Maintaining a comfortable environment in places of assembly

Designing comfort systems for places of assembly—such as auditoriums, gymnasiums, arenas or houses of worship—presents some vexing challenges. Such facilities often have acoustical requirements that place limits on equipment location and air distribution design. Many places of assembly experience diverse loads and occupancy schedules, complicating part-load system control.

By Don Eppelheimer, PE, Applications Engineer, Trane, Piscataway, N.J. October 1, 2007

Designing comfort systems for places of assembly—such as auditoriums, gymnasiums, arenas or houses of worship—presents some vexing challenges. Such facilities often have acoustical requirements that place limits on equipment location and air distribution design. Many places of assembly experience diverse loads and occupancy schedules, complicating part-load system control. But the biggest challenge is occupancy itself and its impact on ventilation and humidity control. Design guidelines that commonly are applied in commercial office space are troublesome here.

A simple example—a school gymnasium during a band concert—can illustrate some of these issues. As this is a talented high school band, both the bleachers and the floor are full. Occupancy is at the fire marshal’s rated seating capacity. The 18,000-sq.-ft. gymnasium is designed for 1,200 people, including use of the gym floor.

A load calculation reveals the following space loads:

Roof 69,600 Btuh
Wall 43,000 Btuh
Glass 10,500 Btuh
Lights 122,900 Btuh
People 300,000 Btuh (sensible)
Totals: 546,000 Btuh (sensible) 240,000 Btuh (latent)

Occupancy is major factor

People constitute a significant portion of the space sensible cooling load, more than 50%. However, it is the impact on humidity that makes occupancy a difficult load to manage. The space sensible heat ratio for this example is 0.69.

If the target space comfort condition is 75°F and 50% relative humidity (RH), and the air distribution system is designed for 55°F supply air, the required supply airflow is more than 4 cfm/sq. ft. of floor area—a huge amount. How did that happen? There are several ways to affect supply airflow.

Humidity ratio

Humidity ratio is grains of water vapor per pound of air. The humidity ratio at 75°F dry bulb and 50% RH results in the desired space condition of 64.7 grains/lb. The 55°F supply air has a humidity ratio of 60.4 grains/lb. At these conditions, each pound of supply air removes 4.3 grains of water vapor. If each occupant contributes 200 Btuh to the latent load, 1,200 people add 230 lbs. of water vapor—1.61 million grains (7,000 grains are in 1 lb.) to the air in the gymnasium. If each pound of supply air removes 4.3 grains of water vapor, it will take 374,000 lbs. of supply air per hour. This equates to approximately 83,000 cfm, or 4.6 cfm/sq. ft. That’s a lot of air.

In this example, 83,000 cfm is required to handle humidity, but only 18,000 cfm of this measurement is outdoor air (OA) for ventilation (assuming 15 cfm of OA per person, following the most recent version of ASHRAE Standard 62.1). With a space sensible cooling load of 546,000 Btuh and a supply-air temperature of 55°F; approximately 25,000 cfm is required to maintain the space temperature at 75°F. In this case, 72% of the supply air must be outdoor air. While this is a high fraction of outdoor air, it is manageable. Ventilation air is not the culprit.

This 25,000 cfm of supply air equates to 1.4 cfm/sq. ft. There still needs to be 83,000 cfm of supply air to control humidity. How do we better equip the supply air to handle the high latent load associated with this many people? Obviously, the supply air needs to be drier. The drier the supply air, which lowers the dew point, the more water vapor it removes from the space. What supply air dew point is required to handle the space latent load?

Calculating specific humidity

The key is another humidity measurement called specific humidity. Specific humidity is expressed as pounds of water vapor per pound of air. Suppose we choose to design the air distribution system for the example gymnasium for 25,000 cfm (114,000 lbs. per hour). The 1,200 people generate 227 lbs. of water vapor each hour. Removing 227 lbs. of water vapor with 114,000 lbs. of air requires that the specific humidity of the supply air be 0.0020 lb. of water/lb. of air drier than the space. The specific humidity at 75°F and 50% RH is 0.0092 lb. of water/lb. of air. So the specific humidity of the supply air must be 0.0072 lb. of water/lb. of air to offset the latent load of the people. This corresponds to a supply air dew point of about 48°F.

So how do we create this 48°F dew point supply air? One common method is to cool all of the supply air to a dry-bulb temperature of about 49°F to 50°F. This should dehumidify the supply air to the 48°F dew point required to offset the latent load due to people.

Desiccant approach

But do we need colder air, or do we just need drier air? The truth is, we need air that is drier. Recent research in desiccants show that a Type III desiccant wheel is able to regenerate at low temperatures, often without the need to add heat. This allows the wheel to be configured in series with a cooling coil. A desiccant wheel delivers cooler supply air than needed with the cooling coil. Use of a desiccant wheel eliminates the need to design a cold-air distribution system, and the required capacity of the cooling load is substantially reduced.

With the cold-air system, the required cooling coil capacity is about 150 tons (based on 1,200 people and 18,000 cfm of outdoor air) and supply fan power is 10 kW. The desiccant wheel reduces cooling coil capacity, but increases fan power because of the higher airflow and additional static pressure from the desiccant wheel. Both are viable options. It is noteworthy that the desiccant wheel may be a means to achieve low supply air dew points with conventional direct expansion equipment.

Don’t forget part load situations

Places of assembly often experience diverse loads. It is wise to evaluate the performance of these systems at part load. There are two part-load conditions we should evaluate. The first occurs when most of the people leave the gym. Perhaps the remaining occupancy is only 40 people instead of 1,200. This is an easy part-load condition to accommodate. The sensible loads drop to 256,000 Btuh and the latent load due to people drops to only 8000 Btuh. The resulting space sensible heat ratio increases to 0.97.

If we supply air at 50°F with the cold air system, the required supply airflow is only 9,400 cfm. This system is called single-zone variable-air-volume (VAV). Supply airflow is reduced to match the reduced sensible cooling load in the space. Single-zone VAV is easy to control. The supply fan airflow is modulated based on space temperature. The 9,400 cfm of 50°F air will remove the 256,000 Btuh of sensible heat and has the potential to remove 116 lbs. of water vapor. However, at this reduced occupancy, people add only 7.6 lbs. of water vapor. The result is that space humidity is lowered to 40% RH. At this condition, the supply-air temperature could be reset upward to save some compressor energy.

Constant volume system problems

What happens if the cold air system is a constant volume design rather than VAV? The reduced sensible cooling load requires a warmer supply air temperature, about 63°F for this example. At this supply-air temperature, the 20,000 cfm of supply air will remove 256,000 Btuh of sensible heat, but less than 7.6 pounds of water vapor. Space humidity rises to 65% RH, well above our target of 50% RH. Not only does a constant volume system use more fan energy at part load, but it is less adept at removing moisture. By comparison, a single-zone VAV system reduces fan energy while adequately removing moisture. Single-zone VAV with cold air provides humidity control at most load conditions, while simultaneously saving fan energy.

Light-sensible loads

The second part-load situation may be even more sinister: reduced building-related sensible loads while the space is fully occupied. What happens with full occupancy when there are no envelope or glass loads? If the only loads in the space are due to lighting and people, the sensible heat ratio drops to 0.63.

If we dim the lights, the situation gets even worse. This reduction in the space-sensible cooling load creates a sensible heat ratio more severe than what the system originally was designed to accommodate. Increasing the supply air temperature or reducing supply airflow in response to the reduced sensible load will hinder the ability to remove moisture. Reheat can help when the sensible heat ratio is lower than design. Both cold air and the desiccant wheel have the abilities to reduce sensible cooling capacity while maintaining a lower supply air dew point.

These two systems perform well at this part-load condition. Single-zone VAV results in a slightly elevated space relative humidity, but still well within the comfort zone. This comfortable condition is achieved without reheat and uses less fan energy. Some reheat and additional fan energy may be needed if more precise humidity control is desired.