and a 5 deg F (2.8°C) rise for cooling, but higher temperatures drops can be used if applicable.
COPPER TUBE, ALUMINUM SHEET, MODULAR RADIANT AC0UST4CAL PANEL
Fig. 13 Metal Ceiling Panels Metallurgical^ Bonded to Copper Tubing
Fig. 14 Extruded Aluminum Panel with Integral Copper Tube
Some ceiling installations require active grids to cover only a portion of the room, while compatible matching acoustical panels are selected for the remaining ceiling area. Extruded aluminum-type panels are often used as long-narrow panels at the outside wall and are independent of the ceiling system.
Embedded Piping in Ceilings, Walls and Floors When piping is embedded in ceilings, one of the following
constructions is generally used: 1. Pipe or tube is embedded in the lower portion of a concrete slab,
generally within an inch (25 mm) of its lower surface. If plaster is to be applied to the concrete, the piping may be placed directly on the wood forms. If the slab is to be used without plaster finish, the piping should be installed not less than 0.75 in. (19 mm) above the under-surface of the slab. Figure IS shows this method of construction. The minimum coverage must comply with local building code requirements.
2. Pipe or tube is embedded in a metal lath and plaster ceiling. If the lath is suspended to form a hung ceiling, the lath and heating coils are securely wired to the supporting members so that the lath is below, but in good contact with, the coils. Plaster is then applied to the metal lath, carefully embedding the coil as shown in Fig. 16.
3. Smaller diameter copper tube is attached to the underside of wire lath or gypsum lath. Plaster is then applied to the lath to embed the tube, as shown in Fig. 17.
4. Other forms of ceiling construction are composition board,
'»:o:o\?:i\%-.?io;„.;o::o-:-.oyi-.:o;;..i.v-.o:
' -"°.• o : 'CONCRETE SLAB,".?• £ \ "•.-.0-'. V." • ° : • ?". '"
o . o ; • -o'. o.v
.HEATING COILS" •.<>-
'o ' . 'o ' : •[: o.-.o.'.
Fig. 15 Coils in Structural Concrete Slab
WIRE TIE TO SUPPORTS
SCRATCH COAT HEATING / EMBEDDING PIPING / PIPES
SUPPORTING MEMBERS ON 3 OR 4 FT CENTERS. USE PIPE OR STEEL STRUCTURAL MEMBERS
3 COAT PLASTER J NMETAL LATH
FINISH FINISHED PLASTER CEILING
SUSPENDED PLASTER C E I L I N G
SCRATCH COAT EMBEDDING PIPES
3 COAT PLASTER I ^METAL LATH
FINISH FINISHEO PLASTER CEILING
PLASTER CEIL ING BELOW J O I S T S
Fig. 16 Coils in Piaster Above Lath
D - l l
Panel Heating and Cooling Systems 8.11
LATH APPLIED' BELOW J O I S T S
C O I L S W I R E -T I E D TO L A T H
3 COAT P L A S T E R F I N I S H
H E A T I N G C O I L S BELOW L A T H
F IN ISHED PLASTER CEILING
Fig. 17 Coils in Plaster Below Lath
ms-m •FOUNDATION
Fig. 18 Coils in Floor Slab on Grade
wood paneling and so forth, with warm water piping, tube or channels built into the panel sections.
Coils are usually the sinuous type, although some header or grid-type coils have been used in ceilings. Coils may be plastic, ferrous or nonferrous pipe or tube, with coil pipes spaced from 4.5 to 9 in. (1 IS to 230 mm) on centers, depending on the required output, pipe or tube size and other factors.
Where plastering is applied to pipe coils, a standard three-coat gypsum plastering specification14 is followed, with a minimum of 0.38 in. (9.6 mm) of cover below the tubes when they are installed below the lath. Generally, the surface temperature of plaster panels should not exceed 120 F (49°C). This can be accomplished by limiting the water temperature in the pipes or tubes in contact with the plaster to a maximum temperature of 140 F (60°C). Insulation should be placed above the coils to reduce reverse loss, the difference between heat supplied to the coil and net useful output to the heated room.
To protect the plaster installation and to assure proper air drying, heat must not be applied to the panels for two weeks after all plastering work has been completed. When the system is started for the first time, the water supplied to the panels should not be Higher than 20 deg F (11°C) above the prevailing room temperature at that time and not in excess of 90 F (32°C). Water should be circulated at this temperature for about two days, then increased at a rate of about 5 deg F (2.8°Q per day to 140 F (60°C).
During the air-drying and preliminary warm-up periods, there should be adequate ventilation to carry moisture from the panels. No paint or paper should be applied to the panels before these periods have been completed or while the panels are being operated. After paint and paper have been applied, an additional shorter warm-up period, similar to first-time starting, is also recommended.
Although not as universally used as ceiling panels, wall panels can be constructed by any of the methods outlined for ceilings. The construction for piping embedded in floors depends on whether the floor is laid on grade or above grade.
1. Plastic, ferrous and nonferrous pipe and tube are used in floor slabs that rest on grade. The coils are constructed as sinuous-continuous pipe coils or arranged as header coils with the pipes spaced from 6 to 18 in. (ISO to 450 mm) on centers. The coils are generally installed
OUTSIDE WALL
CONCRETE EDGE CURBING
ASPHALT IMPREGNATED INSULATION BOARD
POURED SLAB
PRECAST SLAB CONCRETE FLOOR SUPPORTS
/ AND DIRECTIONAL VANES
- A I R PLENUM
-CONCRETE
- INSULATION
-GRAVEL OR ROCK FILL
- E A R T H
FOOTING WALL
Fig. 19 Warm Air Floor Panel Construction
with 1.5 to 4 in. (40 to 100 mm) of cover above the coils. Insulation is recommended to reduce the perimeter and reverse losses. Figure 18 shows the application of pipe coils in slabs resting on grade. Coils should be embedded completely and should not rest on an interface. Any supports used for positioning the heating coils should be nonab-sorbent and inorganic. It is suggested that reinforcing steel, angle iron, pieces of pipe or stone or concrete mounds be used. No wood, brick, concrete block or similar materials should be used for support of coils. A waterproofing layer is desirable to protect insulation and piping.
2. Where the coils are embedded in structural load-supporting slabs above grade, construction codes may affect their position. Otherwise, the coil piping is installed as described for slabs resting on grade.
3. A warm-up and start-up period for concrete panels should be similar to that outlined for plaster panels.
Air-Heated Floors Several methods have been devised to warm interior room
surfaces by circulating heated air through passages in the floor. In some cases the heated air is recirculated in a closed system. In others, all or a part of the air is passed through the room on its way back to the furnace to provide supplementary heating and ventilation. Figure 19 indicates one common type of construction. Compliance with applicable building codes is important.
Electrically Heated Ceilings Several different forms of electric resistance units are avail
able for heating interior room surfaces. These include: (1) electric heating cables that may be embedded in concrete or plaster or laminated in drywall ceiling construction, (2) prefabricated electric heating panels to be attached to room surfaces and (3) electrically heated fabrics or other materials for application to, or incorporation into, finished room surfaces.
Electric heating cables for embedded or laminated, ceiling panels are factory-assembled units furnished in standard lengths of about 75 to 1800 ft (25 to 550 m). These cable lengths cannot be altered in the field. The cable assemblies are normally rated at 2.75 W per linear ft (9 W/m) and are supplied in capacities from 200 to 5000 W in roughly 200-W increments. Standard cable assemblies are available for 120,208 and 240 V. Each cable unit is supplied with 7-ft (2.1-m) non-heating leads for connection at the thermostat or junction box.
Electric cables for panel heating have electrically insulated coverings resistant to medium temperature, water absorption, aging effects and chemical action with plaster, cement or ceiling lath material. This insulation is normally a polyvinylchlo-ride (PVC) covering which may have a nylon jacket. The outside diameter of the insulation covering is usually about 0.12 in. (3 mm).
For plastered ceiling panels, the heating cable may be sta-
D-12
8.12 CHAPTER 8 1984 Systems Handbook pled to gypsum board, plaster lath or similar fire-resistant materials with rust-resistant staples. With metal lath or other conducting surfaces, a coat of plaster (brown or scratch coat) is applied to completely cover the metal lath or conducting surface before the cable is attached. After fastening on the lath and applying the first plaster coat, each cable is tested for continuity of circuit and for insulation resistance of at least 100,000 ohms measured to ground.
The entire ceiling surface is finished with a covering of thermally noninsulating sand plaster about 0.S0 to 0.7S in. (13 to 19 mm) thick or other approved noninsulating material applied according to manufacturer's specifications. The plaster is applied parallel to the heating cable, rather than across the runs. While new plaster is drying, the system should not be energized and the range and rate of temperature change should be kept low by other heat sources or by ventilation until the plaster is thoroughly cured. Vermiculite or other insulating plaster causes cables to overheat and is contrary to code provisions.
For laminated drywall ceiling panels, the heating cable is placed between two layers of gypsum board, plasterboard or other thermally noninsulating fire-resistant ceiling lath. The cable is stapled directly to the first (or upper) lath, and the two layers are held apart by the thickness of the heating cable. It is essential that the space between the two layers of lath be completely filled with a noninsulating plaster or similar material. The purpose of this fill is to hold the cable firmly in place and to improve heat transfer between the cable and the finished ceiling. Failure to fill the space completely between the two layers of plasterboard may allow the cable to overheat in the resulting voids and may cause cable failure. The plaster fill should be applied according to manufacturer's specifications.
Electric heating cables are ordinarily installed with a 6-in. (150-mm) nonhealing border around the periphery of the ceiling. An 8-in. (200-mm) clearance must be provided between heating cables and the edges of the outlet or junction boxes used for surface-mounted lighting fixtures. A 2-in. (SI mm) clearance must be provided from recessed lighting fixtures, trim and ventilating or other openings in the ceiling.
Heating cables or panels must be installed only in ceiling areas which are not covered by partitions, cabinets or other obstructions. However, it is permissible for a single run of isolated embedded cable to pass over a partition.
The National Electric Code requires that all general power and light wiring be run above the thermal insulation or at least 2 in. (51 mm) above the heated ceiling surface, or that the wiring be derated.
In drywall ceiling construction, the heating cable is always installed with the cable runs parallel to the joist. A 2.5-in. (64-mm) clearance between adjacent cable runs must be left centered under each joist for nailing. Cable runs that cross over the joist must be kept to a minimum. Where possible, these crossings should be in a straight line at one end of the room.
Figure 20 shows details of ceiling cable installation practice for plastered construction.
The spacing between adjacent runs of heating cable can be determined using Eq. (10):
s=UA„/C (10) where
s = cable spacing, in. (mm). A„ = net panel heated area, ft2 (m2).
C = length of cable, ft (m). For cable having a watt density of 2.75 W/ft (9 W/m), the
LEAVE 8 IN. (203.2 MM> CLEARANCE BETWEEN BOX AND HEATING WIRE 6IN.(152.4 MM) INSULATION MINIMUM
STAPLE 6 IN. (152.4 MM) FROM TURN 3 IN. (76.2 MM) FROM TURN AND ON RADIUS OF BEND (MAX. STAPLE SPACING 16 IN
IN. (152 4 MMI CLEAR SPACE FROM WALL
FULL THICK -INSULATION
1 . :rJ
i -• -irv
EXTERIOR FINISH
Fig. 20 Electric Heating Panel for Wet Plastered Celling
D-13
Panel Heating and Cooling Systems
minimum permissible spacing is 1.5 in. (38 mm) between adjacent runs. Some manufacturers recommend a minimum spacing of 2 in. (51 mm) for drywall construction.
Net panel area, A„, in Eq. (10) is the net ceiling area available after deducting the area covered by the nonheating border, lighting fixtures, cabinets and other ceiling obstructions. Since, for simplicity, Eq. (10) contains a slight safety factor, small lighting fixtures are usually ignored in determining net ceiling area.
The 2.5-in. (64-mm) clearance required under each joist for nailing in drywall applications occupies one-fourth of the ceiling area, if the joists are 16 in. (400 mm) o.c. Therefore, for drywall construction, the net area, A„, must be multiplied by 0.75. Many installations have a spaaing of 1.5 in. (38 mm) for the first 2 ft (600 mm) from the cold wall. Remaining cable is then spread over the balance of the ceiling.
Prefabricated Electric Ceiling Panels A variety of prefabricated electric heating panels are avail
able for either supplemental or full room heating. These panels are available in sizes from 2 x 4 ft (0.6 x i .2 m) to 6 x 12 ft (1.8 x 3.6 m). They are constructed from a variety of materials such as gypsum board, glass, steel and vinyl. Different panels have rated inputs varying from 10 to 95 W/ft2 (108 to 1023 W/m2) for 120, 208, 240 and 277 V service. Maximum operating temperatures vary from about 100 to about 300 F (38 to 149 °C), depending on watt density. Consult the national and local codes for restrictions on the location of partitions, lights and air grilles adjacent to or near electric panels.
Panel heating elements may be embedded conductors, laminated conductive coatings or printed circuits. Nonheating leads are connected and furnished as part of the panel. Some panels can be cut to fit available space; others must be installed as received. Panels may be either flush or surface mounted and, in some cases, are finished as part of the ceiling. Rigid panels that are about 1-in. (25-mm) thick and weigh about 25 lb (11kg) each are available to fit standard 2 x 4 ft (0.6 x 1.2 m) modular tee-bar ceilings. Always follow the installation instructions furnished by the manufacturer.
Electrically Heated Wall Panels Cable embedded in walls similar to ceiling construction is
occasionally found in Europe. Because of possible damage from nails driven for hanging pictures or from building alteration, most codes in the United States prohibit such panels. Some of the prefabricated panels described in the preceding section are also used for wall panel heating.
Electrically Heated Floors Electric heating cable assemblies, such as those used for
ceiling panels, are sometimes used for concrete floor heating systems. Since the possibility of cable damage during installation is greater for concrete floor slabs than for ceiling panels, these assemblies must be carefully installed. After the cable has been placed, all unnecessary traffic should be eliminated until the concrete covering has been poured and hardened.
Preformed mats are sometimes used for electric floor slab heating systems. These mats usually consist of PVC-insulated heating cable woven in, or attached to, metallic or glass fiber mesh. Such mats are available as prefabricated assemblies in many sizes from 2 to 100 ft2 (6.18 to 9.3 m2) and with various watt densities.
Mineral-insulated (MI) heating cable is another effective method of slab heating. MI cable is a small-diameter, highly durable, flexible heating cable composed of solid electric-resistance heating wire or wires surrounded by tightly compressed magnesium oxide electrical insulation and enclosed by a metal sheath. MI cable is available in stock assemblies in a
8.13
Fig. 21 Electric Heating Cable in Concrete Slab
variety of standard voltages, watt densities and lengths. A cable assembly consists of the specified length of heating cable, waterproof hot-cold junctions, 7-ft (2.1-m) cold sections, UL-approved end fittings and connection leads. Several standard Ml cable constructions are available, such as single conductor, twin,conductor and double cable. Custom-designed MI heating cable assemblies can be ordered for specific installations.
Other outer-covering materials that are sometimes specified for electric floor heating cable include: (1) silicone rubber, (2) lead and (3) tetrafiuoroethylene (Teflon).
Floor Heating Cable Installation When PVC-jacketed electric heating cable is used for floor
heating, the concrete slab is laid in two pourings. The first pour should be at least 3-in. (75-mm) thick and, where practical, should be insulating concrete to reduce downward heat loss. For a proper bond between the layers, the finish slab should be poured within 24 hours of the first pour, with a bonding grout applied. The finish layer should be at least 1.5 in. (38 mm) and not more than 2-in. (Sl-mm) thick. This top layer must not be insulating concrete (see Fig. 21). At least 1 in. (25 mm) of perimeter insulation should be installed as shown in Fig. 12 and 21.
The cable is installed on top of the first pour of concrete not closer than 2 in. (51 mm) from adjoining walls and partitions.
Methods of fastening the cable to the concrete include:
1. The cable is stapled to wood nailing strips fixed in the surface of the rough slab. The predetermined cable spacing is maintained by daubs of cement, plaster of paris or tape.
2. In lightweight or uncured concrete, the cable can usually be stapled directly to the slab using hand-operated or powered stapling machines.
3. Special anchor devices are available that are nailed to the first slab to hold the cable in position while the top layer is being poured.
Preformed mats can be embedded in the concrete in a continuous pour. The mats are positioned in the area between expansion and/or construction joints and electrically connected to a junction box. The slab is poured to within 1.5 to 2 in. (38 to 51 mm) of the finished level. The surface is rough screeded, and the mats placed in position. The final cap is applied immediately. Since the first pour has not set, there is no adhesion problem between the first and second pour, and a monolithic slab results. A variety of contours can be developed by using heater wire attached to glass fiber mats. Allow for circumvention of obstructions in the slab.
MI electric heating cable can be installed* in concrete slab using either one or two pours. For single-pour applications
D-14
8.14 CHAPTER 8 1984 Systems Handbook the cable is fastened to the top of the reinforcing steel before the pour is started. For two-layer applications the cable is laid on top of the bottom structural slab and embedded in the finish layer. Proper spacing between adjacent cable runs is maintained by using prepunched copper spacer strips nailed to the lower slab.
For a given floor heating cable assembly, the required cable spacing is determined from Eq. (10). In general, cable watt density and spacing should be such that floor panel watt density is not greater than 15 W/ft2 (161 W/m2). Higher watt densities [up to 25 W/ft2 (269 W/m2)) are often specified for the 2-ft (0.6-m) border next to cold walls. It is important to check with the latest issue of the National Electric Code and other applicable codes to obtain information on maximum panel watt density and other required criteria and parameters.
GENERAL DESIGN CONSIDERATIONS
The application, design and installation of panel systems have certain requirements and techniques that should be recognized:
1. As wilh any hydronic system, close attention should be paid to the piping system design. Piping should be designed to assure that water of the proper temperature and in sufficient quantity will be available to every grid or coil at all times. Reverse-return systems should be considered to minimize balancing problems.
2. The apparatus dew point of the cooling coils in the air distributing system should be designed for full capacity plus a 10 to 15% safety factor, because most problems occur when the supply air is short on dehumidification capacity.
SHR = RSH - PC
RTH
where SHR = sensible heat ratio RSH = room sensible heat
PC = panel cooling RTH = room total heat
3. Individual ceiling panels can be connected for parallel flow using headers, or for sinuous or serpentine flow. To avoid flow irregularities within a header-type grid, the water channel or lateral length should be greater than the header length. If the laterals in a header grid are forced to run in a short direction, this problem can be solved by using a combination series-parallel arrangement.
4. Noises from entrained air, high velocity or high pressure drop devices or from pump and pipe vibrations must be avoided. Water velocities should be high enough [usually 1.S fps (0.46 m/s) or higher] to prevent separated air from accumulating and causing air binding. • Where possible, avoid automatic air venting devices over ceilings of occupied spaces.
5. Design piping systems to accept thermal expansion adequately. Do not allow forces from piping expansion to be transmitted to ceiling panels. Thermal expansion of the ceiling panels must be considered.
6. In circulating water systems, both steel and copper pipe or tube are used widely in ceiling, wall or floor panel construction. Some types of plastic pipe also may be suitable where codes permit. Where coils are embedded in concrete or plaster, no threaded joints should be used for either pipe coils or mains. Steel pipe should be the all-welded type. Copper tubing should be soft-drawn coils. Fittings and connections should be minimized. Changes in direction should be made by bending. Solder-joint fittings for copper tube should be used with a medium temperature solder of 95% tin, 5% antimony or capillary brazing alloys. All piping should be subjected to a hydrostatic test of at least three limes the working pressure, but not less than ISO psig (1033 kPa). Maintain adequate pressure in piping while pouring concrete.
7. Locate ceiling panels adjacent to the outside wall and as close as possible to the areas of maximum load. The panel area within 3 ft (1.0 m) of the outside wall should have a heating capacity equal to or greater than 50% of the wall transmission load.
8. Ceiling system designs based on passing return air through the
4 0
3 0
2 0
to
1 | !
1 j \y
' y / l
1
CO
-
i /\
/ j i
; 1 i i
NVERSIOR1 f ACTORS - IF - 3 2 l ' l 8 —
h X 0.3MB
I i I 1 ! 1 IOO ISO 700 750 300
PANEL SURFACE TEMP, F
Fig. 22 Suggested Design Ceiling Surface Temperatures at Various Ceiling Heights
panels into the plenum space above the ceiling are not recommended, because much of the panel heat transfer is lost to the return air system.
9. Allow sufficient space above the ceiling for installation and connection of the piping that forms the radiant panel ceiling.
10. Placing the thermostat on a side wall where it can see the outside wall and the warm ceiling should be considered. The normal thermostat cover reacts to the warm ceiling panel, and the radiant effect of the ceiling on the cover tends to alter the control point so that the thermostat controls 2 to 3 deg F (1 to 2°C) lower when the outdoor temperature is a minimum and the ceiling temperature is a maximum. Experience indicates that radiantly heated rooms are more comfortable under these conditions than when the thermostat is located on a back wall.
11. When selecting a ceiling panel surface temperature, mean water temperature or watt density of an electric panel, the design parameters are:
a. Excessively high temperatures over the occupied zone will cause ' the occupant to experience a "hot head effect."
b. Temperatures that are too low can result in an oversized, uneconomical panel and a feeling of coolness at the outside wall.
c. The technique in item 7 above should be given priority. d. With normal ceiling heights of 8 to 9 ft (2.4 to 2.7 m), panels less
than 2 ft (0.6 m) wide at the outside wall can be designed for 235 F (113°Q surface temperature. If panels extend beyond 2 or 3 ft (0.6 or 0.9 m) into the room, the panel surface temperature should be approximately as given in Fig. 22. The surface temperature of concrete or plaster panels is limited by construction.
12. If throttling valve control is used, either the end of the main should have a fixed bypass, or the last one or two rooms on the mains should have a bypass valve to maintain water flow in the main. Thus, when a throttling valve modulates, there will be a rapid response.
13. When the panel chilled water system is started, the circulating water temperature should be maintained at room temperature until the air system is completely balanced, the dehumidification equipment is operating properly and building humidity is at design value.
14. When the panel area for cooling is greater than the area required for heating, a two-panel arrangement (Fig. 23) can be used.
% SUPPLY
RETUI I N -
> *
HC
CO
Vbvi
T
i
i T
i
0V2
—i
- * -
(el TWO-PIPE SYSTEM IM FOUR-PIPE SYSTEM
Fig. 23 Split Panel Piping Arrangement for Two-Pipe and Four-Pipe Systems
D - 1 5
Panel Heating and Cooling Systems 8.15 Panel HC (heating and cooling) is supplied with hot or chilled water year-round. When chilled water is used, the controls function to activate panel CO (cooling only), and both panels are used for cooling.
15. To prevent condensation on the room side of cooling panels, the panel water supply temperature should be maintained at least 1 deg F (0.6°C) above the room design dewpoint temperature. This minimum difference is recommended to allow for the normal drift of temperature controls for the water and air systems, and also to provide a factor of safety for temporary increase in space humidity.
16. Selection of summer design room dewpoint below 50 F (10°Q generally is not economical.
17. The most frequently applied method of dehumidification utilizes cooling coils. If the main cooling coil is six rows or more, the dewpoint of the air leaving will approach the temperature of the water leaving. The cooling water leaving the dehumidifier can then be used for the panel water circuit without danger of condensation during normal operation.
18. Several chemical dehumidification methods are available to control latent and sensible loads separately. In one application, cooling tower water is used to remove heat from the chemical drying process, and additional sensible cooling is necessary to cool the dehumidified air to the required system supply air temperature.
19. When chemical dehumidification is used, hygroscopic chemical-type dewpoint controllers are required at the central apparatus and at various zones to monitor dehumidification.
20. When cooled ceiling panels are used with a variable air volume (VAV) system, the air supply rate should be near maximum volume to assure adequate dehumidification before the cooling ceiling panels are activated.
21. Controls (Heating). Automatic controls for panel heating differ from those for convective heating because of the thermal inertia characteristics of the panel heating surface and the increase in the mean radiant temperature within the space under increasing loads for panel heating. However, many of the control principles for hot water heating systems described in Chapters IS and 16 also apply to panel heating. See Chapter 31 for further information on automatic controls.
Panels such as concrete slabs have large heat storage capacity and continue to emit heat long after the room thermostat has shut off the heating medium supply. In addition, there is a considerable time lag between thermostat demand and heat delivery to the space, since a large part of the heat must first be stored in the thermally heavy radiant surface. This inertia will cause uncomfortable variations in space conditions unless controls are provided to detect load changes early.
In general, the temperature of the heating medium supplied to the panel surface should be varied in accordance with outdoor temperature. However, for embedded pipe panels precautions must be taken to prevent the introduction of excessively hot water, which might damage the panels if controls failed. A manual boiler bypass or other means of reducing the water temperature may be necessary to prevent new panels from drying out too rapidly (see Embedded Piping for Ceiling Panels).
Because the mean radiant temperature (MRT) within a panel heated space must increase as the heating load increases, the air temperature during this increase should be lowered I or 2 deg F (0.5 to l°C)to maintain comfort. In ordinary structures with normal infiltration
• loads, the required reduction in air temperature is small, enabling a conventional room thermostat to be used.
In panel heating systems, lowered night temperatures will produce unsatisfactory results with heavy panels such as concrete floors. These panels cannot respond to a quick increase or decrease in heating demand within the relatively short time required, resulting in a very slow reduction of the space temperature at night and a correspondingly slow pickup in the morning. Lightweight panels, such as plaster or metal ceilings and walls, may respond to changes in demand quickly enough for moderately satisfactory results from lowered night temperatures. Tests on a metal ceiling panel demonstrated the speed of response to be comparable to that of conventional environmental systems.15 However, very little fuel savings can be expected even with light panels unless the lowered temperature is maintained for long periods. If reduced nonoccupancy temperatures are employed, some means of providing a higher-than-normal rate of heat input for rapid warm-up is necessary, or a long warm-up period should be provided, as explained in Chapter 16.
Electric Heating Slab Controls. For comfort heating applications.
the surface of a floor slab {!,) is held to a maximum of 80 to 85 F (26 to 29°Q. Therefore, when used as a primary heating system, thermostatic control devices sensing air temperature should not be used to control the slab temperature, but should be wired in series with a slab-sensing thermostat. The remote sensing thermostat in the slab acts as a limit switch to control maximum surface temperatures allowed on the slab. The ambient sensing thermostat controls the comfort level. For supplementary slab heating, as in kindergarten floors, a remote sensing thermostat in the slab is commonly used to tune in the desired comfort level. Indoor-outdoor thermostats are used to vary the floor temperature inversely with the outdoor temperature. If the heat loss of the building is calculated for 70 to 0 F (21 to -18°C), and the floor temperature range is held from 70 to 85 F (21 to 29°C) with a remote sensing thermostat, the ratio of outdoor temperature to slab temperature is 70:15 (39:8), or approximately 5:1. This means that a 5 deg F (2.8°C) drop in outdoor temperature requires a 1 deg F (0.56°C) increase in the slab temperature. An ambient sensing thermostat is used to vary the ratio between outdoor and slab temperatures. A time clock is used to control each heating zone if off-peak slab heating is desirable.
22. Controls (Cooling). Controlling the panel water circuit temperature by mixing, heat exchange or using the water leaving the dehumidifier is the major consideration in preventing condensation. Other considerations are listed in items 12 and 14. It is imperative to dry out the building space before starting the panel water system, particularly after extended down periods, such as weekends. Such delayed starting action can be controlled manually or by device.
Panel cooling systems require the following basic areas of temperature control: (1) exterior zones, (2) areas under exposed roofs to compensate for transmission and solar loads and (3) control of each typical interior zone to compensate for internal loads. For optimum results, each exterior corner zone and similarly-loaded face zone should be treated as a separate subzone.
Panels are suitable for control systems which are scheduled by elements that sense solar and weather changes before these changes affect the space temperature. Window pane thermocouples have been used to schedule water temperatures in panels under a window sill. Photoelectric cells can be used to divert cold water into a peripheral ceiling panel; e.g., picking up the winter solar load on a south zone. Panel cooling systems have also been zoned to provide individual temperature control in exterior offices, particularly in applications where there is a high lighting load, or for corner rooms with large glass areas on both walls.
The temperature control of the interior air and panel water supply should not be functions of the outdoor weather. The normal thermostat drift is usually adequate compensation for the slightly lower temperatures desirable during winter weather. This drift should be limited to result in a room temperature change of not more than 1.5 deg F (0.8°C). Control of the interior zones is best accomplished by devices that reflect the actual presence of the internal load elements. Frequently, time clocks and current sensing devices are used on lighting feeders.
Because air quantities are generally small, it is not advisable to use volume control in any part of the system. With the apparatus arranged to supply air of appropriate apparatus dewpoint at all times, it is possible to avoid compromising indoor conditions with a panel cooling system throughout the year. As with all systems, to prevent condensation on window surfaces, the supply air dew point should be reduced during extremely cold weather according to the type of glazing installed.
For eeneral information on automatic controls, refer to Chapter 31.
PANEL HEATING SYSTEM DESIGN
Design Steps Panel design requires specifying the following: (1) panel
area, (2) size and location of the heating elements in the panel, (3) insulation on the reverse side and edge of the panel, (4) required input to panel and (5) temperature of the heating elements. The procedure is summarized as follows:
1. Calculate the hourly rate of heat loss for each room. 2. Determine the available area for panels in each room. 3. Calculate the required unit panel output. 4. Determine the required panel surface temperature.
D - 1 6
8.16 CHAPTER 8 1984 Systems Handbook 5. Select the means of heating the panel and the size and location of
the healing elements. 6. Select insulation for the reverse side and edge of panel. 7. Determine panel heat loss and required input to the panel. 8. Determine the other temperatures that are required or devel
oped. 9. Design the system for heating the panels according to conven
tional practice. In the design steps, the effect of each assumption or choice
on comfort should be considered carefully. The following general rules should be followed:
1. Place panels near the cold areas where the heat losses occur. 2. Do not use high temperature ceiling panels in very low ceilings. 3. Keep floor temperatures at or below 85 F(29°C).
Letter Symbols for Examples of Design Methods The following design examples use the letter symbols shown
below.
r. c,=
c, =
/> = Qd~
Ide =
'•</ =
'dc =
'* =
'dmw
'/> =
panel area, ft2 (m2). coefficient of heat transfer from the upper surface of the concrete slab which forms the ceiling panel to air above panel at point tbi Btu/h-ft2-F (W/m2-°C); deg F (°Q represents temperature difference between panel surface and air. coefficient of heat transfer from lower surface of concrete slab to air below the panel at point tb, Btu/h«ft2-F (W/m2«°C); deg F (°C) represents temperature difference between panel surface and air. coefficient of downward and edgewise heat loss of exposed slab, (Btu/h«ffF (W/m«°C); ft (m) represents linear ft (m) of exposed slab perimeter and deg F (°Q represents temperature difference between concrete surface and outdoor air. length of exposed edge of slab, ft (m). downward heat flow from panel, Btu/h* ft2 (W/m2). apportioned downward and edgewise heat flow from panel, Btu/h-ft2 (W/m2). upward heat flow from panel, Btu/h«ft2 (W/m2). total thermal resistance of panel to downward heat flow, ft2-F'h/Btu(m2.°C/W). thermal resistance of material between the underside of the concrete slab and the ceiling surface below, ft2»F«h/Btu (m2'°C/W). thermal resistance of bare concrete panel to downward heat flow, ft2«F-h/Btu (m2«°C/W). total thermal resistance of panel to upward heat flow, ft2«F-h/Btu(m2.',C/W). thermal resistance of floor covering, ft2»F»h/Btu (m2*°C/W). thermal resistance of bare concrete slab to upward heat flow, ft2«F«h/Btu(m2-0C/W). design room air temperature, F(°C). outdoor design air temperature, F (°Q. air temperature above or below panel at point to which U, C|, or C2 is taken, F (°Q. inlet water temperature, F (°C). outlet water temperature, F (°Q. mean water temperature, F (°Q. maximum water temperature permissible for a given construction, F(°C). design mean water temperature (selected for each zone), F (°Q. panel surface temperature (exposed surface), F ("Q-surface temperature of top of concrete slab, F (°Q. overall coefficient of heat transfer for the given construction between room air and the point tb, Btu/h«ft2«F (W/m2«°C).
Procedure for Metal Ceiling Panels Design procedures for metal ceiling panels in a heating ap
plication are included in the example given in the section Design Procedure for Panel Cooling Systems.
Warm Water Panels—Embedded Pipe This section presents a simplified procedure for the thermal
design of embedded pipe, water-heated panels. The procedures are based primarily on data obtained at the former ASHRAE Research Laboratory.
A panel designed by this procedure will maintain the desired room air temperature for the selected outdoor conditions. Room air temperature is the selected criterion of comfort. The design procedure is restricted to situations in which the area-weighted average temperature of unheated surfaces of walls, glass and floor or ceiling does not differ greatly from room air temperature. Room-scale tests, which simulate various conditions of construction and outdoor temperature, have shown that this near-equality of the two temperatures normally prevails.
The procedure is applicable within the following range: Outdoor design conditions: Temperatures as low as -30 F
(-34.4"C). Room air temperature: 70 to 76 F (21.1 to 24.4°C). Air changes: No more than two air changes per hour. Room dimensions: Rooms having normal proportions; ceiling
height between 7 and 12 ft (2.1 and 3.7 m). Room construction: Any type of wall construction and any amount -
of glass area. (Both, however, have an effect on comfort.) Conventional interior finishes and furnishings.
Plaster Ceiling Panels." The procedure for designing a plaster ceiling panel is illustrated by Example 1.
Example 1: Three rooms, A,B and C, have a common water supply temperature; that is, they represent a single zone. They are maintained at 72 F (22.2°Q air temperature when the outdoor air temperature is 0 F (-17.8°C). The ceilings of rooms A and B have floors above them with the space heated to 72 F (22.2°Q and an air- to-air U value of 0.25 Btu/h • ft2 • F (1.42 W/m2 • °C). The ceiling of room C has insulation in the joist spaces and an uninsulated attic space with a combined (/value of 0.05 (0.28) from room C to outdoor air. Step I. Heat Loss
Calculate the heat loss of each room by methods outlined in Chapter 25 of the 1981 FUNDAMENTALS VOLUME, but do not include any heat loss through the area covered by the panel.
Room dimensions and calculated heat losses are as follows:
Room Dimensions ft(m)
Heat Loss Btu/h(W)
Room A 11x12x8(3.3x3.6x2.4) 6300(1850) RoomB 11x12x8(3.3x3.6x2.4) 2500(730) RoomC 15x21x8(4.5x6.3x2.4) 8000(2300)
Step 2. Required Panel Output Divide the heat loss of each room by the maximum ceiling area in
the room that can be used as a heating panel. The result is the minimum heat output per square foot (square metre) of panel that will satisfy the requirements of the room. The panel that requires the highest output per square foot (square metre) will generally control the design, because the temperature of the fluid in the system must be high enough to produce the required output from that panel. The required panel output is given in the following table.
Available Panel Area ft2(m2)
Required Panel Output, qd Btu/h-ft2(W«m2)
Room A 132(12.3) RoomB 127(11.8) RoomC 306(28.4)
6300/132 = 47.7(1850/12.3 = 150) 2500/127 = 19.7 ( 730/11.8 = 62) 8000/306 = 26.1 (2300/28.4 = 82)
Step 3. Panel Surface Temperature From Fig. 8, find the panel surface temperature needed to yield the
required heat output to each room, using the output determined in Step 2 and the design room air temperature. This use of Fig. 8 for Room A is illustrated in Fig. 8A.
The values of tp determined for Rooms A, Band C are as follows:
D-17
Panel Heating and Cooling Systems 8.17
I CONVERSION FACTORS: • C - I F - 321/1.8 t W/m"> Btu/h It-' X3.15
1
Fig. 8A Determination of Panel Surface Temperature for Room A from Fig. 8
Room A /,, = 114F(45.6°C) RoomB t„ = 92F(33.3°C) RoomC t„ = 97F(36.I°Q
Step 4. Upward Heat Flow Determine the upward heai flow from each panel using Fig. 24. The
upward heat flow must be determined to obtain the downward panel resistance and to select the proper size of boiler. The use of Fig. 24 is illustrated for Room A in Fig. 24A. Note that the U values shown on Fig. 24 are those for the entire structure of which the panel is a part and are taken from room air (temperature, ta) to air above the panel (temperature, /ft). For that reason, qu ±U(tp-1/,). U values may be obtained from Chapter 23 in the 1981 FUNDAMENTALS VOLUME.
The values of qu for Rooms A, B and C as determined* from Fig. 24 are as follows:
Room U tp-'b degFfQ Btu/h'ft2(W.m2)
A B C
0.25(1.42) 0.25(1.42) 0.05 (0.28)
42(23) 20(11) 97 (54)
12.5(39.4) 6.0(18.9) 5.0(15.8)
Step S. Downward Panel Resistance Assume tentative pipe or tube spacings for the panel construction to
be used from those listed in Table 3, choosing the closer spacings when higher outputs are required. Also in Table 3, find the resistance of each panel (rd) using heat flow ratios calculated from the flow rates determined in Steps 2 and 4, interpolating as required..
ROOM SPACING HEAT FLOW RATIO RESISTANCE, rd
in. (mm) ft2«F«h/Btu(m*«<'C/W) A 4.5(114) 12.5/47.7 (39.4/150) = 0.26 0.35(0.06) B 9 (229) 6/19.7 (18.9/62) = 0.30 0.90(0.16) C 9 (229) 5/26.1 (15.8/82) = 0.19 0.85 (0.15)
Step 6. Mean Water Temperature For the required panel output (qd) found in Step 2, the panel
resistance (rd) found in Step 5, and the room air temperature (/„), find the mean water temperature (tmw) from Fig. 8. This use of Fig. 8 is illustrated in Fig. 8B. Values of tmw for Rooms A, B and C as determined from Fig. 8 are as follows:
Room A tmw = 131 F (55°C), Room B tmw = 110.5 F (43.6°Q and Room C fmw = 119 F (48.3°Q. Step 7. Design Mean Water Temperature
Select a single design mean water temperature (jtjmw) f° r each group of rooms comprising a zone, choosing the highest mean water temperature (rm„) of the group subject to:
a. If tmw + l(f(- - /0)/2] is equal to or less than /„,„, this mean water temperature (tmw) is an acceptable design mean water temperature ltdmw)- Proceed to Step 8.
b. If fm„ + ((/, - r„)/2] is greater than !„,„, go back to Step 5 and select a panel construction having a lower panel resistance (rd). This can be accomplished by either or both:
1. Reducing the tube spacing. 2. Decreasing the upward heat flow (qa) by increasing the insula
tion above the panel. If the required mean water temperature is still too high, apply either
or both of the following:
lOO '20 140 CANU. SUWAt.l HMP MINUS AIR 7IMP ABOVt TANT I. VP Ub. f DEC
Fig. 24 Upward Heat Flow from Plaster Ceiling Panel
I'll.
o
ct
i
i
CONVERSION FACTORS: C - deg F/1.8
W/m ! • Btu/h • I t ' X3.1S W/nv C » Btu/h ft-" • F X 5 .68^
(U factor)
Fig. 24A Determination of Upward Heat Flow for Room A from Fig. 24
1. Reduce the heat loss of the room. 2. Provide supplementary heating. In Example I, assume t; - ;„ = 15 deg F (8.3°C) and tmax = 140 F
(60°C). Then since tmvl = 131 F (55°C), tm„ + 0.5 (/, - t0) = 138.5 F (59.1°C) and is less than /„„„. Thus, 131 F (55°C) can be used as the design mean water temperature tjmK. Step 8. Design Panel Output
From Fig. 8, find the panel output (qd) for design mean water temperature Crfm„), room air temperature (/„),. and panel resistance (rd)- Fig. 8C shows how Fig. 8 is used to find qd for Room B in this step.
The mean water temperature (tmw) found for Room A in Step 6 was used as the design mean water temperature Udmw)- The panel output determined in Step 2 for Room A is therefore the design panel output
H __477_
l = O j /
^ ^ " t > 4
"*Z?
t 1
1-
CONVERSION FACTORS: - C . | F - 3 2 1 / 1 . 8 WAtr - Btu/h • ft-" X 3.15 m'-'C/W-- f t ! - F • h/Btu X 0.176
Ir value)
Fig. 8B Determination of Mean Water Temperature for Room A from Fig. 8
I ,_ 32.5
•*^r -zZ, -C.o
2§S. -T i
7 0 -72>
CONVERSION FACTORS: "C=|F-32)/1.8 W/m:-Btu/h-ft--X3.1S tn!-'C/W- ftJ• F • h/Btu X 0.176
(rvalue)
4' Fig. 8C Determination of Design Panel
Output for Room B from Fig. 8
D-18
8.18 CHAPTER 8 1984 Systems Handbook for this room. Design panel outputs for Rooms B and C can be found from Fig. 8 using the design mean water temperature (tdmw).
Room Design Panel Output, qj Btu/h»ft2 (W»m2)
A B C
47.7(150) 32.5(102) 33.0(104)
Step 9. Design Panel Area Divide the room heat loss found in Step 1 by the design panel output
found in Step 8.
Room Design Panel Area, A. ft2 (m*)
A 132 (12.3) (unchanged from Step 2) B 77 ( 7.2) C 242(22.5)
Step 10. Total Panel Output Add the heal flow upward (?„) to the design panel output (qd) and
multiply by the design panel area to obtain the total panel output. If the design panel output is different from the panel output (qd) used in Steps 3 and 4, the heat flow upward (qu) should be redetermined.
Room
A B C
Qu
12.5 (39.4) 9.5f(29.9) 5.5f(17.3)
Qd
47.7(150.3) 32.5 (102.4) 33.0(104.0)
Ap
132(12.3) 77 ( 7.2)
242(22.5)
Total Panel Output
Ap (.Qu + Qd) Btu/h(W)
7946(2328) 3234 ( 948) 9317(2730)
t Redetermined.
Step II. Fluid Circuit Design the fluid circuit (panel piping and mains) for a temperature
drop of 10 to 20 deg F (5.6 to 11.1°C) between the water inlet and outlet of the panel (see applicable chapters on hot water systems). Step 12. Boiler Size
Size the boiler according to the method in Chapter 24 of the 1983 EQUIPMENT VOLUME. The net Btu (kj) rating of the boiler should equal or exceed the total output of all panels plus any other loads on the boiler.
Concrete Ceiling Panels. Concrete ceiling panels are distinguished from concrete floor panels in intermediate floors by the position of the tubes in the concrete slabs (see Tables 1 and 2). Both types of panels have heat outputs in two directions in amounts determined by the thermal resistance and the temperature difference in each direction. The effect of these outputs on space heating requirements and the occupants' comfort should always be considered. (See the section on Concrete Floor Panels— Intermediate Slab.)
The procedure for hot water plaster ceiling panels cannot be wholly applied to concrete ceiling panels because some of the simplifying assumptions regarding the upward heat flow from plaster panels are not valid for concrete panels. The plaster ceiling panel procedure must be modified as follows: Steps 1,2 and 3.
These steps are identical to the corresponding steps: for plaster ceiling panels. Step 4. Upward Heat Flow Estimate
a. If the upper surface of the slab is exposed to form a floor. And the heat flow upward from Fig. 9, using the panel surface temperature (tp) found in Step 3, and the air temperature of the space above.
b. I f the upper surface of the slab is not exposed, use:
<7«=C, (t„-tb) (11) Step 5. Upward and Downward Panel Resistance
Follow the procedure for Plaster Panels, Step S, using Table 2 to find both resistances (r„, and rd). Step 6. Mean Water Temperature and Upward Heat Flow
a. Follow the procedure for Plaster Panels, Step 6. b. Find the heat flow upward from the panel (qu) from Fig. 9. Add
to the thermal resistance of the slab to upward heat flow (ru), the resistance to heat flow (/•„.) of any material between the upper surface of the slab and the space above to obtain the resistance (r„) to be used in Fig. 9. The mean water temperature (tmw) found and the air temperature of the space above the panel are the other two factors to be used. Step 7. Design Mean Water Temperature
Select the highest mean water temperature (imw) as the design mean water temperature (/,/««,). Steps 8,9,10. II and 12.
These steps are identical to the corresponding steps for plaster ceiling panels.
Plaster Wall Panels. A design graph has not been prepared for wall panels, but a design can be approximated using the equations of heat transfer from walls together with the thermal resistance properties of plaster ceilings from Table 3. The procedure for plaster ceiling panels is used as a guide. Steps I and 2.
These steps are identical to the corresponding steps for plaster ceiling panels. Step 3. Panel Surface Temperature
Assume a trial panel surface temperature and determine the resulting heat output from Fig. 2 and 4 as explained in the accompanying section of the text. Assume successive trial panel surface temperatures until finding the temperature at which the combined heat transfer from the panel equals the output determined in Step 2. Steps 4 and 5.
These steps are identical to the corresponding steps for plaster ceiling panels. Step 6. Mean Water Temperature
For the required panel output found in Step 2, the panel resistance found in Step 5, and the room air temperature, calculate the required mean water temperature as follows:
tmw='p+rdlQd) (12) Step 7. Design Mean Water Temperature
Follow the procedure for plaster ceiling panels. Step 7. Step 8. Design Panel Output
From Eq. (12), and Fig. 2 and 4, find the panel output for design mean water temperature tdmw by successive trials. Steps 9,10,11 and 12.
These steps are identical to the steps for plaster ceiling panels. Concrete Floor Panels (Slab-On-Grade)
Step I. Heat Loss Calculate the heat loss of each room, but do not include any heat
loss through the area covered by the panel. If very large rooms are involved, the rooms should be subdivided into areas having somewhat similar heat requirements, i.e., separate the interior areas requiring little or no heat input from the exterior areas directly influenced by outdoor weather conditions. Treat each area as a separate room for design purposes. Step 2. Required Panel Output
Divide the heal loss of each room by the maximum floor area in the room that can be used as a heating panel. The result is the minimum heat output per sq ft (square meter) of panel that will satisfy the room requirements. The panel that requires the highest output per ft2(mJ) will generally control the design, because the temperature of the water in the system must be high enough to produce the required output from that panel. Step 3. Panel Surface Temperature
From Rg. 9, find the panel surface temperature needed to yield the required heat output to each room by using the output determined in Step 2 and the design room air temperature.
Floor panel surface temperatures exceeding about 85 F (29°Q are
D-19
Panel Heating and Cooling Systems 8.19 not recommended since they wilt probably cause foot discomfort. If the required heat output cannot be obtained from an 85 F (29°Q floor panel, heat losses should be reduced or supplementary heating should be provided.
Step 4. Downward and Edgewise Heat Flow a. Determine the temperature of the surface of the concrete slab (/,)
by adding to the panel surface temperature the temperature difference caused by the thermal resistance of the floor covering, if any:
's = 'p + (lu * rM) (13) The value of rue for various floor coverings is given in Table 5. b. Determine the downward and edgewise heat loss coefficient C3
from Fig. 11 for the insulation to be used. Insulation with a conductance of 0.4 Btu/h«ft2«F (2.3 W/m2-°Q extending 2 ft (0.6 m) below the slab results in a slab downward and edgewise heat loss coefficient of 0.97 Btu/h-ft2-F(5.5 w7m2-°C).
c. Apportion the downward and edgewise heat loss uniformly across the panel as follows:
Ode=PxCi{ts -toaVA,, (14)
Step 5. Upward Panel Resistance Assume a tentative pipe or tube size and a spacing for each panel,
choosing closer spacings and larger pipe or tube when higher heat outputs are required. From Table 1, find the thermal resistance (rm) of the slab of each panel, using heat flow ratios (qu/Qde) calculated from the flow rates determined in Steps 2 and 4, interpolating as required. To the slab resistance (r^). add the resistance of the floor covering (r^.. Step 4) to obtain the panel resistance to upward heat flow (/•„).
Step 6. Mean Water Temperature For the required panel output (qu) found in Step 2, the panel
resistance (/•„) found in Step 5, and the room air temperature (ta), find the mean water temperature Umw) from Fig. 9.
Step 7. Design Mean Water Temperature Select a single design mean water temperature (tdmw) for each group
of rooms comprising a zone, choosing the highest mean water temperature (/„,„.) of the group. In the example, 116.5 F (46.9°Q is used. StepS. Design Panel Output
From Fig. 9, find the panel output (qu) for design mean water temperature (tdmw), room air temperature (ra), and panel resistance (ru). Also from Fig. 9, find the design panel surface temperature. If it exceeds 85 F (29.4°C) (see Step 3) in a room, go back to Step 5 and choose a wider spacing or smaller pipe or tube for that room. Step 9. Design Panel Area
Divide the room heat loss found in Step I by the design panel output found in Step 8. Step 10. Total Panel Output
Add the apportioned downward and edgewise heat flow (<7rfc) to the design panel output (qu) and multiply by the design panel area to obtain the total panel output. If the design panel output is appreciably different from the panel output (qu) used in Steps 3 and 4, the apportioned downward and edgewise heat flow (q^ should be redetermined. Step 11. Fluid Circuit
Design the fluid circuit (panel piping and mains) for a temperature drop of 10 to 20 deg F (5.6 to 11.1°Q between the water inlet (/,) and outlet (t0) of the panel (see Chapters 15 and 16). Step 12. Boiler Size
Select the boiler size according to the method in Chapter 24 of the 1983 EQUIPMENT VOLUME. The net Btu (kJ) rating of the boiler should equal or exceed the total output of all panels plus any other loads on the boiler.
Concrete Floor Panels (Intermediate Slab). The procedure for designing a hot water concrete floor panel of intermediate slab type is given in the following steps. Steps 1,2 and 3.
These steps are identical to the corresponding steps for slab-on-grade construction. Step 4. Down ward Heat Flow Estimate
a. Follow procedure for slab-on-grade construction, Step 4, part a.
b. Estimate the heat flow downward as follows: 1. If the underside of the concrete slab is exposed to form a ceil
ing, find the heat output downward (qd) from Fig. 8, using the slab surface temperature (ts) found in Step 3 as the panel surface temperature (tp) and the air temperature of the space below.
2. If the underside of the concrete slab is not exposed, use the equation:
Qd = C2 Us ~ lb) C5) Step 5. Upward and Downward Panel Resistance
Follow the procedure for slab-on-grade construction. Step 5, using Table 1 to find both resistances (rm and r^). Step 6. Mean Water Temperature and Downward Heat Flow
a. Follow the procedure for slab-on-grade construction. Step 6. b. Use Fig. 8 to find the downward heat flow (qd). Add to the slab
resistance to downward heat flow (r^) the resistance to heat flow (r^) of any material between the underside of the slab and the ceiling surface, to find the total resistance to downward heat flow (rd) to be used in Fig. 8. The water temperature (tmw) found above and the air temperature (/6) of the space below the ceiling are the other two factors to be used.
If the heat flow downward (qd) differs appreciably from the estimate made in Step 4, repeat Steps 5 and 6 using the calculated value. Steps 7,8 and 9.
These steps are identical to the corresponding steps for slab-on-grade construction. Step 10. Total Panel Output
Follow the procedure for slab-on-grade construction, Step 10, substituting heat flow downward (qd) for the apportioned downward and edgewise heat flow {q^). Steps 11 and 12.
These steps are identical to the corresponding steps, for slab-on-grade construction.
Insulated Electric Ceiling Panels
Electrically heated ceiling panels can be designed using the equations and curves previously presented in this chapter. However, for ceiling panels which have the thermal insulation recommended for electric heating, U = 0.05 Btu/h«ft2«F (0.28 W/m2 ' °C), the reverse heat loss is considered negligible. Therefore, the design heat loss is most commonly calculated in the normal manner, using the design heat transfer coefficients as given in the 1981 FUNDAMENTALS VOLUME.
The design heat loss is then converted to watts in the ratio of 3.413 Btu/h per watt [watt output/watt input (W0/Wj)] and the nearest standard cable assembly is selected. Cable length is ordinarily published as part of the manufacturers* rating data. Cable spacing is calculated from Eq. (10). Safety factors are usually omitted in the design of electric heating systems.
Electric Ceiling Panels on Masonry Slab (Intermediate Floors)
The following steps give the design procedure for ceiling heating panels that consist of electric heating cable embedded in plaster applied to the lower surface of intermediate floor, masonry slab construction:
Step 1. Calculate the heat loss of each room, but do not include any heat loss through the area covered by the panel.
Step 2. Divide the heat loss of each room by the net ceiling area that can be used as heating panel.
Step 3. Determine the combined thermal resistance, Ru, of the slab material or materials above the heating cable. Do not include the surface film resistance.
Step 4. Determine the required watts input per ft2 (m2) of panel area. Use the heat loss per ft2 (m2) of panel area from Step 2 and the upward thermal resistance from Step 3. Use either Fig. 25 or Eq. (16) and (17) as follows:
a. If Fig. 25 is used:
D-20
8.20 CHAPTER 8 1984 Systems Handbook
5 10 IS 2 0
HEAT INPUT, WATTS PER SO FT.W;,
•75 F ambient and 75 AUST above and below slab.
Fig. 25 Ceiling Panel Output for Intermediate Floor, Masonry Slab Construction*
Enter the chart on the vertical scale at the required ceiling panel heating output (from Step 2). At the intersection of the required heat output and the curve of thermal resistance for the particular construction, drop to the horizontal scale and find the necessary Wj/ft2
(Wj/m2). b. If Eq. (16) and (17) are used: Determine the value of m from Eq. (17). With m and the room heat
loss per ft2 (m2) of panel area from Step 2, calculate watts input from Eq. (16). Note: Eq. (17) can only be used for values of Ru between 0.5 and 6.0 (0.09 to 1.06).
Step 5. Determine the approximate floor surface temperature from Fig. 26. If the resulting floor temperature exceeds 85 F (29.4°C), one of the following steps should be taken:
a. Reduce the room heat loss by increasing the building insulation. b. Increase the insulating value of the slab above the insulation.
This may be done by using insulating concrete, floor coverings or other means.
c. Supplement the panel heat output with another form of electric heat.
Step 6. Multiply the required watts input per ft2 (m2) by the net panel area available to obtain the total watts required for the room.
Step 7. Select from manufacturers' ratings the standard cable assembly that most closely matches the required total watts.
Step 8. Determine the proper spacing between adjacent cable runs using Eq. (10).
Watt input can be determined from Eq. (16) and (17).
W, = qc/m (16)
m = 1.344 + 0.854 Ru - 0.175 Rul + 0.013 Ru3 (17)
[0.5 (0.09) S * „ £ 6.0 (1.06)]
where
W, = watts input per ft2 (m2) of panel area. qc => design heat loss per ft2 (m2) of panel area; also ceiling panel
output, Btu/h-ft2(W/m2). m = ceiling panel output factor, Blu/h per W, (Wa/Wi).
5 *
4 1
"J
/.
1 80 f/
/ i ezr/
/ /
/ ..../
85 f /
/ CON\ ' C» - nr B
m : t
1 /
/ '
/ j , 4
i /ERSION FACTORS: f - 3211.8 ( I X 0.0929 -y w - h : F h/BuiX 0.176
Ir value)
CO IS 2 0
HEAT INPUT, WATTS PER SQ F*. * .
* 75 F ambient and 75 AUST above and below slab.
Fig. 26 Upper Surface Temperature of Intermediate Floor, Masonry Slab Structure with Embedded
Cable Ceiling Panel*
Ru = the combined or equivalent thermal resistance of the slab material above the heating cable, ft2-F«h/Btu (m2-°CAV). Ru does not include the surface film resistance.
The heat output to the space above the slab equals the difference between the heat energy input and the ceiling panel output.
Figures 25 and 26 and Eq. (15) and (16) were derived using Eq. (3), (7) and (8). This procedure is based on the following conditions:
1. Steady-state heat flow. 2. A slab of infinite length and width. 3. Ambient temperature of 75 F (23.9°C) both above and below the
slab. 4. AUST of 75 F (23.9°C) both above and below the slab. 5. Electric heating cable embedded in a material having a thermal
resistance below the cable of 0.12 (0.02). This resistance does not include the surface film resistance.
The procedure can be used for small deviations from these conditions. For other conditions, the basic panel heat output equations should be used.
Electric Floor Slab Heating
In electric floor slab heating, it is important to thermally isolate the heated slab from the adjacent floor, exposed edge and earth. A large amount of the heat is lost through the exposed edge of a floor. A portion of this loss is directly through the concrete to the air, and the remainder is lost through the slab for a distance of approximately 3 ft (0.9m) from the exposed edge. This edge loss is proportional to: (1) the difference between indoor and outdoor temperatures, (2) the length of the floor perimeter adjacent to the exposed edge and (3) a heat loss factor (K), based on the amount of heat loss through the floor area included within a 3-ft (0.9-m) border along the exposed edge, in Btu/h (W) per ft (m) of exposed edge, deg F (°C) difference between indoor and outdoor temperatures.'
H^KUti-O (18) where
D-21
Panel Heating and Cooling Systems 8.21 H = heat loss through the exposed edge, W. tf = W/ft.F<\V/m-°C). L = exposed edge, ft (m). /,- = indoor design temperature, F (°Q. t„ = outdoor design temperature, F(°Q.
The heat loss through the inner slab at a distance beyond 3 ft (0.9 m) of the exposed edge is independent of outdoor temperature. Previous work on the subject,17,18 indicates that heat losses through the inner floor [area beyond 3 ft (0.9 m) of the edge] remain nearly constant at approximately 0.6 W/ft2
(6.5 W/m2) or (conservatively) 1 W/ft2 (11 W/m2). Thus, the total heat loss through a slab on grade can be divided into two segments: (1) the heat loss through the exposed edge including the floor area within 3 ft (0.9 m) of the edge and (2) the heat loss through the remaining (inner) floor area, 1 W/ft2 (11 W/m2).
H,=KL(t,-t0)+\A (19)
where H, = total slab loss, W. A - inner floor area = total floor area minus floor area included
within 3-ft (0.9-m) border along exposed edge, ft2 (m2).
The remainder of the heat generated in the inner floor will be stored and, under certain conditions of temperature difference, will dissipate into the area to be heated by radiation and convection heat transfer. The rate of heat loss from the surface of the inner slab is proportional to the temperature difference between the floor surface temperature and air temperature. The rate of heat loss is also affected by the floor material (concrete, asphalt or tile). Assuming a steady-state condition, the heat output of the inner floor equals the heat input minus the downward heat loss.
K = h, -h„ = U(ts - ta) (20)
where h0 = heat output, W/ft2 (W/m2). h, = heal input. W/ft2 (W/m2). hd = downward heat loss, W/ft2 (W/m2). U = heat transmission coefficient, W/ft>F(W/m2•°C). ts = floor temperature, F (°C). ta = air temperature, F(°C).
The heat transmission coefficient U has been determined17
to be 0.5 W/ft2 (5.4 W/m2). Assuming the heat loss for the inner floor to be 1 W/ft2 (11 W/m2):
h,-,= h0 + 1 = 0.5 (/, - / . ) + 1 (21)
In SI, ^f = h0 + 11 = 2.84 (/s - t.) + 11 (21a)
Electric slab heating systems can also be designed by using part of the procedure for hot water panels as a guide. Steps 1,2 and 3.
Follow the procedure for hot water panels. Step 4 Slab Heat Loss
Determine the heat loss from the slab, using the slab surface temperature found in Step 3 and the factors in Chapter 23 of the 1981 FUNDAMENTALS VOLUME.
Using electric slab heating as a primary system, a design range of 10 to 15 W/ft2 (108 to 161 W/m2) of slab area is the maximum input recommended. For slab tempering in kindergartens, bathroom floors, basements and family rooms, a range of 1 to 5 W/ft2 (11 to 54 W/m2) is recommended. For spot heating in unheated, poorly insulated buildings (farrowing, brooding and milking parlors) an input of 30 to 40 W/ft2
(323 to 431 W/m2) is a normal design range.
Warm Air Panels The first three steps in the design of warm air panels are the
same as those outlined for hot water panels and the same per
formance curves can be used. The balance of the design can be determined from the data in Chapter 11.
PANEL COOLING SYSTEM DESIGN
Design Steps for Metal Ceiling Panels Panel design requires specifications of panel area, panel
type, supply water temperature, water flow rate and panel arrangement. Panel performance is directly related to room conditions. Air-side design also must be established. The procedure is:
1. Determine room design dry-bulb temperature, relative humidity and dewpoint.
2. Calculate room sensible and latent heat gains. 3. Select mean water temperature for cooling. 4. Establish minimum supply air quantity. 5. Calculate latent cooling available from the air. 6. Calculate sensible cooling available from the air. 7. Determine panel cooling load. 8. Determine panel area for cooling. 9. Calculate room heat loss.
10. Select mean water temperature for heating. 11. Determine panel area for heating. 12. Determine water flow rate and pressure drop. 13. Design the panel arrangement.
Step I: Design Conditions Referring to Example 1 for design procedures for a plaster ceiling
heating panel system, the three rooms. A, B and C, are each to be maintained at 78 db and 45% rh, resulting in a- dewpoint of 55 F (12.8°C). If room db and rh are allowed to drift for energy conservation purposes, the highest expected dewpoint should be used. -
Step 2: Heat Gains Calculate the room heat gains by the methods outlined in the 1981
FUNDAMENTALS VOLUME. The room heat gains are:
Dimensions, ft (ro)
Sensible Gain. Blu/h(W)
Latent Cain, Btu/h (W)
A I l x | 2 x 8 ( 3 . 4 x ] . 7 x 2 . 4 > 3.500(1026) 400(117) B 1 1 x 1 2 x 8 ( 3 . 4 x 3 . 7 x 2 . 4 ) 4400(1319) 500(147) C 1 5 x 2 1 x 8 ( 4 . 6 x 6 . 4 x 2 . 4 ) 14.000(4 102) 1.000(293)
9 30
>
A ' • & -
j .
X *
s*
A
-yd? \X
&^-—
» /
j fc>
s vV
^ *$r
^
CONVERSION FACTORS: -C-degF/1.8 War - BtttA) • H* X 3.tS
m-'-C/W-lf fMBt»X0. i ;6 bate)
S 10 15 20 MEAN WATER TEMPERATURE DIFFERENCE.
25 F 0EG.
•Derived from Fig. 10. Fig. 27 Metal Ceiling Panel Design Graph Showing Panel
Surface Temperature and Mean Water Temperature Difference vs. Panel Cooling Performance*
D - 2 2
8.22 CHAPTER 8 1984 Systems Handbook Step 3: Select Mean Water Temperature
Room dewpoint temperature.
Inlet water temperature. Assume a 5 deg F (2.8°Q temperature rise giving a mean water temperature.
F(°Q = 55(12.8)
1 ( 0.5) / , = 56(13.3)
, = 58.5 (14.7)
Step 4: Minimum Supply A ir Quantity Minimum supply air quantities should be based on the recommend
ed practices as given in the ASHRAE HANDBOOK and as dictated by local codes. Minimum supply air quantities are:
Room Minimum Supply Air Quantity, cfm (L/s)
A B C
80(38) 80(38)
190(90)
Step 5: Latent Cooling Assume a central station air system supplying conditioned air to the
rooms at 58 F (14.4°Q db and 53.5 F (U.9°C) wb, resulting in a moisture content of 54.0 gr/lb (7.72 g/kg) of dry air, and thereby a moisture difference between room and supply air conditions of 10.0 gr/lb (1.43 g/kg) of dry air. The latent load capacity ofthe'air is:
Room
Air Flow Rale.
cfro(LA)
(' Room \ _ /" Supply Air \ Humidity I I Humidity I
gr/lb (g/kg) air
Supply Air
Latent Cooling,
Biu/h(W)
80(18) 80(18)
190(90)
(64.0 - 54.0)0.68 (64.0 - 54.0)0.68 (64.0-54.0)0.68
544.0(159.4) 544.0(159.4)
1.292.0(178.6)
In all three rooms, the moisture pickup of the air is sufficient to offset the room latent gain. If this were not the case, adjustments in supply air quantity or design temperatures would have to be made accordingly.
Step 6: Sensible Coolingfrom Air
Air Flow Rate
cfm (L/s)
( Room \ / Supply Air \ Temp I I Temp I
F f O
Supply Air
Sensible Cooling,
Blu/h(W)
80(18) 80(18)
190(90)
(78.0 - 58.0) I.I (78.0-58.0)1.1 (78.0-58.0)1.1
1.760(516) 1.760(516)
4.180(1225)
Step 7: Panel Load The panel load is the room sensible load minus the cooling done by
the air.
( Room \ X s u p p l y A i r \
Sensible 1 _ I Sensible 1 Heat Gain, I I Cooling. I Blu/hfW) / V B i u / h f W ) /
Pand Cooling Load,
Btu/h(W)
3.500(1 026) 4,500(1 119)
14,000(4 102)
- 1,760( 516) - 1,760 ( 516) - 4,180(1 225)
1,740 ( 510) 2,740 ( 801) 9.820(2 877)
Step 8: Panel Area The panel area is a function of the mean water temperature and the
type of panel used. Fig. 27 shows design panel performance for three types of ceiling panel systems. The room temperature minus the mean water temperature is 78 - S8.5 = 19.5 deg F(I0.8°Q. The panel cooling performance at 19.5 deg F (10.8°Q temperature difference for the three types of panels is 18.0, 24.0 and 34.0 Btu/h-ft2 (56.7, 75 6 and 107.1 W/m2), respectively.
Based on the available ceiling area, the panel cooling available for Room A is:
Pand Resistance
rt2>F-h/Btu (m3-*C/W)
Pand Performance.
Btu/h.ft2
(W/m2)
Available Pand AreaL
f t 2 ' '(m2>
Pand Coating
Available. Btu/h (W)
0.61 (0.107) 0.11 (0.055) 0.071 (0.012)
18.0 ( 56.7) 24.0 ( 75.6) 14.0(107.1)
112(12.1) 112(12.1) 112(12.1)
2.176 ( 697) 1.168 ( 910) 4.488(1 117)
All three types of panels exceed the panel cooling load, reducing the design panel area as follows:
ROOM A Pand
Resistance, f l 2 -F .h /Btu (m2 -'C/W)
Panel Cooling Load.
Btu/fc(W)
Pand Cooling
Performance. B m / h - f t 2 ( W . m 2 )
Pand Area.
ft2 (m2)
0.61 (0.107) 0.61 (0.055)
0.071 (0.012)
1.740(510) 1.740(510) 1.740(510)
I8.0( 56.7) 24.0( 75.6) 14.0(107.1)
97(9.0) 71(6.8) 5114.8)
Similarly for Room B:
ROOMB Panel
Resistance. f !2 .F.h/Btu (m2 -"C/W)
0.61 (0.107) 0.11 (0.055) 0.071 (0.012)
Panel Performance.
Blu/h.fc2
(W/m2)
18.0 ( 56.7) « 24.0 ( 75.6) " 34.0(107.1) «
Available Panel Area,
ft2 (in21
127(11.8) 127(11.8) 127(11.8)
Pand Cooling
Available. Btu/h (W)
2.2861 670) 1.048| 891) 4.118(1 265)
Panel type [r = 0.61(0.107)] is deficient in capacity and panel type lr - 0.31 (0.055)1 exceeds the cooling required. Since these panels have the same appearance and construction, a combination panel may be used. The calculations are: Panel Cooling Load = 2,740 Btu/h (803 W) Panel Cooling Available,
from typer = 0.61 (0.107)
Panel Cooling Required from typer = 0.31 (0.055)
Area of Panel Typer = 0.31 (0.055)
= -2,286 Btu/h (670 W)
= 454 Btu/h (133 W) = 4S4
^ ^ 76 ft2 (7.1m2) (24-18)
Thus, the panel area for Room B consists of 81 ft2 (7.5 m2) of panel type r = 0.31 (0.055) plus 127 - 81 = 46 ft2 (4.3 m2) of panel type r = 0.61 (0.107). K panel type r = 0.071 (0.012) is used, the cooling available exceeds the panel cooling load, and the panel area is reduced to:
2,772/34.0 = 82 ft2 (7.6 m2) The data for Room C are:
Panel Resistance,
f t 2 .h -F /Btu (m2 . 'C/W)
0.61 (0.107) 0.11 (0.055) 0.071 (0.012)
ROOMC Panel
Performance. Btu/h-ft2
(W-m 2 )
18.0 ( 56.71 24.0 ( 75.6) 14.0(107.1)
XX
X
Panel Area
Available. f t 2 (m 2 )
106(28.4) 106(28.4) 106(28.4)
Panel Cooling
Available. Btu-hlW)
5.508(1 612) 7.344(2 148)
10.404(1 045)
The panel cooling available from either panel type, r = 0.61 (0.107) or r = 0.31 (0.055) does not satisfy the panel cooling load for Room C in Step 7. If either of these two ceiling types is still desired for architectural reasons, additional cooling must be provided by increasing the supply air quantity. Assuming panel type r = 0.31 (0.055), the calculation is:
Room sensible load = 14,000 Btu/h (4102 W) Panel cooling = - 7,344
Air cooling required = 6,656 Btu/h (1950 W)
Revised air flow rate = 6 ' 6 5 6 = 308 cfm (145 L/s) (20xl.i)
If panel type r = 0.071 (0.012) is used, the cooling available satisfies the panel load and no increase in air flow rate is necessary.
Step 9: Room Heat Loss See Step I for plaster ceilings. Heat losses are the same.
D - 2 3
8.24 CHAPTER 8 1984 Systems Handbook
C.S. Leopold: The mechanism of heat transfer panel cooling heat storage (Refrigerating Engineering, July 1947, p. 33).
C.S. Leopold: The mechanism of heat transfer panel cooling heat storage—Part II. Solar radiation (Refrigerating Engineering, June 1948, p. 571).
C.S. Leopold: Design factors in panel and air cooling systems (ASHAE Transactions, Vol. 57.1951. p. 61).
L.F. Schutrum, John Vouris, and T.C. Min: Preliminary studies of heat removal by a cooled ceiling panel (ASHAE Transactions, Vol. 61,1955. p. 95).
L.F. Schutrum and T.C. Min:. Lighting and cooled air effects on panel cooling (ASHAE Transactions, Vol. 64,1958, p. 189).
R.E. Boyar: The influence of radiant energy transfer on human comfort (Heating, Piping & Air.Conditioning, June 1966, p. 109).
T.C. Min, L.F. Schutrum, G.V. Parmelee and J.D. Vouris: Natural convection and radiation in a panel-heated room (ASHAE Transactions, Vol. 62,1956, p. 337).
W.F. Spiegel: A water-cooled luminaire in a panel-air system (ASHAE Transactions, Vol. 64,1958, p. 351).
L.F. Schutrum and T.C. Min: Cold wall effects in a ceiling-panel-heated room (ASHVE Transactions, Vol. 63,1957, p. 187).
Cyril Taster, C M . Humphreys, G.V. Parmelee, and L.F. Schutrum: The ASHVE environment laboratory (ASHVE Research Report No. 1444, ASHVE Transactions, Vol. 58.1952, p. 139).
P.O. Fanger: Calculation of thermal comfort: Introduction of a basic comfort equation (ASHRAE Transactions, Vol. 73, Part 11, 1967, p. 111.4.1).
P.E. McNall, Jr., and R.E. Biddison: Thermal and comfort sensations of sedentary persons exposed to asymmetric radiant fields (ASHRAE Transactions, Vol. 76,1970, p. 123).
J .C. Schlegel and P.E. McNall, Jr.: The effect of asymmetric radiation on the thermal and comfort sensations of sedentary subjects (ASHRAE Transactions, Vol. 74,1968, p. 144).
A.P. Gagge, G.H. Rapp, and J.D. Hardy: The effective radiant field and operative temperature necessary for comfort with radiant heating (ASHRAE Journal, May 1967, p. 63).
D-25
Panel Heating and Cooling Systems
t'wrtoe) " I I I I I 1 I I 1 1 100 120 140 160 180 200
TP 0 R TM»T
* Derived from Figs. 2 and 4, based on 70 F and 70 AUST.
Fig. 28 Metal Ceiling Panel Design Graph Showing Panel Surface Temperature and Mean Water Temperature
vs. Panel Heating Performance*
Step 10: Mean Water Temperature Typical design panel performance at various mean water tempera
tures is given in Fig. 28, which is similar to Fig. 8 except the performance is based on the convection output from Fig. 4 for the Design curve.
The suggested basis for selecting the mean water temperature is that the area of panel adjacent to and within approximately 2 to 3 ft (0.6 to 0.9 m) of the outside wall should have an output of not less than 50% of the wall transmission heat loss. For example. Room A transmission loss equals the heat loss of 6,300 Btu/h (1846 W); and the panel area within 4 ft (1.2 m) of the outside wall is 4 x n = 44 ft2 (1.2 x 3.3 = 4 m2). A water temperature is selected to give a panel output of approximately (6300 x 0.5/44) = 71.6 Btu/h-ft2 [(1846 x 0.5/4) = 2300 W/m2).
From Fig. 28, a mean water temperature of 164 F (73.3°Q for ceiling type r = 0.61 (0.11) will satisfy the requirement. This represents a suggested minimum mean water temperature for comfort. The temperature may be increased using experience and judgment if desired. By inspecting the data for rooms B and C and Fig. 28, a mean water temperature of 164 F (73.3°C) is selected.
Step 11: Heating Panel Area The heating panel area for each room is given below. For purposes
of simplification, assume Room A ceiling to be aluminum panels with pipes 12 in. (305 mm) on center. Room B ceiling to be aluminum panels with pipes 6 in. (152 mm) on center, and Room C ceiling to be aluminum sheet with copper tubes4 in. (102 mm) on center.
Room
A B C
Heai Loss,
Btu/h <W)
6.300(1850) * 2.500 ( 730) * 8.00012300) v
Panel Performance.
© . „ , „ - 1 6 4 F(73.fC) Btu/n.f l2 (W/m2 )
71 (224) 90(284)
119(375)
Panel Area.
ft2(ra2)
89(8.3) 27 (2 J ) 67(6.2)
Step 12: Hydraulic Data
The water flow rate may be based on a 20 deg F (1 l.l°C) temperature difference between inlet and outlet temperatures or any temperature difference suitable to the design of the piping distribution system. The panel pressure drop must be obtained from manufacturer's data.
Step 13: Panel Arrangement The panel and piping arrangement must be designed to accommo
date the various elements in the ceiling such as lights, air outlets, speakers, sprinkler and smoke devices. This should be done in cooperation with the architect and manufacturer. As can be seen from Steps
8.23
8 and 11, the panel area for heating may or may not be the same for cooling.
Room
A B
C
Cooling Panel Area,
ft2 (mJ)
98(9.1) 46(4.3) 81 (7 J )
291 (27.0)
Healing Panel Area,
ft2(m2>
89(8.3)
27(2.5) 67(6.2)
Ceiling Type
r
0.61 (0.107) 0.61(0.107) 0.31 (0.055) 0.071 (0.012)
For Room A, the areas are close enough so that the larger area is used as a common panel for heating and cooling. For Room B, two panels are designed; one panel for heating and cooling = 27 ft2 (2.5 m2), and one panel for cooling only = (81 - 27) + 46 = 100 ft2 (9.2 m2). See Fig. 23 for typical split-panel arrangements. Room C is similar to Room B.
REFERENCES 1 Standard 55-74, Thermal Comfort Conditions (ASH RAE, 1974). 2 P.O. Fanger: Thermal Comfort Analysis and Applications in En
vironmental Engineering (McGraw Hill, Inc., New York, NY, 1972). 3F. W. Hutchinson: Influence of gaseous radiation in panel heating
(ASHVE Transactions, Vol. 53,1947, p. 285). 4T.C. Min et al: ASHAE Research Report No. 1576-Natural con
vection and radiation in a panel heated room (ASHAE Transactions, Vol. 62,1956, p. 337).
5G.V. Parmelee and R.G. Huebscher: Forced convection heal transfer from flat surfaces (ASHVE Transactions, Vol. 53, 1947, p. 245).
6G.B. Wilkes and C.M.F. Peterson: Radiation and convection from surfaces in various post ions (ASHVE Transactions, Vol. 44, 1938, p. 513).
7L.F. Schutrum and CM. Humphreys: ASHVE Research Report No. 1499—Effects of non-uniformity and furnishings on panel heating performance (ASHVE Transactions, Vol. 60,1954, p. 121).
»L.F. Schutrum and J.D. Vouris: ASHVE Research Report No. 1516—Effects of room size and non-uniformity of panel temperature on panel performance (ASHVE Transactions, Vol. 60,1954, p. 455).
9 L.F. Schutrum, G.V. Parmelee, and CM. Humphreys: ASHVE Research Report No. 1473—Heat exchanges in a ceiling panel heated room (ASHVE Transactions, Vol. 59,1953, p. 197).
10 L.F. Schutrum, G.V. Parmelee, and CM. Humphreys: ASHVE Research Report No. 1490—Heat exchanges in a floor panel heated room (ASHVE Transactions, Vol. 59,1953. p. 495).
"H.H. Macey: Heat loss through a solid floor (Institute of Fuel Journal, 22-128, p. 369).
12 E.L. Sartain and W.S. Harris: Performance of covered hot water floor panels. Part 1-Thermal characteristics (ASHAE Transactions, Vol.62,1956, p.55).
13 A Subcommittee of the TAC on Panel Heating and Cooling, R.L. Maher, Chairman; W.P. Chapman; H.T. Gilkey; P.B. Gordon; E.F. Snyder; and J.M. Van Nieukerken; and by ASHAE Laboratory Staff Members, L.F. Schutrum and CM. Humphreys: ASHAE Research Report No. 1600—Thermal design of warm water concrete floor panels (ASHAE Transactions, Vol. 63,1957, p. 239).
14Standard Specifications for Gypsum Plastering, Including Requirements for Lathing and Plastering (American Standards Association, A 42.1,1946).
13 R.E. Boyar: Room temperature dynamics of radiant ceilings and air conditioning comfort systems (ASHRAE Transactions, Vol. 69, 1963, p. 37).
'*A Subcommittee of the TAC on Panel Heating and Cooling, R.L. Maher, Chairman; W.P. Chapman; H.T. Gilkey; P.B. Gordon; E.F. Snyder; and J.M. Van Nieukerken; and by ASHAE Laboratory Staff Members, L.F. Schutrum, G.V. Parmelee, and CM. Humphreys: ASHAE Research Report No. 1559—Thermal design of warm water ceiling panels (ASHAE Transactions, Vol. 62,1956, p. 71).
"H.D. Bareither, A.N. Fleming, and B.E. Alberty: Temperature and Heat-Loss Characteristics of Concrete Floors Laid on the Ground (Research Report'48-1, Small Homes Council, University of Illinois).
I 8F.C Houghten et al: Heat loss through basement floors and walls (ASHVE Transactions, Vol. 48,1942, p. 369).
BIBLIOGRAPHY C.S. Leopold: Hydraulic analogue for the solution of problems of
thermal storage, radiation, convection and conduction (ASHVE Transactions, Vol. 54,1948, p. 389).
D-24
