HPAC Magazine

Why And How To Do Radiant Cooling

February 2, 2014 | By Robert Bean

"Have you ever considered that most naysayers of hybrid radiant-based HVAC systems say the problem is that you need two systems, one for comfort and one for ventilation; and yet many ventilation experts agree that independent ventilation systems are preferred - go figure."

When it comes to radiant there is no shortage of myths, we address at least 45 of them at www.healthyheating.com. 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. Translation: it facilitates the use of 100 per cent dedicated, ducted and distributed outdoor air (DOAS). This system has many advantages in that its sole existence is for the exclusive tasks of dehumidification, deodorization and decontamination.

In comparison to all-air systems, the air part of a hybrid is designed and assembled around significantly reduced air flows leading to smaller 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 regulating the environmental conditions necessary for controlling microbial population, 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 55F to 70F (13C to 21C) for high performance buildings using masonry type flooring. This makes 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 60F to 75F (16C to 24C) also maximizes efficiency from cooling plants and reduces transmission gains.

Radiant systems also serve the needs of architects and interior designers through greater freedoms with space, the ability to use low VOC materials and superior capacity in handling direct solar load with a quieter and more pleasant solution.


As much as industry might want there to be a Radiant Cooling for Dummies – there is not. In fact, I would be disturbed to think that such a book would be published. We have enough stupid radiant heating tricks out there. We do not need to pile on radiant cooling fiascos. The designer and contractor must understand the interactions and connections between buildings, the indoor environment and HVAC systems and controls. It is not difficult but it does require skill sets beyond the typical hydronics only or air only technician. It requires a hybrid radiant-based HVAC designer and contractor.

Abridged how it is done

First, understand that 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. The dry supply air will act as a sponge to maintain space operating conditions below the dew point of the radiant panel. Panel surface and radiant asymmetry temperature limits will be 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 (qs) calculated to be 28 584 Btuh (8.4kW) and space conditions maintained at 74F (23.3C) operative temperature (top ) and 50 per cent relative humidity (rh). From the psychrometric chart, this gives a dew point temperature of ≈ 54F (12.2C) and a moisture content of ≈ 0.00896 lbH2O/lbdry .

1. Using a 100 per cent outdoor air supply per person of 20 cfmiii, ventilation flow rate (Qv) becomes (all calcs in IP units):

Qv=30 persons•20 cfm pp = 600 cfm [1]

2. The latent load (qL) due to occupants is calculated, using an estimate of 200 to 220 Btuh/per person (approximation from ASHRAE activity tables ) as:

qL=30 occupants x 220 Btuh/pp = 6600 Btuh [2]

3. The humidity ratio differential (Δω) due to ventilation is calculated for the occupant latent load as:

qL=latent heat of vaporization (Lv )•air flow rate (Q)•Δ in humidity ratio (Δω) [3]

qL=(60 min/h•1076 Btu/lb water•0.075 lb air/ft3 )•600 cfm•Δω

qL=4840 Btu-min/ft3-hr•600 cfm•(Δω)

4. From formula [2] the occupant latent load, qL = 6600 Btuh, and now Δω can be calculated using [3];

6600Btuh=4840 Btu-min/ft3-hr•600 cfm•(Δω),

Rewritten to solve for Δω;

Δω = qL/Lv•Q [4]

Δω = 6600 Btuh/(4840•600 cfm)

Δω ≈ 0.00227 lbH2O/lbdry

0.00227 lbH2O/lbdry represents the humidity ratio differential for 30 people.

5. For determining dew point, calculate the maximum anticipated humidity ratio (ωoccupied) starting with a space operating condition of 74F (23.3C) @ 50 per cent rh adding a 30 person load (infiltration and other possible latent loads ignored for this example):

ωoccupied = ωoperating + ωpeople …+ ωother [5]

ωoccupied = ω(74°F@50%rh) + ω(30 people @ 220 Btu/pp)

ωoccupied = 0.00896 lbH2O/lbdry + 0.00227 lbH2O/lbdry

ωoccupied = 0.01123 lbH2O/lbdry

From the psychrometric chart (Figure 1),

ωoccupied = 74F @ 62.5 per cent rh equals a dew point of 60.4F

6. Establish the minimum allowable surface temperature of the radiant panel (tp); based on good practice, select for ≈ 2F to 3F (1C to 1.5C ) minimum Δt above the dew point :

tp = ωoccupied-dp + 3F [6]

tp = 60.4F + 3F = 63.4F (17.5C)

Notwithstanding radiant asymmetry and comfort, 63.4F (17.5C) represents the lowest allowed panel surface temperature with a sufficient safety margin to prevent surface condensation.

7. Calculate sensible supply air capacity (qs): with an operating space dry bulb of 74F (23C) and designer choice supply dry bulb of 55Fv , the sensible air capacity becomes:

qs=60 min/h•(specific heat, Cp)•(density, ρ)•(air flow rate, Q)•Δt [7]

qs=(60 min/h•0.244 Btu/lbF•0.075 lb/ft3 )•cfm•Δt

qs = 1.08•cfm•Δt

qs = 1.08•600•(74F – 55°F)

qs = 12 517 Btuh

12 517 Btuh represent the sensible air cooling capacity of the supply air. This value, deducted from the 28 584 Btuh total sensible required, is what the radiant cooling panel must absorb.

8. The sensible cooling load placed on the radiant panel becomes:

qs,panels  = Total load (sensible) – air cooled (sensible) [8]

qs,panels  = 28 584 Btuh – 12 517 Btuh

qs,panels  = 16 069 Btuh

This 16 069 Btuh 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:

qflux = qs, pane
l/Aavailable panel area   [9]

qflux = 16 069 Btuh/(30 ft x 30 ft)

qflux = 17.85 Btuh/ft2

This 17.85 Btuh/ft2 can be absorbed from any type of radiant panel.

10a. For radiant ceiling cooling, the surface temperature (ts) becomes:

ts=top – (qflux/heat transfer coefficientviii) [10a]

ts = 74F – (17.85 Btuh/ft2 /1.94 Btuh/ft °F)

ts=64.8F > 63.4F (safety margin temperature) > 60.4F (occupied dew point) = good

Since 64.8F (18.2C) is above the safety margin limit of 63.4F (17.4C) and more than the 60.4F (15.8C) occupied dew point this would be an acceptable solution.

10b. For radiant floor cooling, the surface temperature (ts ) becomes:

ts = top – (qflux / heat transfer coefficientv) [10b]

ts = 74F – (17.85 Btuh/ft2/1.23 Btuh/ft°F)

ts = 59.5F (15.3C)

ts=59.5F < 60.4F (occupied dew point) < 63.4F (safe) & < 66F min.

Since 59.5F (17.5C) is below the occupied dew point of 60.4F (15.8C) and below the acceptable 66F (19C) surface temperature for thermal comfort, this would be an unacceptable solution without modifications to rework design for; higher operative temperature (top) 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 (ts) becomes:

ts = top – (qflux/heat transfer coefficientv) [10c]

ts = 74F – (17.85 Btuh/ft2/1.41 Btuh/ft °F)

ts = 61.3F

ts= 61.3F > 60.4F (occupied dew point) < 63.4F (safety margin temperature)

Since 61.3F (16.3C) is above the occupied dew point of 60.4F (15.8C) but below the acceptable safety margin surface temperature of 63.4F (17.4C) this is a riskier application and would be an unacceptable solution without modifications to rework the design for; higher operative temperature (top) 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 (qs), latent (ql) and total load (qt) for cooling and dehumidification load for the dedicated outdoor air system to take 600 cfm of 100 per cent outdoor air from an example of 85F (29C) @ 80 per cent rh (h1, ω1) to a supply air of 55F (13C) @ 50 per cent rh (h2, ω2).

State point conditions from the psycrometric chart:

h1 = 43.44 Btu/lbdry

ω1 = 0.0210 lbH2O/lbdry

h2 =  18.17 Btu/lbdry

ω2= 0.0046 lbH2O/lbdry

h3 =  25.45 Btu/lbdry

ρ  = 0.0750 lb/ft3


hn = enthalpy at state point

ωn = humidity ratio at state point

ρ    = air density

12. Sensible (qs), latent (ql) and total load (qt) on coil is calculated using:

q= 60 min/h•density (ρ)• air flow rate (Q)•enthalpy differential (Δh) [12]

q = (60 min/h•0.075 lb/ft3)•cfm•Δh

q = 4.5•cfm•Δh

qs= 4.5•600•(h3 – h2) = 2700•(25.45 – 18.17) = 19 656 Btuh

ql= 4.5•600•(h1 – h3) = 2700•(43.44 – 25.45) = 48 573 Btuh

qt= 4.5•600•(h1 – h2) = 2700•(43.44 – 18.17) = 68 229 Btuh

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 re-heat or heat recovery potentials of the system. A competent HVAC engineer or technician would be able to describe all the necessary processes for each application and choice in DOAS equipment.

It is 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 realized by radiant-based HVAC systems.ix,x

Robert Bean, R.E.T., P.L.(Eng.) is president of Indoor Climate Consultants Inc. and a director of www.healthyheating.com. He serves on ASHRAE Committees: T.C.61. (CM), T.C.6.5 (VM), T.C. 7.04 (VM), SSPC 55 (VM). www.healthyheating.com

i Radiant Mythology: myths about low temperature radiant heating and high temperature radiant cooling <http://www.healthyheating.com/Radiant_Mythology/Radiant_Floor_Heating_Myths_.htm#.UrHdYeLDseI>

iiSensible loads: ASHRAE 55, ISO 7730, CSA F280 and Ventilation loads: ASHRAE 62.1, 62.2 and CSA 326

iiiThe 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/CSA-F326-M91 (R2010)

iv2009 ASHRAE Fundamentals Handbook, Chpt.18, Section 4

vThis is purely based on experience but you can see how changing the target number will have a “flow through” effect on the design.

viThe 28 584 Btuh total load came from the heat gain calculation (not described in this example).

viiBean, 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

x  Hydronic Cooling, idronics Journal of Design Innovation for Hydronic Professionals, Caleffi, July 2013 



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