chilled beams; designing for thermal comfort - loudermilk

58 ASHRAE Jou rna l ash rae .o rg O c t o b e r 2 0 0 9

By Ken Loudermilk, P.E., Member ASHRAE

Designers also should devote ad-equate consideration to the specification and maintenance of space humidity levels where chilled beams are to be used. Designing for space humidity levels that are unnecessarily low can result in primary airflow rates that are higher than necessary, and neglecting space latent loads can lead to excessive

space humidity levels and potential condensation issues.

Principles of OperationFigure 1 illustrates an active chilled

beam. Primary air that has been cooled and dehumidified (1) is ducted from the central air-handling unit to a dis-tribution plenum within the beam from

which it is injected (2) through a series of nozzles. The velocity of the primary air jets entrain room air (3) through the beam’s integral heat transfer coil where it is reconditioned (4), then sub-sequently mixed with the primary air before this mixture is discharged to the space (5). The volume of induced room air is typically two to five times that of the primary air, depending on the size and design of the induction nozzles used, thus the discharge air mixture is three to six times the primary airflow rate. The ratio of the induced airflow rate to the primary (ducted) airflow rate is referred to as the beam’s induction ratio (IR).

About the AuthorKen Loudermilk, P.E., is the vice president of technology development for TROX USA, Cum-ming, Ga. He is vice chair of ASHRAE Technical Committee 5.3, Room Air Distribution, and chairs the TC 5.3 subcommittee on chilled beams.

DESIGNING

Chilled Beams Thermal Comfort

FOR

Active chilled beams have been used in Europe since the mid-1990s but

only recently have become a popular alternative to all-air HVAC systems

for nonresidential buildings in North America. Although active chilled beams

are an in-room space conditioning device, they are also the room air diffusion

device, so sizing and locating the beams is vital to providing acceptable levels

of occupant thermal comfort as established by ASHRAE’s thermal comfort

standard (55-2004).1

This article was published in ASHRAE Journal, October 2009. Copyright 2009 American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. Posted at www.ashrae.org. This article may not be copied and/or distributed electronically or in paper form without permission of ASHRAE. For more information about ASHRAE Journal, visit www.ashrae.org.

Octobe r 2009 ASHRAEJou rna l 59

The sensible cooling coil within the beam is supplied with chilled water whose supply temperature is maintained at, or above, the space dew-point temperature to prevent condensa-tion. The sensible heat removed by the coil typically constitutes 50% to 75% of the required space sensible heat removal. As a result, the primary airflow rate required to accomplish the space sensible cooling can be reduced accordingly.

Although primary (ducted) airflow rates associated with chilled beams are considerably lower than those in all-air sys-tems, their discharge airflow rate to the room is always greater. Since the chilled water supplied to the beam is maintained above the space dew point, the beam’s off-coil temperature will be higher than the primary air temperatures used in all-air systems. The resultant temperature of the beam’s discharge mixture is typically 3°F to 6°F (2°C to 3.3°C) warmer than that of all-air systems. Therefore, a proportionally higher (20% to 30%) discharge airflow rate to the space must be provided. This higher discharge flow rate often contributes to greater draft risks, which may compromise occupant thermal comfort levels.

Designing for Occupant Thermal ComfortStandard 55-20042 defines the occupied zone as the portion of

a space where occupants normally reside. It is further quantified as the volume of the room that is (1) no closer than 3.3 ft (1 m) from any outside walls or windows nor within 1 ft (0.3 m) of any internal wall and (2) is vertically bounded by the floor and the head level of the predominant space occupants. Although the head level is often accepted to be 67 in. (1.7 m) for stand-ing occupants, the standard allows the designer to define that height according to the space occupancy.

For example, if a space is predominantly occupied by seated persons, the occupied zone height could be considered as 42 in. (1.1 m). Chapter 20 of the 2007 ASHRAE Handbook— HVAC Applications3 predicts the percentage of occupants who might express thermal dissatisfaction for various combinations of local air speed and temperatures. Figure 2 (from that chapter) can be used to predict the percentage of occupants that will object to various air speeds and temperatures at the neck and ankle regions. As active chilled beams are normally mounted overhead, the neck region is usually the most critical. Comfort cooling applications should strive to minimize dissatisfaction levels, and in all cases limit the percentage of occupants object-ing to these local conditions to 20% or less.

Room Air Distribution Active chilled beams distribute air within the room in a

manner consistent with that of linear slot diffusers. As such, relationships between airstream terminal velocities and thermal decay of the supply airstream that apply to linear slot diffusers also apply to active chilled beams. Upon discharge to the open space, velocity and temperature differentials between the sup-ply air mixture and the room begin to diminish due to room air entrainment. Linear slot diffusers exhibit relatively long throw characteristics and their velocity and temperature differentials

diminish at a rate that is proportional to the distance the air has traveled within the space.

Manufacturers publish throw values that allow designers to estimate the travel distance of the airstream before it reaches a given terminal velocity. Most manufacturers present such data using isothermal air for terminal velocities of 150, 100 and 50 fpm (0.75, 0.5 and 0.25 m/s). These data can be used to map the airstream and predict the local velocity at the point where it enters the occupied zone. As the room-to-supply-air-differential decays at a similar rate, its temperature also can be predicted at the entry point based on the initial temperature difference (ΔTO) between the beam discharge temperature and that of the room into which it is introduced. Manufacturers supply selection software that can be used predict the value of local velocities and temperatures at critical locations such as that where the airstream enters the occupied zone.

Figure 1 illustrates a space being served by two active beams with two-way discharge patterns delivering identical primary (QP) and discharge (QS) airflow rates. The discharge airflow rate is a function of the induction ratio of the nozzles chosen and is calculated by multiplying the primary airflow rate by the induction ratio. Assume a beam produces an induction ratio of 2.5 and is sized to deliver 100 cfm (170 m3/h) of 55°F (13°C) primary air to a 75°F (24°C) room. Also, assume that chilled water enters the beam at 57°F (14°C) and leaves at 61°F (16°C). The discharge airflow rate to the space will be 3.5 times the

Figure 1: Application of active chilled beams.

1

2

4

5

3

A

Occupied Zone(Height Determined By Designer)

Beam A Beam BQA QA

H1

VL ’ TL VH1 ’

TH1

VO ’ TO

QS

VO ’ TO

QSVC

1 m (0.3 ft)


primary airflow rate or 350 cfm (595 m3/h). The temperature (TOC) of the air leaving the beam’s integral cooling coil can be conservatively estimated as 1°F (0.6°C) warmer than its mean chilled water temperature, which is the average of the entering and leaving chilled water temperatures. In fact, the leaving air temperature would typically be at least 2°F to 4°F (1°C to 2°C) higher than the coil mean water temperature. The beam manufacturer has this information as well as the beam’s IR. Upon identifying the primary air temperature (TPA), the temperature (TZ) leaving the beam as well as the initial room to supply air temperature difference (ΔTZ) can be estimated using Equations 1 and 2.

TZ

= [TPA + (TOC × IR)] / (IR+1) (1)

ΔTZ = TROOM – TZ (2)

or, for this example,

TZ = [55°F + (60°F × 2.5)] / (2.5 + 1) = 58.6°F and ΔTZ = 75°F – 58.6°F = 16.4°F

The initial velocity (VO ) of a supply airstream leaving the discharge slot can be calculated by dividing the supply airflow rate leaving each slot (for two-way beams this is 0.5 x QS) by the effective area of that slot. If the effective area is not available, VO can be conservatively estimated as 450 fpm (2.3 m/s) for the purposes of this calculation. The local temperature difference (ΔTX) between the room and the supply airstream at any point along its travel can be estimated by Equation 3.4

ΔTX= 0.8 × ΔTZ × (VX / VO) (3)

For a beam with an initial discharge velocity (VO ) of 450 fpm (2.3 m/s) and an initial supply to room temperature differential (ΔTZ) of 16.4°F (9.1°C), the local temperature differential (ΔTX) at the point coincident with a 50 fpm (0.25 m/s) terminal veloc-ity is about 1.4°F (0.9°C). Referring to Figure 2, this predicts that less than 20% of the occupants would be dissatisfied with these local velocity/temperature conditions.

As the region near the outside walls is not defined as part of the occupied zone, local velocities and temperatures do not generally affect occupant thermal comfort. Care should still be taken that they are not so high that they can affect processes (e.g., fume hoods) along the outer wall and that they are suf-ficient to provide adequate heating where applicable. Chapter 56 of the 2007 ASHRAE Handbook—HVAC Applications5 recom-mends that outlets used for perimeter heating be selected and located such that their isothermal throw to 150 fpm (0.75 m/s) extends at least halfway down the outside surface or to a level 5 ft (1.5 m) above the floor, whichever is greater.

The area of greatest draft risk usually occurs directly below the point where two opposing airstreams collide. Figure 1 il-lustrates such a point and defines the collision velocity as VC . If the velocity (VC ) of the colliding airstreams is of sufficient velocity (greater than 100 fpm or 0.5 m/s), some of the velocity and temperature differential of the airstreams is dissipated by the collision and the velocity (VH1 ) at the point where the airstream

enters the occupied zone is reduced accordingly. Figure 3 can be used to estimate the velocity at the entry point for various collision velocities and vertical distances (H1) between the point of collision and the top of the occupied zone. This figure illustrates that the velocity (VH1 ) entering the occupied zone directly below the point of collision of two airstreams will be less than half the collision velocity (VC ) if distance H1 (the distance between the ceiling and the top of the designated oc-cupied zone) is greater than or equal to 3.5 ft (1.1 m).

In cases where the collision velocity (VC ) is 100 fpm (0.5 m/s) VH1 would be less than or equal to 50 fpm provided H1 is greater than 3.5 ft (1.1 m). In cases where H1 is greater than 3.5 ft (1.1 m), active chilled beams should be sized and located so that their throw to a terminal velocity of 100 fpm (0.5 m/s) does not exceed half the distance to another beam with an opposing blow.

In cases where the collision velocity (VC ) is 150 fpm (0.75 m/s), VH1 is less than or equal to 50 fpm provided H1 is greater than 6 ft (1.8 m). Active beams for which H1 is greater than

Figure 2: Percentage of occupants objecting to drafts in air.3

www.info.hotims.com/25209-34


or equal to 6 ft (1.8 m) can be sized such that their throw to a terminal velocity of 150 fpm (0.75 m/s) does not exceed half the distance to a beam with an opposing discharge.

Figure 3 can be used to determine the maximum collision velocity that limits velocities within the occupied zone to 50 fpm (0.25 m/s) or less for any given distance H1 using the fol-lowing equation:

VC = 50 / (VH1 / VC ) (4)

where VH1 / VC is the value from Figure 3 that corresponds to the distance H1.

For example, if H1 is equal to 4 ft,

VC = 50 / 0.43 = 116 fpm

Room Humidity Design Considerations The sensible cooling contribution with chilled beams af-

fords the designer an opportunity to significantly reduce the primary airflow rate compared to all air systems. As 50% to 75% of the sensible heat gains are typically removed by the chilled water coil, proportional reductions in the primary airflow rates within the system may be achievable. However, this should be done with caution as the beam must also de-liver sufficient ventilation air and maintain acceptable space humidity levels. The primary airflow rate to the room must be the greater of that required to (1) ventilate the space in conformance to ASHRAE Standard 62.1-20076 (or other applicable ventilation codes); (2) offset space latent gains to control the room humidity level within ASHRAE Stan-dard 55-2004 recommendations; and (3) provide sufficient sensible cooling to complement the sensible heat removed by the chilled water coil.

In most common interior space applications, the primary airflow rate required to offset space latent gains will exceed both the ventilation airflow rate and the airflow rate required to complement the coil’s sensible cooling. The space airflow rate will be determined by the latent gains and the design room humidity ratio (WROOM). In perimeter spaces, the primary airflow rate to the beams will be driven by the sen-sible load (laboratories may be exceptions due to their high mandated ventilation rates). The use of beams whose water side cooling capacity contributes to more than about 65% of the total space load may be impractical due to architec-tural constraints that limit the installed beam quantities and lengths in these spaces. The goal of the design should be to reduce the primary airflow rate to as close to the required ventilation rate as possible.

Designing chilled beam systems to maintain room air hu-midity levels lower than necessary can result in considerably higher primary air requirements. Figure 4 is presented in Standard 55-20047 and prescribes acceptable ranges of room temperatures and humidity ratios. Assuming a clothing level of 1.0 clo (0.15 m2 · k/W), this diagram defines the thermal comfort window for a dry-bulb temperature of 75°F (24°C) to include room dew-point temperatures as high as 62°F

(16.7°C). Where chilled beams are applied, the room dew-point temperature must not exceed the chilled water supply temperature, so design for dew points above about 57°F (14°C) is not recommended.

Most conventional HVAC systems condition delivered air to about a 52°F (11°C) dew-point temperature, which coincides with a humidity ratio (WPRIMARY) of 58 grains (3.8 g). The airflow rate (QPRIMARY) to offset the space latent heat gains (qLATENT) can be determined by Equation 5.

QPRIMARY = qLATENT / [0.68 × (WROOM – WPRIMARY)] (5)

Using this equation and assuming the humidity ratio of the primary air is 58 grains (3.8 g), the primary airflow requirement

Figure 3: Velocities entering the occupied zone.

7.0

6.5

6.0

5.5

5.0

4.5

4.0

3.5

3.0

2.5

Dis

tanc

e H

1 (f

t)

VH1/VC

0.3 0.4 0.5 0.6 0.7

Figure 4: ASHRAE Summer and Winter Comfort Zones. Acceptable ranges of operative temperature and humidity with air speed ≤ 40 fpm (≤0.20 m/s) for people wearing 1.0 and 0.5 clo (0.15 m2 · k/W and 0.08 m2 · k/W) clothing during primarily sedentary activity (≤1.1 met [≤63.9 W/m2]).8

64 ASHRAE Jou rna l O c t o b e r 2 0 0 9


for a space whose latent gains total 400 Btu/h (117 W) can be calculated for various design humidity ratios:

If WR OOM = 65 grains (50% RH) → QPRIMARY = 84 cfm



In this case, designing for 75°F (24°C) and 50% relative humidity would result in a primary airflow rate that is 58% higher than that required to maintain 53% RH in the space. As the 53% RH is within Standard 55-2004 recommendations and results in a space dew-point temperature of 57°F (14°C), it is probably a reasonable design goal.

Chilled beams are often used with central HVAC equipment that includes heat recovery and enthalpy wheels. Lower space dew-point temperatures can be achieved when the dew-point temperature is further suppressed by these processes. In cases where room dew-point temperatures below 55°F (13°C) are desired, using such equipment is recommended.

SummaryActive chilled beams can be selected to remove large amounts

of sensible heat while substantially reducing primary airflow requirements. However, this must be done with consideration

of the occupant thermal comfort and space dehumidification. Because producing high levels of thermal comfort is the primary objective of any comfort cooling application, beams should be selected, sized and located with that in mind.

While primary airflow reduction opportunities are an inherent characteristic of chilled beams, the reduction of such should be limited to that required to provide adequate space humid-ity control. All-air systems almost always deliver a sufficient amount of dry air to satisfy the space sensible load, therefore, engineers often do not consider space latent loads in their selec-tion. Individual space latent loads should be considered when designing chilled beam systems.

In conclusion, the following design guidelines should be ob-served when selecting, sizing and locating active chilled beams:

• Chilled beams should not be used in low ceiling height applications where the distance between the ceiling and the top of the occupied zone is less than 3 ft (0.9 m).

• When applied in lobbies, atriums or other areas with high and/or uncontrollable infiltration rates, provide adequate condensation prevention strategies.

• To maintain high levels of thermal comfort (veloci-ties within the occupied zone no greater than 50 fpm or [0.25 m/s]), active chilled beams were mounted at least 3.5 ft (1.1 m) above the designated occupied zone should be sized and located such that their throw to a terminal velocity of 100 fpm (0.5 m/s) does not exceed half the distance between them and another beam with an opposing blow. Active beams mounted 6 ft (2 m)or more above the designated occupied zone may be located such that their throw to a terminal velocity of 150 fpm (0.75 m/s) is as much as half the distance between the beam and an adjacent beam with an opposing discharge.

• Smaller nozzles result in higher induction ratios and higher sensible cooling capacities per cfm (m3/h) of primary air. However, the use of smaller nozzles gen-erally results in higher noise levels and inlet pressure requirements for a given primary airflow rate that increases the number of beams required.

• Designing for space humidity levels lower than that actually required may result in significantly higher primary airflow rates.

References1. ANSI/ASHRAE Standard 55-2004, Thermal Environmental

Conditions for Human Occupancy. 2. Standard 55-2004, p. 3.3. 2009 ASHRAE Handbook—Fundamentals, p. 20.13.4. Koestal, A. 1954. “Computing temperature and velocities in

vertical jets of hot or cold air.” ASHVE Transactions 60:385.5. 2007 ASHRAE Handbook—HVAC Applications, p. 56.4.6. ANSI/ASHRAE Standard 62.1-2004, Ventilation for Acceptable

Indoor Air Quality, Table 6-1.7. Standard 55-2004, p. 5.2.1.1.8. 2009 ASHRAE Handbook—Fundamentals, p. 9.12.

chilled beams; designing for thermal comfort - loudermilk

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