Part II: How to use the ASHRAE Design Graph for Sensible Heating and Cooling with Floor and Ceiling Panels
Copyright © 2012, Robert Bean, R.E.T., P.L.(Eng.). All rights reserved. Edited and originally published in HPAC Canada, March, 2012 under a 2 part article called: "Together Forever"

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See Part I: Using the National Research Council of Canada's IA-QUEST VOC Emission and Simulation Tool

For a background discussion see: thermal and optical properties of flooring and other interior finishes.

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Introduction

Part II of my presentation explaining why HVAC designers cannot operate in isolation from interior designers (see Part I) will use Figure 1, the ASHRAE Design Graph for Sensible Heating and Cooling with Floor and Ceiling Panels.¹

We suggest you open the thumbnails in a new window so you can view it along with the text.

Figure 1, the ASHRAE Design Graph for Sensible Heating and Cooling with Floor and Ceiling Panels¹ reprinted with permission from Section/page 6.8, Panel Heating and Cooling, 2012 ASHRAE Systems and Equipment Handbook. Floor and ceiling finishes are an integral part of radiant design.

Some of you might look at this and think that the psychrometric chart looks like child's play next to this graph; but like the psychrometric chart, they are both simple to use once you know how to navigate around the various components.

So let me start out with the geography of the graph. There are essentially six segments, of which the designer needs to establish three through calculation; that being the flux (q, Figure 2), average unconditioned surface temperatures (AUST, Figure 3) and characteristic panel resistances (r_u, Figure 5); the first two (q and AUST) drive the temperature difference between the floor and space (|t_p-t_a|, Figure 4) to emit or absorb the flux at design conditions; the latter (r_u) along with flux (q) and the selection in tube spacing (M, Figure 6) drives the average design fluid temperature in the pipes (t_w, Figure 7).

Your choice in fluid delta t (Δt) in the flow calculation establishes the return temperature (t_return= t_w - (Δt/2)) and thus the efficiency of the boiler, heat pump, solar system or chiller; and it establishes the upper end of the reset curve for programming your fluid controller (t_supply = t_w + (Δt/2)). Obviously all of the factors above are not trivial, as all affect the capital and operating cost of the system, the quality of the indoor environment, efficacy of the surface temperature and integrity of the panel materials, which is (again) why radiant should never be treated as a science experiment. Let’s look at each of the steps in using the design graph.

Step 1. Heating or Cooling Flux (q), Btuh•ft²(W/m²) (Figure 2)

Flux in heating is the energy released from the floor through radiation and convection per unit area. In cooling, it is the energy absorbed from convection and radiation. At the end of the day the performance of a radiant-conditioned building is judged by these values. High numbers represent bad buildings, while low numbers represent good buildings. I have categorized flux into four regions, as shown in Table 1.

Table 1. Enclosures level performance: based on combined convective and radiative flux per unit of floor area (sensible loads only in the long wave range).
Enclosure performance level	Thermal flux, Btu/ft2·hr, (W/m2)
Terrific (high)	< 10 (31)
Transitional (good)	10 to 20 (31 to 63)
Traditional (moderate)	20 to 30 (31 to 94)
Terrible (poor)	> 30 (94)
I came up with these parameters by tying surface flux back to boiler efficiencies based on potential return temperatures using conductive floors and tight tube spacing in radiant floor heating systems. I’m not saying it should be used universally but hey it’s as good as any definition that I have seen elsewhere

The short strokes are the higher the flux the hotter the temperatures in heating and the colder the temperatures in cooling. Hot temperatures in heating and cold temperatures in cooling destroy plant efficiency.

If you have to operate an embedded pipe radiant system at temperatures typical of baseboards, fan/coils, panel radiators and free standing radiators, you have missed the entire point of using the floor, walls and ceilings for conditioning people and spaces. Low temperatures in heating and high temperatures in cooling are the end game for sustainability – full stop.

Step 2. Average Unconditioned Surface Temperatures. (AUST), F(C) (Figure 3)

Any surface that is cooler than the heating panel will draw energy out of that heating panel. Likewise, a cooling panel will draw energy out of any warmer surface. All of the unconditioned surfaces combined (those without pipes), represent the radiant cooling or heating load on the conditioned panel. These unconditioned surfaces along with the conditioned surface represent the mean radiant temperature (MRT) calculated as an area weighted average. The value of the MRT averaged out with the dry bulb temperature (t_db), represents the operative temperature (t_op = (MRT + t_db)/2) as described in ASHRAE Standard 55 – Thermal Environmental Conditions for Human Occupancy. The calculation of AUST and MRT is beyond the scope of this article, but for terrific buildings (0-10 Btuh•ft² (0-31 W/m²)) the AUST can in some cases be assumed to be at or near space temperatures.

Step 3. Positive Difference Between Effective Panel Surface Temperature and Dry-Bulb Room Air Temperature |t_p-t_a|, F (C) (Figure 4)

The greater the flux, the greater the temperature differential required between the panel and space temperature. Surface temperatures are regulated by ASHRAE Standard 55 Thermal Environmental Conditions for Human Occupancy based on health and comfort research. In heating 85°F(29°C) and cooling 66°F(19°C) are the accepted boundaries for occupants wearing normal footwear. For bare or stocking feet, refer to the 2009 ASHRAE Fundamental Handbook, Section 9.15, Warm or Cold Floors.

Step 4. Characteristic Panel Resistance (r_u), ft²•h•F/Btu (m²•K/W) (Figure 5)

This is an intimate element of interior design due to finishes, which affect the efficiency of the heating and cooling plant; and is where the meat and potatoes of radiant design take place. It is also the one that is taken for granted especially by the DIY crowd. Radiant panels are on site fabricated heat exchangers and ultimately the operating temperatures of the fluid are a function of what is known as the fin efficiency of the panel; which is a function of the resistance of: the flooring layers; conductivities of the pipe and encasing materials; pipe surface area; fluid flow characteristics; and log mean temperature differences in the exchanger design. You may have seen some of the finite element analysis (FEA) John Siegenthaler and I have presented on radiant systems. FEA is an excellent way of modelling the heat exchanger to evaluate its efficiency and effectiveness. The calculation procedure again is beyond the scope of this article but the procedures are in the 2008 ASHRAE HVAC Systems and Equipment Handbook, Section 6.2.

Step 5. Spacing (M), in(cm) (Figure 6)

As noted above, spacing plays a significant role in the effectiveness and efficiency of the panel. It also has a major affect on the surface efficacy, that being the quality of the surface temperature of the panel. To this day it astounds me that individuals try to save money on jobs by reducing the amount of pipe when it has a major affect on operating costs and quality of the system – this despite PEX pipe being the least costly component in a radiant system. When it comes to heat exchanger designs, pipe is good and the more the better. Since servicing pipe after it is installed is not on anyone's “bucket list,” put in the highest quality pipe you can find. Lots of good quality pipe equals low temperatures in heating and high temperatures in cooling, which adds up to the highest efficiency and good comfort.

Step 6. Average Fluid Temperature (tw), F (C) (Figure 7)

Average Fluid Temperature is the “keep your eye on the fry” component. Remember it is the average – meaning it is not the supply or return, it is the in-between temperature. That is why they call it the average – go figure. The lower the average temperature in heating and higher in cooling the better. How does this happen? It happens with good to high performance buildings with conductive floors and lots of pipe. It is not any more complicated than that. Here are two examples:

Example, Part 1 Radiant Floor Heating (Figure 8)

1. From the flux side of the graph, q = (25 Btuh•ft²)(79 W/m²), draw a horizontal line straight across.

2. Deduct the desired air space temperature (t_a=70°F) from, in this case a hypothetically calculated AUST of 68°F for a result of (-2°F) (1°C), and draw a line straight down to where it intersects the horizontal flux line established in Step 1.

3. Running parallel with the sloped lines for heating, take the line down from the intersection to the boundary and read the differential (approximately 13°F)(7°C). For a space temp of 70°F(21°C), the floor temperature becomes 70°F+13°F = 83°F (28°C), which is within ASHRAE Standard 55. For those familiar with heat transfer coefficients, you would get similar floor temperature results with:
[(25 Btuh•ft²)/(2 Btuh•ft•°F)]+70°F = 82.5°F (28°C).

4. Now select a pipe spacing (M) and again this is not the place to be cheap – more of the highest quality is better! In this case we have selected 9"o.c.(23cm)

5. From the intersection of (in this example) a hypothetical calculated Characteristic Panel Resistance (r_u) of 0.2 ft²•h•F/Btu (0.03 m²•K/W), representing conductive low VOC flooring, we take a line down to the 9"(23cm) spacing line.

6. At the above intersection, follow the sloped lines to the boundary and read the value of 25°F(14°C).

7. Add this value 25°F(14°C) to the space temperature of 70°F(21°C) for an average fluid temperature (t_w) of 95°F(35°C).

8. For comparisons do the same with a Characteristic Panel Resistance of
2.5 ft²•h•F/Btu (0.5 m²•K/W) representing a less conductive floor and likely having higher VOC emissions.

Message: The difference between low VOC conductive flooring and less conductive higher VOC flooring is, in this case, the differences between average temperatures of 95°F(35°C) and 150°F(66°C) or a 55°F(31°C) spread. Now think about this…with the low VOC conductive flooring and a 20°F (11°C) delta t, the return temperature becomes 85°F(29°C); which puts you into the high 90s in boiler efficiency and above a COP 4 for a heat pump – whereas the more resistive flooring puts you out of the heat pump league and drops your boiler efficiency into the mid range regardless of its high efficiency capabilities. All of this is based on 9"(23cm) o.c. which means it could get worse if you decided to go cheap and use wider spacings. On the plus side, if you do your job and convince the builder and client to improve the building enclosure to get that load down from 25 Btu/hr•ft²(79W/m²) to 10 Btu/hr•ft² (32 W/m²), you can get that return temp down to 75°F(24°C) and that does not get any better from an energy efficiency perspective.

To put some magnitude around 75°F(24°C), that would be a nominal 20°F(11°C) cooler than your blood temp, ergo hot water heating is NOT! There is absolutely nothing hot about these temperatures – tepid yes, but hot not...got it?

Example 2 Radiant Floor Cooling (Figure 9)

In this example we are showing a higher performing building with a low flux (approximately 12 Btuh/ft²) (38 W/m²), and again, a low VOC conductive flooring with the same 9" (23 cm) o.c. spacing. Using the same process except working towards the upper part of the graph for floor cooling, the average cooling water temperature becomes t_w = t_a-12°F = 77°F-12°F = 65° F with a return temperature of 62.5°F(17°C) (based on an approximate 5°F (2.8°C) differential). Also make note that based on a space temperature of 77°F, the required floor temperature for cooling is only a nominal 77°F(25°C) -10°F(6°C)= 67°F (19°C) or approximately 4.5°F(3°C) above the return fluid temperature.

At an approximate surface temperature of 67°F(19°C) you would be within ASHRAE Standard 55 for floor temperature comfort. Controlling the space to 50 per cent RH and 77°F dry bulb (i.e. delivering a lean mixture (h_op=0.01476–h_load) to the space from the ventilation system) addresses microbial concerns (virus bacteria, moulds), provides for both stability of hygroscopic materials (woods), and occupant respiratory and thermal comfort. It also makes the condensation on the radiant cooling panel a moot point.

The FINAL WORD

The message in Part I regarding the use of the IA-Quest tool for source control over interior finishes is part and parcel of IAQ and CSA F326 Residential Mechanical Ventilation Systems,  ASHRAE 62.1 -Ventilation for Acceptable Indoor Air Quality, and  ASHRAE 62.2 - Ventilation and Acceptable Indoor Air Quality in Low-Rise Residential Buildings. It is also in the domain of the interior designer. As you can see from using the ASHRAE design graph, flooring also plays a big role in system efficiency and comfort, as well as IAQ. HVAC and interior design are glued together no matter how you cut it.

But here is the upside: clients like their interior designers. In fact, they like them more than they like their HVAC contractors (I know that hurts); and the typical client of an interior designer is an educated female with above average income.

Working with interior design professionals and her clients is a no-brainer for the progressive HVAC designer.

References
Figure 1 Design Graph for Heating and Cooling with Floor and Ceiling Panels is reprinted with permission from Section/page 6.8, Panel Heating and Cooling, 2012 ASHRAE Systems and Equipment Handbook.

Resources
i www.healthyheating.com/Boiler-efficiency.htm
iihttp://www.healthyheating.com/Effects-of-tube-depth-on-radiant-systems/Effects-of-tube-depth-on-radiant-systems.htm
iii Flooring affects boiler efficiency