TABLE OF CONTENTS

hammercoupleMécanique

22 févr. 2014 (il y a 3 années et 3 mois)

447 vue(s)

Chapter 53 reformatted july 4th

Error! Reference source not found.

Peter Simmonds

Page
i

TABLE OF CONTENTS
1.0 INTRODUCTION

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....

1

2.0 ASPECTS OF RADI
ANT SYSTEMS

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................................
.....

1

3.0 RADIANT HEAT EX
CHANGE

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................................
................................
...............

3

4.0 THERMAL COMFORT


................................
................................
................................
.........................

3

5.0 RADIANT TEMPERA
TURE ASYMMETRY

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................................
..........................

4

5.1

V
ERTICAL AIR TEMPERAT
URE DIFFERENCE

................................
................................
....................

4

5.1.1 Architecture of Radiant Ceilings
................................
................................
...............

4

5.2

D
ESIGN AND
D
IMENSIONING

................................
................................
................................
.........

5

5.3

C
OOLING

................................
................................
................................
................................
.....

6

5.4

H
EATING

................................
................................
................................
................................
.....

6

6.0 DESIGN ASPECTS
OF RADIANT CEILING S
YSTEMS

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................................
......

7

6.1

H
YDRAULICS

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................................
................................
...............................

7

6.1.1 Hydraulic loop connect
ions

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................................
......................

7

6.2

A
COUSTIC
F
EATURE OF
R
ADIANT
C
EILING
P
ANELS

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................................
........

7

6.2.1 Acoustic Inlay Mats

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................................
................................
..

7

6.2.2 Acoustic Fleece

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................................
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........

7

6.2.3 Panel Perforation

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................................
.....

7

6.3

C
ONTROLS

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................................
................................
................................
.

8

6.3.1 Two port control valves

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................................
............................

9

6.3.2 Controlling water temperature / Injection circuit

................................
....................

11

6.4

E
NERGY
S
AVINGS WITH
R
ADIANT
C
EILING
S
YSTEMS
................................
................................
....

12

6.5

S
ELECTING
P
ROCEDURE OF
R
ADIANT
C
EI
LING
P
ANELS

................................
...............................

12

6.5.1 Design Example: Classroom
................................
................................
..................

12

6.5.2 Design Example: Office

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.........................

16

6.5.3 Condensation Control

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................................
............................

18

7.0 EMBEDDED SYSTEM
S

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................................
................................
......................

20

7.1

INTRODUCTION


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................................
................................
................................
..........

20

7.2

B
ACKGROUND

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................................
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...........................

20

7.3

F
UNDAMENTAL CALCULATI
ONS
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....

21

7.3.1 Convection:

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................................
...........

21

7.3.2 Long wave radiation:

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................................
..............................

21

7.3.3 Short wave radiat
ion:

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................................
.............................

22

7.4

C
ONCEPT OF THE METHOD

TO DETERMINE THE HEA
TING AND COOLING CAP
ACITY

.........................

22

7.4.1 The heat transfer coefficient is combined convection and radiation.

...................

23

7.5

S
IMPLIFIED CALCULATIO
N METHODS FOR DETERM
INING HEATING AND CO
OLING CAPACITY OR
SURFACE TEMPERATURE

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................................
................................
.....

24

7.5.1 Universal single power function

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................................
.............

25

7.5.2 Thermal resistance methods
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................................
.................

26

7.6

U
SE OF BASIC CALCULAT
ION PROGRAMS

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................................
....................

28

7.6.1 Basic calculation programs

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................................
...................

28

7.6.2 Calculation of the heating and cooling capacity
................................
....................

28

7.6.3 Calculation of the heat flow intensity

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................................
.....

29

7.
6.4 Reference heating systems

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................................
...................

29

7.6.5 Limits of heat flow intensity

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................................
....................

32

7.6.6 Thermal conductivity of screed with fixing inserts
................................
..................

34

7.7

G
ENERAL RESISTANCE ME
THOD

................................
................................
................................
..

48

7.7.1 Pipes embedded in wooden construction

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..............................

53

7.7.2 Determination of heat exchange by calculati
on

................................
.....................

53

7.7.3 The calculation model

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................................
...........................

53

Chapter 53 reformatted july 4th

Error! Reference source not found.

Peter Simmonds

Page
2

7.7.4 Calculation procedure for determination of equivalent heat transmission coefficient
53

7.7.5 Maximum heat flow density to the room

................................
................................

54

7.7.6 Thermal resistance above the heat conduction layer

................................
............

55

7.7.7 Thermal resistance

on the back
-
side of the heat conducting layer

.......................

55

7.7.8 Thermal resistance between the heat source and the heat conducting layer

.......

56

7.7.9 Pipe coils

................................
................................
................................
................

57

7.8

C
ALCULATION METHODS

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................................
.............

62

7.8.2 Simplified model based on finite difference method (FDM)

................................
...

68

7.8.3 Limits of th
e method

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................................
...............................

71

7.8.4 Dynamic building simulation programs

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................................
..

71

7.8.5 Simplified diagrams

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................................
................................

72

7.8.6 Calculation m
ethod

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................................
................................

76

7.9

E
MBEDDED
S
YSTEMS
C
ONTROLS

................................
................................
................................
.

3

7.9.1 Central control

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................................
................................
..........

3

7.9.2 Individual control

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................................
......

4

7.9.3 Room thermostats/sensors

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......................

4

7.9.4 Floor mass

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...............

5

7.9.5 Time delay, time response

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.......................

5

7.9.6 The duty cycle

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................................
..........

5

7.9.7 Self
-
regulating effect

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................................
................................

6

7.10

R
ADIANT
C
OOLING
S
YSTEM

................................
................................
................................
........

6

7.10.1 C
ontrols

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................................
................................
..................

6

7.10.2 Control of TABS

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................................
.....

6

7.10.3 Temperature Differences and Flow Rates

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.............................

7

7.11

R
ADIANT
F
LOOR
H
EADE
R

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................................
................................
..........

8

7.11.1 Control strategy

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................................
................................
......

9

7.11.2 Control strategy for office buildings

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................................
.....

12
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1.0

I
NTRODUCTION

RADIANT based HVAC systems use low temperature fluid for heat
ing and high temperature fluid for cooling to directly
condition surfaces which can be externally mounted or embedded pipe arrangements where the surface temperatures
are controlled to below 300°F
1

and where the exchange of energy from or to the surface ex
ceeds 50% by radiation.
2




Radiant energy is transferred in the absence of solid mass, between different surfaces of different temperatures
by short and long electromagnetic waves.



This energy transfer occurs in straight lines, and can be reflected, absorb
ed, reradiated and transmitted.



Absorbed energy is converted to heat; emitted (reradiated) energy is a release of heat.



Transmitted energy can be interstitial heat transfer between spaces and or heat loss and heat gain through the
building enclosure.

Purp
oses

The deliberate control of radiant energy absorbed or emitted from a surface is for the purposes of directly conditioning
the occupants and building mass; and for using the lowest possible fluid temperature in heating and highest temperature
in cooling
.In both cases there is delivered


thermal comfort, energy efficiency and exergy effectiveness.



Thermal Comfort


radiant based HVAC system work effectively to satisfy the radiant comfort elements as
defined by the requirements of ANSI/ASHRAE Standard 55


Thermal Environmental Conditions for Human
Occupancy and ISO 7730
-

Ergonomics of the thermal environment
--

Analytical determination and
interpretation of thermal comfort using calculation of the PMV and PPD indices and local thermal comfort
criteria.



E
nergy Efficiency


due to the typically low operating temperatures in heating, and high operating temperatures
in cooling, radiant based HVAC systems enable combustion and compression equipment to achieve its
engineered maximum rated potential as defined b
y the objectives of ASHRAE Standard 90.1 and have lower
transmission losses as per ASHRAE 152
-

Method of Test For Determining The Design and Seasonal
Efficiencies of Residential Thermal Distribution Systems. These equipment friendly fluid temperatures can

be
further optimized by the application of ASHRAE 189.1 where increases in building efficiency promote even
further reductions in heating fluid temperature and increases in cooling fluid temperatures.



Exergy effectiveness
-

due to the low operating temper
atures in heating, and high operating temperatures in
cooling, radiant based HVAC systems can be coupled without combustion or compression to low grade sources
of energy (geothermal or solar) or facilitate the drawn down of high quality source temperatures

(generated by
combustions) through cascading thermal systems afforded through power production and waste heat recovery.
3


Back ground considerations



Radiant heating and cooling ‘system’ consists of thermally conditioned surfaces, a source of heated and ch
illed
fluid, distribution systems and controls.

2.0

A
SPECTS OF
R
ADIANT
S
YSTEMS

While different buildings have different requirements, the most important requirement remains the well being of
occupants. Proper control not only prevents problems associated wit
h an unhealthy environment, it provides occupants
with the best possible environment in which to work.

Many factors, such as health, activity levels, clothing, and social situations influence the well
-
being and productivity of
building occupants. Environm
ental conditions that are influenced by environmental systems and controls can include:



Air quality, temperature, velocity, turbulence, humidity



Temperature distribution



Temperature of surrounding areas



Noise and Lighting



Arrangement of the room (size,
color, view)

Radiant
has

been used extensively in
Asia and Europe for centuries.

There has been an
adoption

of radiant
applications in the US
beginning in the early 1900’s.
4

.

When designing
modern
radiant
systems for heating
and cooling
there are

few
considerations that haven’t already be
en
resolved including the management of
moisture

in cooling applications

and the requirements of ventilation with
dedicated outdoor air systems.
.





1


There is design consideration with radiant cooling for dew point control.

2


For higher temperature systems see Chapter 53, R
adiant Heating And Cooling, 2007

ASHRAE Handbook, HVAC Applications

3


ASHRAE Technical Group 1
-

Exergy Analysis for Sustainable Buildings

4


Bean, R., Olesen, B., Kim, K., History of Radiant Heating and Cooling, ASHRAE Journal, Part 1, Jan. 2010, Part 2,

Feb.
2010
.

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Background

There have been many previously published ASHRAE papers describing the performance and control of radiant
systems. Udagawa 1993, Zweizel and Kochenz 1993 described the radiant cooling per
formance of
suspended
ceiling
panels. Brunk 1993 described the energy savings as a result of using
such
systems
. Kochendorfer 1996 described the
method of testing ceiling panels to validate their performance. Braun 1990 showed that energy c
osts could be reduced
by using the building mass. There have been many papers written on radiant
floor and ceiling
slabs that describe their
performance and effectiveness such as Athienitis and Show 1991, Gerpen and Shapiro 1981, Kalisperis, Steinman,
Summ
ers and Olesen 1990, Leigh 1991, Lingh and Defenbach 1990, MacCluer 1989, 1991, Olesen 1994.

Papers have been written on the performance of radiant ceilings and their effectiveness in creativity and controlling a
comfortable indoor climate for occupants, S
immonds, 1991 (a), 1991 (b), 1992, 1993 (a), 1993 (b), 1993 (c), 1994, 2003
(a), 2003 (b).

Harmon and Yu 1993 discussed condensation issues with cold air distribution systems. Mehlomakulu, Chambers,
Gautrey and Simmonds 2005 described control strategies to

operate radiant
panels and avoid condensation.



(radiant is not a “concept”, is not “new” and it does not need to have “potential” also we
need to focus on controlling moisture with less emphasis on the panel temperature


the
panel didn’t put th
e moisture in the space so why do we keep focusing on the panel as the
problem? Control the moisture for health (mold, mites, bacteria, viruses) and comfort (skin
wettedness) benefits, benefits to hygroscopic materials (dimensional stability)


my edits
b
elow are a reflection of this. RBean)


Applications
Based on
the global
drive for high performance buildings,
current application
knowledge

and readily
available

equipment

there are few limitations for radiant heating and cooling systems.




To obtain the maximum performance of
panels especially in cooling mode the
space as well as the
surface temperature
of
panels

to ensure that the panel surface temperature does not drop below the dew point temperature of the space.


The control of space humidity levels and panel surface temperature can be controlled by traditional HVAC
methodologies
. The suppl
y air temperature and volumetric flow can be selected to balance space air moisture content.



Figure
1

shows a typical layout of a
suspended
ceiling panel

system.


Most air conditioning systems use the principle of forced air. In forced air systems, warm room air is recirculated, chilled,

and thrown back into the space. With the increased use of electronic equipment and other office technology, higher and
h
igher heat loads need to be removed from these environments. To remove this additional heat, a higher volume of
cooler air is blown into the working environment to maintain comfort, making draft
-
free cooling with chilled, primary air
increasingly more diff
icult.


Radiant systems contain heated or cooled water running through pipes that are bonded to the non
-
visible side of the
panels or embedded in a material such as concrete.. The radiant systems function as heat exchangers between the
room air and the ea
ting or cooling water. The radiant system emits or absorbs heat from heat sources in a room and
exchanges it with the circulating water.



The heat emitted in the room is radiated from a warm body to the cool surface or from warm surface to a cool body.

3.0

R
ADIANT HEAT EXCHANGE

With radiant cooling, heat exchange in the room happens mainly via radiation and convection. Radiation of energy takes
place between objects with different surface temperatures. In all cases, the warmer object radiates heat to the coo
ler
object. Just as the hot sun radiates to the cooler earth, in a radiant cooled environment, computers, people, and other
sources of heat radiate that heat to the cooler surface of the ceiling. As an added benefit, humans perceive heat transfer
via radia
tion as particularly comfortable. Convection occurs when the room air is cooled as it flows beneath the cooling
panels. The cooler air is heavier than the warmer air rising from the heat sources, which creates natural, high volume,
low
-
velocity air current
s.

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4.0

T
HERMAL
C
OMFORT

Thermal comfort can be defined as “the psychological condition that expresses satisfaction with the thermal
environment.” (ASHRAE Standard 55
-
2004) Therefore, thermal comfort would be evaluated by asking all the occupants
if they are s
atisfied with their thermal environment. However, in order to design and control radiant heating and cooling
systems, it is necessary to predict the thermal comfort in a room without resorting to a polling result. To provide an
acceptable thermal environm
ent to the occupants, the requirements for general thermal comfort (PMV, operative
temperature, etc.) and local thermal comfort (surface temperature, vertical air temperature differences, radiant
temperature asymmetry, draft, etc) must be taken into accoun
t.

In order to provide acceptable thermal conditions, two parameters, air temperature and mean radiant temperature,
should be taken into account. The combined influence of these two temperatures is expressed as the operative
temperature. This means that a
ir temperature and mean radiant temperature have an equal importance with respect to
the level of thermal comfort in a space. Compared with a convective heating and cooling system, a radiant heating
system can achieve the same level of operative temperatur
e at a lower air temperature and a radiant cooling system at
a higher air temperature.

Operative temperature is defined as the temperature of a uniform isothermal black enclosure in which the occupant
exchanges the same amount of heat by radiation and c
onvection as in the actual non
-
uniform environment. In a physical
sense, operative temperature is the temperature that the occupant perceives his or her surroundings based on
convection and radiation. Operative temperature can be said as the weighted avera
ge of the air temperature and the
mean radiation temperature (MRT). The radiant systems have low MRT which improves comfort conditions.

5.0

R
ADIANT TEMPERATURE A
SYMMETRY

In all practical thermal environments, a radiation field has an asymmetric feature to some

degree. If the asymmetry is
sufficiently large, then it can cause discomfort. For example, discomfort might be felt by persons exposed to the
asymmetric radiation such as the open door of a furnace, direct sunlight, heated ceilings or cooled windows or wa
lls.

The radiant temperature asymmetry is defined as the difference between the plane radiant temperatures of the two
opposite sides of a small plane element. And the detailed calculation method of percentage dissatisfied due to the
radiant temperature asy
mmetry is described in ASHRAE Handbook Fundamentals Chapter 4.

Human body is most sensitive to radiant asymmetry caused by warm ceiling or cool walls
-
windows. Thus, in case of
making ceilings heated by applying the radiant heating, the radiant temperature
asymmetry should be maintained less
than 5°C (in relation to a small horizontal plane 0.6m above the floor). In case of making walls
-
windows cooled by
applying the radiant cooling, the radiant temperature asymmetry should be less than 10°C(in relation to a

small vertical
plane 0.6m above the floor). Figure 1 shows discomfort level (percentage dissatisfied) due to radiant temperature
asymmetry in case of ceiling cooling/heating and walls
-
windows cooling/heating.



Figure
2

local Th
ermal discomfort caused by radiant asymmetry

5.1

V
ERTICAL AIR TEMPERAT
URE DIFFERENCE

One of the important features of radiant heating and cooling system is that it is possible to get the uniform temperature
conditions from floor to ceiling. According to measur
ements, the application of the floor heating/cooling and the large
wall panel heating units under the window have a uniform temperature profile. The more convective systems (baseboard
under window, warm air system or blackwall) or high temperature systems
result in 2
-
3°C gradients between floor and
ceiling and even up to 7°C in the much severe case. For systems relying on much more convection, the temperature
profile becomes less uniform.

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Thermal stratification that results in the air temperature at the hea
d level being warmer than at the ankle level may
cause thermal discomfort. Thermal stratification in the opposite direction is rare, but is preferred to by occupants.
Therefore it is not addressed in this standard.

The differences in air temperature form
the ankle level to the head level are recommended to be within 3°C. Figure 2
can be used in conjunction with the PPD limit for vertical temperature differences to determine the allowable ranges of
vertical temperature differences.

5.1.1

A
RCHITECTURE OF
R
ADIANT
C
EILINGS

Ceiling radiant cooling panels (CRCP) are generally built as an architectural finish product (with necessary acoustical
qualities, color, and pattern), compatible with the traditional drop ceiling “tee grid” system or as a free hanging element.

Typ
ically a copper tube is embedded into an extruded aluminum saddle which is permanently affixed to the back of a
architectural metal ceiling panel. The process how the copper coil is thermally bonded with the radiant ceiling panel is
crucial especially in c
ooling application.. Panel piping arrangements are generally in a serpentine pattern; however,
parallel header arrangements are also available. Typical panel construction is illustrated in Figure 1. As installed, the
“drop in” radiant panels weigh 1.6
-

2
lb/ft2A radiant ceiling panel also features acoustic value. The acoustic signature
can be achieved with a variety of perforations. In addition a fiber glass blanket is placed on the pack of the panel.



Figure
3

shows a typical

composition of a radiant ceiling panel

The lightweight construction results in a transient response “time constant” of only about 3
-
5 minutes. That means they
respond rapidly to changing space sensible load conditions. Hydraulically, the ceiling panels ar
e most frequently
connected with flexible
-
push on coupling hoses for fast and safe installations, as illustrated in Figure 2.


Figure
4

shows the flexible connections to a radiant ceiling panel.

Panels can be moved aside without

disconnecting the hoses, for easy access above. They can also be easily removed
and reconnected for either extensive maintenance or evolving space use requirements without breaking normal
threaded or sweat solder plumbing connections.

Radiant ceiling pane
ls promote architectural freedom in several ways. First, radiant cooling ceilings can be designed to
be visually indistinguishable from regular ceilings to maintain aesthetic appeal. Secondly, the hardware in a radiant
cooling system is smaller and more fl
exible. The size of the ceiling panels, the arrangement of the appliances within the
ceiling, and the partitioning of the room all are flexible. Minimal airflow requirements often permit designs where
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common mechanical air distribution devices such as diff
users and ductwork are made much smaller or are eliminated
altogether.

5.2

D
ESIGN AND
D
IMENSIONING

(?? MB text) intro to chapter section


Figure
5

SHOWS TYPICAL CONNECTIONS TO RADIANT CEILING PANELS

5.3

C
OOLING

In practice, the design
cooling capacity per unit panel area (Btu/h ∙ ft2 [W/m2]) is determined from the panel
manufacturer’s catalog data rated for the test standard, such as EN 14240 and ANSI/ASHRAE Standard 138
-
2005,
Method of Testing for Rating Ceiling Panels for Sensible Hea
ting and Cooling.8
i

The unit panel cooling capacity can be
selected from the design capacity tables provided by the panel manufacturer based on the difference between the room
temperature and the mean panel surface temperature (or mean water temperature M.
W.T.). However, the
manufacturer’s data rated in their standard test chamber is usually 5% to 30% less than the actual capacities measured
in the real space after installation because the tests are performed in a test chamber under ideal conditions (no
mec
hanical ventilation and adiabatic walls). Consequently, the required panel area may be overestimated, and the initial
and operating costs for the CRCP system also may be increased.

This over
-
design problem in the CRCP system can be avoided by estimating t
he design panel cooling capacity for the
real operating conditions. Recently, the unit cooling capacity of top insulated panels lying on conventional false ceiling T
-
bar grid in a mechanically ventilated space was proposed
ii

as a function of eight variable
s including tube spacing (w),
panel thickness (δ), panel thermal conductivity (k), panel inlet chilled water temperature (Tfi), room temperature (Tsp),
room position (interior or perimeter space with or without fenestration), diffuser discharge air velocit
y (V), and diffuser
characteristic width (W).

5.4

H
EATING

In many climates, radiant cooling ceilings can also provide heat. The structure should be well insulated, and outside
temperatures should not get extremely low. In these cases, radiators are not neede
d, saving costs and making more
floor space available.

Because humans sense heat from a hot ceiling more quickly than from hot air blown in from ducts, surface temperatures
should not exceed 95°F (35°C). The reason the surface should

NOT EXCEED 95°F (35°
C) is to permit humans to radiate a small amount of heat from their heads that are normally at
99°F (37°C) to a slightly cooler surface. If the surface temperature exceeds 95°F (35°C), humans in the space will
experience discomfort and headaches. This is o
nly if the panel is right on the head even a foot or two separation should
allow higher temperatures of 110
-
120 max for normal ceiling heights of 9
-
11 ft

Radiant panels radiate heat to the surrounding surfaces including people, furnishings, and the cold in
terior surface of the
window. The cold window surface, without radiant heating, loses heat via transmission to the cold outside. Without
radiant heating, the cold window surface cools the air passing over it, creating a significant convection down draft.
N
ormally In a baseboard radiator system the baseboard radiators normally balance this action. With radiant heating, the
interior surface of the glass is warmed, significantly reducing the convection down draft produced. The small down draft
current produced
, typically less than 20
-

30 fpm (10
-

15 cm/sec), creates a vacuum effect under the radiant heat
panels. The low
-
pressure area formed draws the warm room air at the ceiling towards this area where it is further
warmed by the radiant heat panels. As a res
ult, the air becomes even more buoyant, and it has a tendency not to
convect down the outer wall. The net effect is a surface boundary layer of “dead air” created very near the window. The
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down draft is below the minimum threshold felt by humans. Therefore
, the baseboard radiators can often be eliminated,
saving capital costs and increasing the usable floor area.

Another consideration is that, in the heating mode, convection is lower, and the ceiling’s capacity is proportionally
reduced. In most cases, the

internal heat loads can provide sufficient heat. To prevent cold airdrop from the windows,
the radiant heating portion of the ceiling needs to be installed along the perimeter of the room (typically within 3 feet of
the exterior wall).

6.0

D
ESIGN
A
SPECTS OF
R
ADIANT
C
EILING
S
YSTEMS

Technically speaking, a radiant cooling ceiling is simply a large heat exchanger suspended from a room’s ceiling. It
exchanges energy via with the room via radiation and convection. Accordingly, cooling ceilings can be rated.

Heat
transfer from the room to the ceiling surface is a function of the average ceiling surface temperature and the room
temperature. The objective of every radiant cooling ceiling developer is to get the ceiling surface temperature as close
as possible to the
water temperature.

The smaller the temperature difference between chilled water and ceiling surface, the more efficient the system.

Heat exchange between ceiling panels and water piping obeys the following equation:

Q = k * A * Dt.

Where Dt is the small
est possible temperature difference, the removed heat load, Q, shall be as large as possible.
Therefore, the k
-
value and the energy conducting areas, A, must be made as large as possible.

Panel dimensions can be chosen freely. The only design constraint t
o the architect is that the non
-
visible side must be
flat.

6.1

H
YDRAULICS

6.1.1

H
YDRAULIC LOOP CONNEC
TIONS

Flexible hoses hydraulically connect the ceiling panels. Flexible hose with quick couplings may be applied with radiant
panel system. These fittings are high
-
quality designed, precision machined, push
-
on couplings to ensure a leak
-
proof
system and enable simple, fast and safe installation.


Figure
6

shows the push pull flexible connections for radiant panels

Hose with working pressu
re is typically less than 25 psi (1.6 bar).push
-
on couplings A special hose tested to 220 psi (15
bar) is available at higher cost for special applications.

6.2

A
COUSTIC
F
EATURE OF
R
ADIANT
C
EILING
P
ANELS

Room acoustics also are handled conventionally. The fol
lowing addresses commonly applied strategies for improving
acoustics in rooms with cooling ceilings.

6.2.1

A
COUSTIC
I
NLAY
M
ATS

Normally, mineral fiber mats with 30mm (1.18 in) thickness and 2.5 lbs/ft3 (40 kg/m3) weight are used. The mats are
wrapped and sealed

for air tightness in plastic PE foil to prevent the formation of rub
-
off fiber particles.

6.2.2

A
COUSTIC
F
LEECE

In environments such as clean rooms and hospitals where fiber rub
-
off is not expected under any circumstances,
acoustic fleece may be used. The blac
k fleece is bonded to the rear of the perforated ceiling panel. To prevent a
decrease in the heat exchange between the heat conducting rails and the ceiling panel, the fleece is bonded between
the heat
-
conducting rails. Acoustic fleece must have a minimum
plenum height of 12 inches (30cm) to be effective.

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6.2.3

P
ANEL
P
ERFORATION

The ceiling panels might feature certain perforation pattern to enhance acoustic performance or personalize the visual
aspect of the ceiling. The perforation pattern is often specified w
ith hole diameter and free area (open cross sections
etc.).

6.3

C
ONTROLS

The design of control system shall take into account the building, its intend use and the effective functioning of the
heating/cooling system, efficient use of energy and avoiding heat
ing/cooling the building to full design conditions when
not required. This shall include keeping distribution heat losses as low as possible, e.g. reducing flow temperature when
normal comfort temperature level is not required. Control and operation of th
e system help us to handle the conditioning
systems with savings of operational costs and en able the maintenance of required indoor environmental conditions.

Radiant cooling controls have two primary functions: controlling room temperature and preventing
condensation on the
ceiling surface. For proper operation and maximum energy savings, radiant ceiling systems require the use of precision
electronic or direct digital controls.








Figure
7

shows a typical control schematic
for a radiant system




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Figure
8

shows an advanced control system for a radiant system (some items removed for clarity)

Controls are generally applied to influence both pressure (flow) and temperature.

Accordingly, there are t
wo control strategies:



Using a 2 port control valve to control water flow and



Using an injection circuit to control water temperature.

6.3.1

T
WO PORT CONTROL VALV
ES

The two
-
port valve controls the space temperature by permitting more or less water to flow thr
ough the valve and
through the radiant ceiling. The supply water temperature stays constant. A humidity sensor closes the valve as soon as
the chilled water supply temperature reaches dew point.

This type of control is very affordable and simple.

The only

disadvantage is dew point control. The cooling ceiling must be turned off as soon as a risk of condensation
occurs.














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Figure
9

Figure
10


























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6.3.2

C
ONTROLLING WATER TEM
PE
RATURE
/

I
NJECTION CIRCUIT

Using an injection circuit


requires a circulating pump to provide a constant flow of water through the ceiling. Depending
on the heat loads, a two
-

or three
-
way valve injects more or less chilled water to the ceiling supply. T
he same water
quantity is sent to the system return. The supply water temperature for the ceiling is controlled by the quantity of the
water injected. If acute danger of condensation occurs, the temperature of the supply water can be raised. The ceiling
lo
ses some of its capacity but can be kept in operation.






Figure
11

Figure
12





The injection circuit ensures both safe operation and the maximum possible mean temperature difference for cooling
ceil
ings. As each control zone needs a pump, a humidity sensor and a controller (which constantly calculates dew point
and compares it with the supply water temperature), this type of control is more costly than the 2
-
port control scheme.



Is identical to Fi
gure 9/10



Figure
13

Figure
14








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Figure
15

Figure
16

6.4

E
NERGY
S
AVINGS WITH
R
ADIANT
C
EILING
S
YSTEMS

6.5

S
ELECTING
P
ROCEDURE OF
R
ADIANT
C
EILING
P
ANE
LS

The concept of a dedicated outdoor air system (DOAS) with parallel sensible cooling was born from the decoupled
system concept, which can be summarized as decoupling of ventilation and air
-
conditioning functions, or decoupling of
sensible and latent loa
d functions. First, remove the latent loads from the outside air (OA) intake and generated in
spaces using a 100% OA ventilation system (i.e., DOAS). Second, remove the space sensible loads using a parallel
mechanical cooling system, such as fan coil units
, conventional variable air volume, and ceiling radiant cooling panel
(CRCP) independent of the ventilation system.

Load calculations are produced with the Trane Trace program for example. From the results the sensible cooling load is
transferred to room
detail sheet. The sensible loads then divided into two parts:



The sensible provided by the radiant panels.



The sensible load provided by the ventilation air.


The output from the radiant ceiling is deducted from the sensible room load; the remaining load
(if any) is to be removed
by the ventilation air. At this point the balance between minimum outside air requirements determined by the number of
occupants at 20 cfm per person and the air required to meet the load. Whichever is the larger air supply is use
d. In the
event that the outside air ventilation air is larger than the depicted volume, then it is possible to reduce to the area and
output from the radiant ceiling to meet the required load.

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6.5.1

D
ESIGN
E
XAMPLE
:

C
LASSROOM

Load calculations require that the
space sensible load of 28,584 BTUH. The space has 30 occupants.

A 30’ x 30’ x 10’ classroom and a maximum occupancy of 30 people, is to be maintained at 74°F and a 50% RH

(see
S.P.

2,
F
igure ___)
. From the psychometric chart, this gives a dew point tempera
ture of

54.2
°F
(
see
S.P. 3)
and a
moisture content of


0.0090
lb of moisture/lb of dry air.

Step 1.

Determine the sensible and latent hourly heat gain for the room.

The sensible and latent hourly heat gains are found using accepted procedures f
ound in the ASHRAE handbooks. For
this example, we will assume:

Sensible hourly heat gain: 28,584 BTUH

Step 2.

Determine the minimum air supply required for the room.

According to ASHRAE tabulated data, the recommended air supply per person is

20 cfm per
person x 30 persons = 600 cfm

Step 3.

Determine the latent load capacity of the primary air.

30 occupants at 220 BTU/pp = 6600 BTU/pp the moisture gain is 6600/4840 x 600 = 0.0022 lb/lb. The moisture content
of the
entering supply air

is
-
0.0090

0.0022 =
0 0.0068

lb/lb.

(
see
S.P. 1)



Specify

1.

the dry bulb and
calculated
relative
humidity

of the entering supply air

conditions

(
see
S.P. 1)


from the
psychrometric chart

2.

55°F
t
db

@ 73.9

RH

= entering conditions of th
e supply air
the


dew point
and
operating

dry bulb surface
temperature
of the
radiant cooling panel
.

a.

Select for

2
°F
or greater,
above the space operating
dew point
conditions of 54
.2
°F/54.2°F

(
see
S.P.
3
)

b.

54.2
°F + 3°F = 57.2°F

= lowest operating
surface
temperature of the cooling panel

(
see
S.P. 4
-
5)





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50
55
60
65
70
75
80
DRY BULB TEMPERATURE - °F
.002
.004
.006
.008
.010
.012
.014
.016
.018
10% RELATIVE HUMIDITY
20%
30%
40%
50%
60%
70%
80%
90%
45
50
50
55
55
60
60
65
65
70
70
HUMIDITY RATIO - POUNDS MOISTURE PER POUND DRY AIR
2
3
1
4
5
Example: Classroom, Legend
S.P. 1 Supply air conditions to space
S.P. 2 Operating air conditions
S.P. 3 Dew point temp. of operating air
S.P. 4 - 5 Operating and dew point temp. of the cooling panel
R
R
ASHRAE PSYCHROMETRIC CHART NO.1
NORMAL TEMPERATURE
BAROMETRIC PRESSURE: 29.921 INCHES OF MERCURY
Copyright 1992
AMERICAN SOCIETY OF HEATING, REFRIGERATING AND AIR-CONDITIONING ENGINEERS, INC.
SEA LEVEL

Figure
17

shows a psychrometric chart used to determine dew point temperatures of space temperatur
es based on
operating conditions.

The space condition is calculated by including the sensible and latent loads to the space. From this condition the dew
point temperature is shown on the psychometric chart. The panel surface operating temperature is select
ed about 3°F
higher than this temperature.

Step 4.

Determine the sensible load capacity of the primary air




60



spply
room
p
s
t
t
QpC
q

Qs = 600 cfm x 0.075lb/ft3 x 0.24 BTU/lb°F x (74
-
55) x 60

Qs = 12,312 BTUH.

The sensible cooling required from the panels = 28,584


12,312 = 16,
272 BTUH.

Step 5.

Determine the supply water temperature

(This
section
needs
a review

panel temp 3°F above space dew point
also =

fluid
supply
temp
?

Marcus? Peter?
)

It is recommended to allow an offset between the space dew point and chilled water supply
temperature. We select a
panel supply temperature 3°F higher to the space operating dew point temperature. Dew point is
54.2°F
(
S.P. 3)
).

Chilled water supply temperature:
54.2
+ 3 =
57.2
°F.


Step 6.

Determine the mean water temperatur
es M.W.T.

The ASHRAE Design Graph uses the Mean Water Temperature (MWT). The MWT is defined according to the follow
equation: (?? Equ)

tMEAN


Mean water temperature F

tEWT Entering water temperature F

tLWT

Leaving water temperature F



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The mean w
ater temperature MWT is the difference between supply and return water temperature. Therefore we
determine next the water inlet and the outlet water temperature. Typically the MWT namely the difference between
supply and return water temperature is selecte
d between 2°F
-

8°F. We select a return water temperature that is 65°F.

The MWT = (tEWT + tLWT) / 2 = (61 + 65) / 2 = 63°F

The MWT has an impact on several parameters such as the water velocity, the pressure drop, and the approach
temperature that eventu
ally determine the radiant ceiling panel performance.

Water velocity: when we decrease the differential temperature (supply
-
return water) we increase the water velocity
(assuming everything else remains in the system), as we increase water velocity we incr
ease the pressure drop. The
higher the water velocity the higher the friction loss and more pump energy is required. Typically a hydronic system is
designed at pressure drop rates between 2.1


3.7 psi. If flow rate results in pressure drop beyond 3.7 psi
the additional
capacity gained by the radiant ceiling panel is insignificant but the pressure drop increases significantly, second power
(that takes energy).


Approach temperature: the mean water temperature impacts the approach temperature which impacts t
he panel
performance. E.g. if the supply water temperature remains 61F but the mean water temperature is lowered from 4F to
2F the approach temperature dT is 75F


((61F+63F)/2) = 13F compare to 75F


((61F+65F)/2) = 12F. Note: see
consequences of pressure

drop above.

As result the radiant panel loop is design according to the pressure drop, flow rate, and water velocity. The maximum
water flow velocity is typically determined by the type of hose and fittings (hydronic connections between radiant ceiling
pa
nels) connections due to flow noise. The water flow rate is determined by the pressure drop as function of the
Reynolds number (to maintain turbulent water flow) and water temperature difference of supply and return.

Radiant ceiling panel manufacturer Typ
ically pressure drop for radiant systems: 2.1


3.8 psi

Step 8.

Determine the radiant panel output

Using manufacturer selection charts or characteristic graphs we determine the radiant panel performance based on our
design fundamentals. From the Figure 3
, we can see


Figure
18

SHOWS THE PANEL OUTPUT IS 30 btuH/FT2 AT A ROOM TEMPERATURE OF 72F and a MWT of 59F

Step 9.

Determine the required panel area

From figure 3, the cooling capacity is determined to be 26 BTUH/ft2. The req
uired sensible cooling output from the panel
is 16,272 BTUH; therefore, the required panel area is = 16,272 / 26 = 625 square
-
feet.

The classroom has a ceiling area of 30’ x 30’ = 900 square
-
feet. The panels cover 69% of the ceiling area. The required
sens
ible cooling output from the panel is 16,272 BTUH and therefore the required panel area is = 16,272 / 26 = 625sf.



i


ANSI/ASHRAE Standard 138
-
2005,
Method of Testing for Rating Ceiling Panels for Sensible Heating and Cooling.

ii


Jeong, J.W. and S.A. Mumma. 2004. “Simplified cooling capac
ity estimation model for top insulated metal ceiling radiant coolin
g
panels.”
Applied Thermal Engineering
24(14):2055

2072.


ASHRAE. 1992. Thermal environmental conditions for human occupancy. ANSI/ASHRAE
Standard
55
-
1992.


Athienitis, A.K. and J.D. Dale. 1987. A study of the effects of window night insulation and low emi
ssivity coating on heating load and
comfort.
ASHRAE Transactions

93(1A):279
-
294.

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NOT BE OTHERWISE POS
TED OR DISTRIBUTED.









Boyd, R.L. 1962. Application and selection of electric infrared comfort heaters.
ASHRAE Journal

4(10):57.


Buckley, N.A. and T.P. Seel. 1987. Engineering principles support a
n adjustment factor when sizing gas
-
fired low
-
intensity infrared
equipment.
ASHRAE Transactions

93(1):1179
-
1191.


Fanger, P.O. 1973.
Thermal comfort.

McGraw
-
Hill, New York.


Fanger, P.O., L. Banhidi, B.W. Olesen, and G. Langkilde. 1980. Comfort limits for
heated ceiling.
ASHRAE Transactions

86(2):141
-
156.


Gagge, A.P., G.M. Rapp, and J.D. Hardy. 1967. The effective radiant field and operative temperature necessary for comfort wit
h radiant
heating.
ASHRAE Transactions

73(1):I.2.1
-
I.2.9; and
ASHRAE Journal

9(
5): 63
-
66.


Greffet, J
-
J., R. Carminati, K. Joulain, J
-
P.
Mulet, S. Mainguy, and Y. Chen. 2002. Coherent emission of light by thermal sources.
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-
64.


ISO. 1994. Moderate thermal environments

Determination of the PMV and PPD indices and specific
ation of the conditions for thermal
comfort.
Standard

7730
-
1994. International Standard Organization, Geneva.


Isoardi, J.P. and J.P. Brasselet. 1995. Surface cooling in hospitals: Use of computer
-
aided design packages.
ASHRAE Transactions

101(2):717
-
720.


Madsen, T.L. 1976. Thermal comfort measurements.
ASHRAE Transactions

82(1):60
-
70.


Petráš, D. 2001. Hybrid heating and ventilating large industrial halls connected to district energy systems.
ASHRAE Transactions

101(1):390
-
393.


Rapp, G.M. and A.P. Gagge.

1967. Configuration factors and comfort design in radiant beam heating of man by high temperature infrared
sources.
ASHRAE Transactions

73(3):1.1
-
1.8.


Scheatzle, D.G. 2003. Establishing a baseline data set for the evaluation of hybrid (radiant/convective
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Research
Project

RP
-
1140, Final Report.


Scholander, P.E. 1958. Cold adaptation in the Australian aborigines.
Journal of Applied Physiology

13:211
-
218.


Stevens, J.C., L.E. Marks, and A.P. Gagge. 1969. The quantitative assessment of
thermal comfort.
Environmental Research

2:149
-
65.


Suryanarayana, S. and R.H. Howell. 1990. Sizing of radiant heating systems, Part II

Heated floors and infrared units.
ASHRAE Transactions

96(1):666
-
675.


Wilkins, C.K. and R. Kosonen. 1992. Cool ceiling sy
stem: A European air
-
conditioning alternative.
ASHRAE Journal

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-
45.


Xiang, W., S.A. Tassou, and M. Kolokotroni. 2001. Heating of church buildings of historic importance with direct gas
-
fired heating systems.
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364.


Zme
ureanu, R., P.P. Fazio, and F. Haghighat. 1987. Thermal performance of radiant heating panels.
ASHRAE Transactions

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-
27.


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-
1989 Fundamentals. Atlanta: American Society of Heating, Refrigerating, and Air
-
Conditioning Engineers, Inc.



ASHRAE Standard 55
-
1992, “Thermal Environmental Conditions for Human Occupancy”. Atlanta: American Society of Heating,
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-
Conditioning Engineers, Inc.



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2.



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-
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RING THIS CHAPTER OF

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Last saved by
John Gautrey

17

2/22/2014

6.5.2

D
ESIGN
E
XAMPLE
:

O
FFICE

The office has a sensible cooling load of 6.824 BTUH which is obtained from the load calculations.

A 13’ x 16.5’ x 8’ interior office with a 2’ x 4’ T
-
Bar ceiling and a maximum occupancy of 7 people, is to be maintained at
75°F and a 45% RH

(S.P. 2, Figure ____)
.


From the psychrometric chart, this gives a dew point temperature of

52.2

(S.P. 3)

and a moisture
content of
.0083
lb of moisture/lb of dry air
.

Step 1.

Determine the sensible and latent hourly heat gain for the room.

The sensible and latent hourly heat gains are found using accepted procedures found in the ASH
RAE handbooks. For
this example, we have calculated;

Sensible hourly heat gain: 6,824 BTUH

Latent hourly heat gain: 662 BTUH

Step 2.

Determine the mean water temperature required for cooling.

Supply water temperature: = 55°F

Assuming a temperature rise of

5°F, add half of this temperature rise to the inlet water temperature. MWT: 56°F + 8°F/2
= 59°F

Step 3.

Determine the minimum air supply required for the room.

According to ASHRAE tabulated data, the minimum air supply for an office is:

20 cfm per person

x 7 persons =
140

cfm

Step 4.

Determine the latent load capacity of the air with the following equation:




60



spply
room
fg
L
Qph
q



7 occupants at 220 BTU/pp = 1,540 BTU/pp the moisture gain is 1540/4840X
140
=

0.00227
lb/lb. The moisture content
of the
entering

supp
ly air
condition is

0.0083

-

0.00227

=
0.00
60

lb/lb.


Specify

1.

the dry bulb and calculated relative humidity of the entering supply air conditions (
see
S.P. 1) from the
psychrometric chart

a.

55°F
t
db

@ 65.3 RH =
entering conditions of the supply air

2.

the dew point and operating dry bulb surface temperature of the radiant cooling panel.

a.

Select for

2
°F or greater, above the space operating dew point conditions of 5
2
.2°F/5
2
.2°F (
see
S.P. 3)

b.

5
2
.2°F + 3°F = 5
5
.2°F =

lowest operating
surface
temperature of the cooling panel (
see
S.P. 4
-
5)



























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50
55
60
65
70
75
80
DRY BULB TEMPERATURE - °F
.002
.004
.006
.008
.010
.012
.014
.016
.018
10% RELATIVE HUMIDITY
20%
30%
40%
50%
60%
70%
80%
90%
45
50
50
55
55
60
60
65
65
70
70
HUMIDITY RATIO - POUNDS MOISTURE PER POUND DRY AIR
2
3
1
4
5
Example: Office, Legend
S.P. 1 Supply air conditions to space
S.P. 2 Operating air conditions
S.P. 3. Dew point temp. of operating air
S.P. 4-5 Operating and dew point temp. of the cooling panel
R
R
ASHRAE PSYCHROMETRIC CHART NO.1
NORMAL TEMPERATURE
BAROMETRIC PRESSURE: 29.921 INCHES OF MERCURY
Copyright 1992
AMERICAN SOCIETY OF HEATING, REFRIGERATING AND AIR-CONDITIONING ENGINEERS, INC.
SEA LEVEL





(Note from RB


my ASHRAE psy tool (set for sea level) is consis
tently
different

than the numbers used

in this
text....so please check them)


Figure
19

shows a psychrometric chart used to determine dew point of space based on operating temperatures


The dewpoint temperature of the space operat
ing temperature is shown on the psychrometric chart in figure 4. The
operating surface temperature of the ceiling panels is kept about 3°F higher than the space dewpoint temperature.

Step 5.

Determine the sensible load capacity of the air with the followi
ng equation:



60



spply
room
p
s
t
t
QpC
q

60
)
59
75
(
/
24
.
0
/
075
.
0
140
3






F
F
lb
BTU
ft
lb
cfm
q
s



BTUH
q
s
418
,
2


The sensible cooling required from the panels = 6,824


2,418 = 4,405 BTUH

Step 6.

Select a panel surface temperature at least 3°F higher than the operating dew temperature =
52.2
+ 3 = 5
5.2
°F


Step 7.

From the
55.2
F panel operat
ing temperature or panel mean water temperature =
55.2
°F

(panel op surf temp =
t
mean
??

Conflicts with classroom example)
,

We can derive the supply water temperature and water temperatures to and from the panels. Typically the difference
supply and return

water temperature is 8°F so the panel water supply temperature in this case =
55


4 = 5
1
°F and the
panel return water temperature = 5
5

+ 4 =
59
°F

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Step 8.

From the Figures 7, 8 and 9 below, we can see the panel output, the flow rate and the coolin
g output in
comparison with the air temperature


M.W.T.

(need to redo the figure to capture the lower mean temps.)


Figure
20

shows that the paneloutput just below 26 BTUH/ft2 at a room temperature of 75.2F and MWT of 61F.

Step

9.

The required sensible cooling output from the panel is 4,405 BTUH and therefore the required panel area is =
4,405/30 = 146sf. The office has a ceiling area of 13’ x 16.5’ = 215sf. The ratio of ceiling panel to ceiling is 68%.

6.5.3

C
ONDENSATION
C
ONTROL

Now

we have determined the dew point temperature of the space air, the surface temperature of the radiant ceiling can
be controlled to be above the dew point and therefore avoiding the risk of condensation. Monitoring the space air
temperature and the space h
umidity levels will provide the space moisture content. The graph below shows how the
panel surface temperature is increased to remain above the space dew point temperature. In simple terms the supply
water temperature to the panels must be controlled to a
void the possibility of condensation.


Note: ASHRAE Std 62
.1 Ventilation for Acceptable Indoor Air Quality

specifies maintenance of mechanically air
conditioned space to have a relative humidity less than 60%
(The Standard wa
s republished in 2010


is this still a true
statement? I couldn’t find it in my hard copy…doesn’t mean it’s not there


just couldn’t find it

RBean
)
which at sea level
is a dew point of 58F so any sytem controlled in accordance with Std 62 should be able
to operate with chilled water
temperatures of 60F or greater with no danger of condensation occurring.



The only possibility of condensation occurring is when radiant cooled ceilings are used in a space with operable
windows. From
practice

it is
know
n

that operable windows can induce up to 6 ach through a space

(reference)
. For a
particular application in say New York City the maximum summer design conditions are 89°F dry bulb and 73°F wet bulb
and has a moisture content of
0.0138
lb/lb. The dew point temperature of this outside air condition is
66.
1
°F, so if this
condition were to come into contact with a radiant panel with a surface temperature of 59°F the inevitable would happen
and condensation would occur. A very simple contr
ol methodology to avoid condensation is to elevate the panel surface
temperature to say 3°F above the outside air dew point temperature of
66.1
°F which would give a panel surface
temperature of
69.1
°F. From the data shown on figure 3, adjusting the pan
el supply and return temperatures to maintain
a panel surface temperature of
69
°F, would give a
73
°F return temperature. At the same time with 6 ach of outside the
space temperature would also rise to somewhere around
81
-

83
°F which would enable
the panels to provide more than
26 BTUH/sf of cooling to the space. This might be considered pushing the limits of control possibilities but it does show
that radiant cooled ceilings can be operated in high moisture content spaces.


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50
55
60
65
70
75
80
85
90
DRY BULB TEMPERATURE - °F
.002
.004
.006
.008
.010
.012
.014
.016
.018
10% RELATIVE HUMIDITY
20%
30%
40%
50%
60%
70%
80%
90%
45
50
50
55
55
60
60
65
65
70 WET BULB TEMPERATURE - °F
70
75
HUMIDITY RATIO - POUNDS MOISTURE PER POUND DRY AIR
OSA
OSAdp
tp1dp
tp2dp


Op Air
Open Window1
Open Window2
For low mass conductive radiant panels
In the event that it is necessary to raise the panel surface
temperature from tp1dp to tp2dp due to open windows,
(Op Air to OSA) the ensuing differential
temperatures between the rising room
temperature and rising panel temperature still
allows for cooling based on the heat transfer coefficients.
R
R
ASHRAE PSYCHROMETRIC CHART NO.1
NORMAL TEMPERATURE
BAROMETRIC PRESSURE: 29.921 INCHES OF MERCURY
Copyright 1992
AMERICAN SOCIETY OF HEATING, REFRIGERATING AND AIR-CONDITIONING ENGINEERS, INC.
SEA LEVEL


Figure
21

shows a psychrometric chart used to determine dew point based on operating temperatures

The psychometric diagram shows the normal operating condition of a radiant ceiling for cooling. Should the space
moisture content be increased by allowi
ng warm moist air to enter the space and increase the moisture content of the
space air and then the radiant ceiling supply water must be elevated to maintain a panel surface temperature at least
3°F higher than the dew point temperature of the space air.

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7.0

E
MBEDDED
S
YSTEMS

7.1

INTRODUCTION


A radiant system described in this section is a surface
cooling and heating system where water tube as well as
electric cables are integrated within the floor, the wall or
the ceiling. Radiant systems have been successfully
used
in Europe for heating and cooling of buildings. A Radiant
system provides a very good method of discharging high
specific cooling loads while maintaining thermal comfort at
relatively low operation costs. The mainly use of radiant
systems are the radi
ant floor heating, however ceiling as
well as wall heating are used. Since the end of the 80’,
radiant cooling systems were developed in Europe.

R
adiant systems are dimensioned in accordance to the
ensuing radiant heat exchange in the space. Radiant
sys
tems usually designed as a hydronic system therefore
the amount of space necessary for the installation is
considerably smaller than a conventional air conditioning
system. Because of the low plenum height necessary to
accommodate the installations, more a
rchitectural
freedom is provided. The radiant system can be installed
so that both individual and zone control can be achieved.


7.2

B
ACKGROUND

Many hospitals have radiant systems, but commercial
buildings have been slow to utilize the potential of radiant
f
loors and ceilings. In Europe several buildings such as
the PGEM in Arnhem and the Groninger Museum have
successfully used radiant systems to control the indoor
environment. There has also been an advancement in
simulation programs which permit a more de
tailed
analysis of the indoor environment. With these advanced
simulation tools the individual elements necessary for the
creation of a comfortable indoor climate using the
predicted mean vote as determined by Fanger (1972),
radiant heat exchange can be s
tudied. Because each
individual surface temperature and its relationship (i.e.
position to the other surfaces) can be determined a
solution to the comfort balance equation can easily be
found.

When incorporating a radiant system and a constant
volume vent
ilating system, the ventilation system may only
be dimensioned to supply of outdoor air for each person
and to remove the latent as well as the material load if the
radiant system is selected to remove the remaining
cooling loads. Simmonds (1994) has repor
ted on some of
these designs and how effective they are in providing an
effective means of comfort climate.

Since cooling surfaces make no contribution to air renewal they may always be operated in conjunction with a ventilating
or air
-
conditioning plant,
which also ensures the probably necessary dehumidification. However, also a combination with
a natural ventilation system as well as operable windows may also be possible.

Radiant cooled floors have been successfully used on many projects over the past ten

years and many ASHRAE and
other peer reviewed papers have been written on these projects. However, many questions still arise regarding the
performance of radiat cooled floors when subjected to solar radiation. The convective cooling performance of a cool
ed
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floor has been reported by Olesen, Simmonds; Borressen (1994) reported on the solar absorption of a radiant cooled
floor, this was further reported by Simmonds, Gaw, Holst and Reuss (1996). Even with this research many questions are
still being raised r
egarding the performance of a radiant cooled floor when absorbing solar radiation. This paper explains
the performance of the floor including some limitations of the presented calculation method, but also includes
information on the controllability.

This c
hapter will outline the performance of a radiant cooled floor subjected to different solar radiation intensities and will
look further into the controllability of the radiant cooled floor to maintain certain conditions such as space temperatures
and floor
surface temperatures. The influence of floor coverings on radiant cooled floors will also be discussed.

Radiant systems are dimensioned in accordance with the ensuing radiant heat exchange in the space. Radiant systems
are usually designed as a hydronic
system therefore the amount of space necessary for the installation is considerably
less than a conventional air conditioning system. Because of the low plenum height necessary to accommodate the
installations, more architectural freedom is provided. The r
adiant system can be designed to provide both individual and
zone control.

Previous papers by Simmonds (1994) have shown that radiant cooled floors are capable of removing between 35
-
40
W/m2 (11
-
12.5 btuh/sf) from spaces. Borressen (1994) and Simmonds, Gaw
, Holst and Reuss (1996) have shown that
radiant cooled floors are capable of removing up to 85 W/m2 (27.0 btuh/sf) of energy from a space 35 W/m2 (11.1
btuh/sf) by convection and 50 W/m2 (16 btuh/sf) by solar absorption. This paper uses a simple steady st
ate equation to
explain the performance of a radiant cooled floor when performing at its maximum capacity of both reducing the space
air temperature and absorbing solar radiation that is reaching the floor. There are several dynamic simulation programs
tha
t accomplish the dynamic performance of a radiant floor. But a steady state model is presented in this paper to
simplify the explanation. There are future plans to write a paper on dynamic simulation of a radiant cooled floor. Many
papers have been writt
en on the performance of radiant floor for heating MacCluer, Athienitis, Simmonds, Olesen,
Meirhans have reported on the performance of active concrete systems.

There has been advancement in simulation programs that have permitted a more detailed analysis
of the indoor
environment. With these advanced simulation tools the individual elements necessary for the creation of a comfortable
indoor climate using the predicted mean vote (PMV) as determined by Fanger (1972), and radiant heat exchange can be
studied
. Because each individual surface temperature and its relationship (i.e. position to the other surfaces) can be
determined a solution to the comfort equation can be found. The PMV/PPD comfort equation derived by Fanger and
shown in ASHRAE Standard 55
-
2004
, can be influenced by the control or balance of the radiant heat exchange in a
space. When operating a radiant cooled floor correctly the surface temperature of the floor can be regulated. Absorbing
a major portion of the solar radiation entering the spa
ce also prevents the floor from emitting thermal energy, in the form
of long wave radiation, back into the space and onto other surfaces.

When incorporating a radiant system and a constant volume ventilating system, the ventilation system need only be