"Have you ever considered that most naysayer of
the hybrid radiant based HVAC systems say the problem is you
need two systems, one for comfort and one for ventilation; and
yet many competent ventilation experts agree that independent
ventilation systems are preferred – go figure."
Introduction
When it comes to radiant there are no shortages
of
myths, we address at least 45 of them at the website[i];
one undisputable whole truth about radiant is its role as an
“enabler”. Specifically when we talk about the hybrid radiant
cooling system  it
enables the preferred separation of
thermal comfort from ventilation.[ii]
Translation: it facilitates the use of
100% dedicated, ducted
and distributed outdoor air. This system has many advantages in
that its sole existence is for the exclusive tasks of
dehumidification, deodorization and decontamination, i.e. there
is no conflict in purpose. In comparison to all air systems, the
air part of a hybrid is designed and assembled around
significantly reduced air flows leading to smaller sized air
handlers, filters, ducts, dampers and fabrication and
installation accessories; all of the above translates to a more
effective system for less capital cost and lower operating and
maintenance costs. Additionally these dedicated duty systems are
very effective at controlling the environmental conditions
necessary for controlling microbial, hydrolysis, swelling in
hygroscopic materials and in promoting respiratory and thermal
comfort.
From an
energy and exergy efficiency perspective,
the sensible part of the hybrid radiant cooling systems is
associated with tepid fluid temperatures in the range of 55°F to
70°F (13°C to 21°C) with high performance buildings using
masonry type flooring; making them ideal for direct ground
coupled exchangers, evaporative cooling with or without night
sky radiation and promote the possibility of compressorless
cooling systems; or at the very least the ability to bypass the
compressor for all but peak loads. The high return temperature
range of 60°F to 75°F (16°C to 24°C) also enables maximizes
efficiency from cooling plants and reduces transmission gains.
Radiant system in general also serve the needs of
the architect and interior designer with greater freedoms with
space, the ability to comfortably use low VOC materials, as well
as superior capacity in handling direct solar load with a
quieter and more pleasant solution.
Design Considerations
As much as industry might want there to be a,
“Radiant Cooling for Dummies”  there isn’t – in fact I’d be
disturbed to think that such a book would be published and end
up…well…in the hands of a dummy. That’s a nice way of saying we
have enough stupid radiant heating tricks out there we don’t
need to pile on with radiant cooling fiascos. This is one type
of system that requires the designer and contractors to
understand the interactions and connections between buildings,
the indoor environment and various HVAC systems and controls.
It’s not difficult but it does require skills sets beyond the
typical hydronics only or air only technician. It…well…yes…
requires a hybrid radiant based HVAC designer and contractor.
So here is an abridged version of how it’s done.
First understand the objective in the hybrid
design is to introduce dedicated lean ventilation air to the
space reflecting the anticipated latent loads from occupants,
infiltration and other sources. You will be using the dry supply
air to maintain space operating conditions
below the dew point of the radiant panel; and have panel surface
and radiant asymmetry temperature limits within the range
established by ANSI/ASHRAE 55  Thermal Environmental Conditions
for Human Occupancy.
Let’s look at a simplified example of a small
30' x 30’ x 10'
classroom with a maximum occupancy of 30 people and a space
sensible cooling load (q_{s }) calculated to be
28,584 Btu/hr (8.4kW) and space conditions maintained at 74°F
(23.3°C) operative temperature (t_{op }) and 50%
rh. From the psychrometric chart, this gives a dew point
temperature of ≈ 54°F (12.2°C) and a moisture content of ≈
0.00896 lb_{H2O}/lb_{dry }.
1) Using a 100% outdoor air
supply per person of 20 cfm[iii],
the ventilation flow rate (Q_{v}) becomes (all
calcs in IP units):
Q_{v
}= 30
persons · 20 cfm per person = 600
cfm
[1]
2) The latent load (q_{L})
due to occupants is calculated, using an estimate of 200 to 220
Btu/hr/per person (approximation from ASHRAE activity tables[iv])
as;
q_{L}
= 30 occupants x 220 Btu/hr/pp = 6,600 Btu/hr [2]
3) The humidity ratio
differential (Dω)
due to ventilation is calculated for the occupant latent load
as;
q_{L}
= latent heat of vaporization (L_{v} ) · air flow
rate (Q) · D
in humidity ratio (Dω) [3]
q_{L}
= (60 min/h · 1076 Btu/lb water · 0.075 lb air/ft^{3} )
· 600 cfm · Dω
q_{L}
= 4840 Btumin/ft^{3}hr ·
600 cfm · (Dω)
4) From formula [2] the
occupant latent load, q_{L }= 6,600 Btu/hr, and
now Dω can
be calculated using equation [3];
6,600 Btu/hr = 4840 Btumin/ft^{3}hr
·
600 cfm · (Dω),
Rewritten to solve for
Dω;
Dω
= q_{L }/ L_{v} · Q
[4]
Dω
= 6,600 Btu/hr / (4,840 ·
600 cfm)
Dω
≈ 0.00227 lb_{H2O}/lb_{dry}
0.00227 lb_{H2O}/lb_{dry
}represents the humidity ratio differential for 30 people.
5) For determining dew point,
calculate the maximum anticipated humidity ratio (ω_{occupied}
) starting with a space operating condition of 74°F (23.3°C) @
50%rh adding a 30 person load (infiltration and other possible
latent loads ignored for this example):
ω_{occupied}
= ω_{operating}
+ ω_{people}
…+ ω_{other
}
[5]
ω_{occupied}
= ω_{(74°F@50%rh)}
+ ω_{(30
people @ 220 Btu/pp)
}
ω_{occupied}
= 0.00896 lb_{H2O}/lb_{dry }+ 0.00227 lb_{H2O}/lb_{dry}
ω_{occupied}
= 0.01123 lb_{H2O}/lb_{dry}
From the psychrometric chart
(Figure 1),
ω_{occupied}
= 74°F @ 62.5%rh equals a dew point of 60.4°F
Figure 1. Lean supply air at 55°F @ 50% rh
delivered to space to control moisture conditions below dew
point of radiant panels. Shown is minimum floor surface
temperature of 66°F based on ANSI/ASHRAE 55.
6) Establish the minimum
allowable surface temperature of the radiant panel (t_{p});
based on good practice, select for ≈ 2°F to 3°F (1°C to 1.5°C )
minimum Dt
above
the dew point :
t_{p
}= ω_{occupieddp}
+ 3°F
[6]
t_{p }= 60.4°F + 3°F = 63.4°F (17.5°C)
Notwithstanding radiant
asymmetry and comfort, 63.4°F (17.5°C) represents the lowest
allowed panel surface temperature with a sufficient safety
margin to prevent surface condensation.
7) Calculate sensible supply
air capacity (q_{s}): with an operating space dry
bulb of 74°F (23°C) and designer choice supply dry bulb of 55°F[v],
the sensible air capacity becomes:
q_{s}
= 60 min/h · (specific heat, C_{p}) · (density,
ρ) ·
(air flow rate, Q) · Dt
[7]
q_{s}
= (60 min/h · 0.244 Btu/lb°F · 0.075 lb/ft^{3} ) · cfm
· Dt
q_{s}
= 1.08 · cfm · Dt
q_{s}
= 1.08 · 600 · (74°F  55°F)
q_{s}
= 12,517 Btu/hr
12,517 Btu/hr represent the
sensible air cooling capacity of the supply air; this value
deducted from the 28,584 Btu/hr total sensible required, is what
the radiant cooling panel must absorb.[vi]
8) The sensible cooling load
placed on the radiant panel becomes:
q_{s,
panels } =
Total load (sensible) – air cooled
(sensible) [8]
q_{s
, panels }
= 28,584 Btu/hr  12,517 Btu/hr
q_{s
, panels }
= 16,069 Btu/hr
This 16,069 Btu/hr can be
assigned to a radiant ceiling, wall or floor or combination of
cooling panels if necessary.[vii]
9) The required radiant panel
surface flux (heat absorption) becomes:
q_{flux
}= q_{s,
}_{panel} / A_{available panel
area}
[9]
q_{flux
}= 16,069
Btu/hr / (30 ft x 30 ft)
q_{flux
}= 17.85
Btu/hr/ft^{2}
This 17.85 Btu/hr/ft^{2}
can be absorbed from any type of radiant panel.
10a) For radiant ceiling
cooling, the surface temperature (t_{s} )
becomes:
t_{s}
= t_{op} – (q_{flux }/ heat
transfer coefficient[viii])
[10a]
t_{s}
= 74°F – (17.85 Btu/hr/ft^{2}_{ }/ 1.94 Btu/hr
ft^{2} °F)
t_{s}
= 64.8°F > 63.4°F (safety margin temperature) > 60.4°F (occupied
dew point) = good
Since 64.8°F (18.2°C) is
above the safety margin limit of 63.4°F (17.4°C) and more than
the 60.4°F (15.8°C) occupied dew point this would be an
acceptable solution.
10b) For radiant floor
cooling, the surface temperature (t_{s} )
becomes:
t_{s}
= t_{op} – (q_{flux }/ heat
transfer coefficient^{v})
[10b]
t_{s}
= 74°F – (17.85 Btu/hr/ft^{2}_{ }/ 1.23 Btu/hr
ft2 °F)
t_{s}
= 59.5°F (15.3°C)
t_{s}
= 59.5°F < 60.4°F (occupied dew point) < 63.4°F (safe) & < 66°F
min.
Since 59.5°F (17.5°C) is below the occupied dew
point of 60.4°F (15.8°C) and below the acceptable 66°F (19°C)
surface temperature for thermal comfort, this would be an
unacceptable solution without modifications to rework design
for; higher operative temperature (t_{op}) or increase
panel surface area (A), add peak cooling panels or second stage
cooling coils or improve zone enclosure to get sensible loads
down.
10c) For radiant wall
cooling, the surface temperature (t_{s} )
becomes:
t_{s}
= t_{op} – (q_{flux }/ heat
transfer coefficient^{v})
[10c]
t_{s}
= 74°F – (17.85 Btu/hr/ft^{2}_{ }/ 1.41 Btu/hr
ft2 °F)
t_{s}
= 61.3°F
t_{s}
= 61.3°F > 60.4°F (occupied dew point) < 63.4°F (safety margin
temperature)
Since 61.3°F (16.3°C) is above
the occupied dew point of 60.4°F (15.8°C) but below the
acceptable safety margin surface temperature of 63.4°F (17.4°C)
this is a riskier application and would be an unacceptable
solution without modifications to rework the design for; higher
operative temperature (t_{op}) or increase panel surface
area (A), add peak cooling panels or second stage cooling coils
or improve zone enclosure to get sensible loads down.
Now that we have determined
the sensible load on the panel system and sensible and latent
load on the air system, we need to determine the capacity of the
cooling coil in the air handler.
11) Calculate
sensible (q_{s}),
latent (q_{l})
and total load (q_{t})
for cooling and dehumidification load for the
dedicated outdoor air system to take 600 cfm of 100% outdoor air
from an example of 85°F @ 80%rh (h_{1},
ω_{1})
to a supply air of 55°F @ 50%rh (h_{2},
ω_{2});
State point conditions from the
psycrometric chart:
h_{1}
= 43.44 Btu/lb_{dry}
ω_{1
}= 0.0210
lb_{H2O}/lb_{dry}
h_{2}
= 18.17 Btu/lb_{dry}
ω_{2}=
0.0046 lb_{H2O}/lb_{dry}
h_{3}
= 25.45 Btu/lb_{dry}
ρ
=
0.0750 lb/ft^{3}
Where,
h_{n}
= enthapy at state point
ω_{n}
= humidity ratio at state point
ρ
= air density
12) Sensible (q_{s}),
latent (q_{l})
and total load (q_{t})
on coil is calculated using;
q_{
} = 60 min/h · density (ρ)·
air flow rate (Q) · enthalpy differential (Dh)
[12]
q_{
} = (60 min/h · 0.075 lb/ft^{3}
) · cfm · Dh
q_{
}= 4.5 · cfm ·
Dh
q_{s}=
4.5 · 600 · (h_{3 }– h_{2}) =
2700 · (25.45 – 18.17) = 19,656 Btu/hr
q_{l}
= 4.5 · 600 · (h_{1 }– h_{3}) =
2700 · (43.44 – 25.45) = 48,573 Btu/hr
q_{t}
= 4.5 · 600 · (h_{1 }– h_{2}) =
2700 · (43.44 – 18.17) = 68,229 Btu/hr
Figure 2. Primary load on the DOAS coil for
cooling and dehumidification. Note: this is a simplified
overview and does not represent all possible loads including
heat recovery.
Figure 3. Overview of the operating
characteristics for the hybrid radiant based HVAC system in the
cooling mode.
In a nut shell, that is the why and how process
for doing a radiant cooling system; albeit a simplified
description since it does not describe reheat nor heat recovery
potentials of the system but a competent HVAC engineer or
technician would be able to describe all the necessary processes
for each application and choice in DOAS equipment.
It’s sufficient to say that radiant cooling is
becoming a big thing especially for commercial buildings. There
is no need to pay attention to the myths nor is there any need
to experiment. The applications and calculation procedures are
proven and the working projects all over the world are
demonstrating the energy and comfort features and benefits
enabled by radiant based HVAC systems.[ix]^{,[x]}
[ii]
Sensible loads: ASHRAE 55, ISO 7730, CSA F280 and
Ventilation loads: ASHRAE 62.1, 62.2 and CSA 326
[iii]
The actual cfm/person is determined by a method
acceptable to the authority having jurisdiction, sample
procedures are defined in ANSI/ASHRAE Standard 62.1 ,
62.2 and CAN/CSAF326M91 (R2010)
[iv]
2009 ASHRAE Fundamentals Handbook, Chpt.18, Section 4
[v]
This is purely based on experience but you can see how
changing the target number will have a “flow through”
effect on the design.
[vi]
The 28,584 Btu/hr total load came from the heat gain
calculation (not described in this example).
[vii]
Bean, R. Together Forever (Using the ASHRAE Radiant
Design Nomograph), HPAC Canada, March 2012.
[viii] *Heat transfer
coefficients (htc)  are empirical values determined
from experiments (ref.: ASHRAE, REHVA and ISSO)
[ix]
Radiant Cooling Design Manual, Embedded Systems for
Commercial Applications, Uponor, 2013
