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Role of safety factors in the design of dedicated outdoor-air systems.

INTRODUCTION

Safety factors are commonly used by engineers when designing various types of HVAC systems for use in all types of buildings. Excessive use of safety factors in the design process can result in larger-than-necessary equipment, inflated installed costs, and sometimes excessive energy use. This is especially true when safety factors are used during several steps of the design process, which "compounds" their impact along the way.

However, many of the decisions made during the design process are based on incomplete information, assumptions that may turn out to be invalid, or valid assumptions that may no longer be valid one, five, or ten years after the system or equipment is installed. Therefore, safety factors are important tools for design engineers, allowing the HVAC system to be designed with "reserve capacity" to accommodate unexpected loads and the need for increased airflow or dehumidification capacity.

A dedicated outdoor-air system (DOAS) uses a separate piece of equipment to condition (filter, heat, cool, humidify, dehumidify) all of the outdoor air brought into the building for ventilation. This conditioned outdoor air is then delivered either directly to each occupied space or to local HVAC units serving those spaces. Meanwhile, the local units (such as fan-coils, water-source heat pumps, PTACs, small packaged units, VAV terminals, chilled ceiling panels, or chilled beams) located in or near each space provide cooling and/or heating to maintain space temperature (Coad 1999, Shank and Mumma 2001)

Treating the outdoor air separately can make it easier to verify that sufficient ventilation airflow reaches each occupied space and can help avoid high indoor humidity levels. The latter is accomplished by dehumidifying the outdoor air to remove the entire ventilation latent load and most (or all) of the space latent loads, leaving the local HVAC units to primarily handle space sensible cooling loads. Some types of local HVAC equipment, such as chilled ceiling panels or chilled beams, must operate dry and avoid condensation. This limits their duty to handling sensible loads only.

Figure 1 shows several example DOAS configurations. Some deliver the conditioned outdoor air (CA) directly to each zone (Mumma 2008), while other configurations deliver the air to the intakes of local, single-zone units (such as fan-coils, water-source heat pumps, dual-duct VAV terminals, small packaged rooftop units, or single-zone air handlers) or to centralized, multiple-zone units (such as floor-by-floor VAV air handlers or self-contained units).

[FIGURE 1 OMITTED]

In addition, there are many types of dedicated outdoor-air equipment available (Figure 2). Dehumidification is usually provided by direct-expansion (DX) refrigeration, a chilled-water coil, a desiccant-based dehumidification device, or some combination of these technologies. Often, the dedicated outdoor-air unit includes an exhaust-air energy recovery device (such as a total-energy wheel, fixed-plate heat exchanger, coil runaround loop, or heat pipe), which can reduce energy use and allow for downsizing of the cooling and heating equipment. In fact, ASHRAE Standard 90.1 requires the use of an exhaust-air energy recovery device for many DOAS applications (ASHRAE 2007).

[FIGURE 2 OMITTED]

SAFETY FACTORS WHEN DETERMINING DESIGN AIRFLOW

This section discusses how safety factors impact the calculation of the design airflow for the dedicated outdoor-air unit.

Calculating the Design Airflow of the Dedicated Outdoor-Air Unit

In most applications, the design airflow for a dedicated outdoor-air unit is dictated by the amount of ventilation air required by industry standard or local code (Stanke 2004). In some cases, the owner or design team may choose to deliver more than code-minimum ventilation airflow to improve indoor air quality or to earn the "Increased Ventilation" credit when certifying a project using the LEED Green Building Rating System[TM] (USGBC 2009). Finally, in applications with very low ventilation requirements or very high indoor latent loads, the design engineer may chose to increase the airflow delivered by the dedicated outdoor-air unit so that the conditioned outdoor air can be delivered at a higher dew point (not as dry).

Table 6-1 of ASHRAE Standard 62.1-2007 (ASHRAE 2007) prescribes two ventilation rates for each occupancy category: one for people-related sources of contaminants (Rp) and another for building-related sources (Ra). Equation 6-1 from ASHRAE 62.1 is used to determine the minimum outdoor airflow (Vbz) that must be delivered to each breathing zone:

Vbz = Rp x Pz + Ra x Az (1)

where

Vbz = outdoor airflow required in the breathing zone of the occupiable space, cfm (L/s)

Rp = outdoor airflow rate required per person, cfm/person (L/s*person)

Pz = largest number of people expected to occupy the zone during typical usage

Ra = outdoor airflow rate required per unit area, cfm/[ft.sup.2] (L/s*[m.sup.2])

Az = occupiable floor area of the zone, [ft.sup.2] ([m.sup.2])

Next, Equation 6-2 and Table 6-2 from ASHRAE 62.1 are used to account for zone air distribution effectiveness (Ez), and to calculate the design outdoor airflow for the zone (Voz). This is the outdoor airflow that must be provided to the zone by the air distribution system (that is, through the supply-air diffusers).

Finally, for a 100% outdoor-air system in which one air handler supplies only outdoor air to one or more zones, Equation 6-4 from ASHRAE 62.1 is used to calculate the system-level outdoor air intake flow (Vot), by summing the zone outdoor airflows of all zones served by the dedicated outdoor-air unit:

Vot = [[summation].sub.all zones]Voz (2)

In some system configurations, the dedicated outdoor-air unit provides conditioned OA to the intakes of local or centralized HVAC units, rather than directly to each zone. In these configurations, the dedicated OA unit must be sized to deliver the sum of the outdoor air intake flows (Vot) required by each of the systems being served. If the conditioned OA (CA) is delivered to local HVAC units that each serve a single zone (bottom left configuration in Figure 1), Equation 2 above still applies since Equation 6-3 from ASHRAE 62.1 defines Vot equal to Voz for single-zone ventilation systems. However, if the conditioned OA (CA) is delivered to central air handlers that each serve multiple zones (bottom right configuration in Figure 1), Vot must be determined for each air handler using Section 6.2.5 (Multiple-Zone Recirculating Systems) of ASHRAE 62.1 (Stanke 2005) and then summed together to determine the airflow delivered by the dedicated OA unit.

Makeup Air Applications. In some applications, the design airflow for the dedicated outdoor-air unit is dictated by the need to replace air that is being exhausted from the building. This is common in laboratories, commercial kitchens, or other applications with very large exhaust requirements. In this case, the design airflow is the sum of all exhaust airflows plus any air needed for positive building pressurization.

Safety Factor for Zone Population

The first common use of a safety factor occurs when determining Pz, the number of people expected to occupy the zone. The definition of this term in ASHRAE 62.1 states that this is "the largest number of people expected to occupy the zone during typical usage." This is not the largest number of people that could conceivably be in the zone under any special circumstance (such as a tornado warning or a brief retirement or birthday celebration).

If actual zone population is not known, or if the owner and design team are not comfortable estimating it, Table 6-1 from ASHRAE 62.1 includes default values for occupant density. A note included under Table 6-1 clarifies that design engineers are not required to use this default, but may choose to if the actual occupant density is not known.

Some engineers use an overly-conservative estimate for Pz as a safety factor. In addition, some engineers or code officials use occupant load (or exit population), which is intended for use in designing a means of egress to comply with the local fire code. Occupant load is typically much higher than the expected zone population used for designing the ventilation system (Table 1).
Table 1. Comparison of Occupant Density and Occupant Load

                 Default Occupant        Minimum Occupant
                   Density, (1)             Load, (2)
              people/1000 [ft.sup.2]  people/1000 [ft.sup.2]
              (people/100 [m.sup.2])  (people/100 [m.sup.2])

Classroom               35                      50

Office space            5                       10

Retail sales            15                      33

(1) Table 6-1, ASHRAE Standard 62.1-2007 (ASHRAE 2007)

(2) Table 1004.1.1, 2006 International Fire Code (ICC 2006)


Using a larger-than-necessary value for zone population can result in significant over-ventilation and excessive energy use. In addition, as explained later in this paper, a larger-than-necessary value for zone population also impacts the calculation of the space latent cooling load, and the resulting dehumidification capacity of the dedicated OA unit.

The recommendation of this author is to avoid applying safety factors to zone population (Pz), instead using the best estimate for expected occupancy during typical usage. Hold off on applying safety factors until later in the process, as will be explained later in this paper.

Safety Factor for Future Expansion or Change of Use

The other common use of a safety factor when determining the design airflow of the dedicated outdoor-air unit occurs when accounting for future expansion or a change in use of the facility. If the facility is expanded in the future, and the dedicated OA unit will be expected to serve the expansion, it would be prudent to select a unit with some amount of extra (or reserve) airflow capacity. Or, if the use of the facility is likely to change in the future (from an office space to a group of meeting rooms or a retail area, for example) the ventilation requirements for the zones served by the dedicated OA unit may change. Again, it would be prudent to select a unit with some amount of extra airflow capacity.

The recommendation of this author is to calculate the design airflow as accurately as possible, without using safety factors. Then select a dedicated outdoor-air unit that has reserve airflow capacity, rather than a unit operating near its maximum allowable airflow. The impact of this will be discussed later in this paper.

SAFETY FACTORS WHEN DETERMINING DEHUMIDIFICATION CAPACITY

This section discusses how safety factors impact the calculation of required dehumidification capacity of the dedicated outdoor-air unit.

Calculating the Dehumidification Capacity of the Dedicated Outdoor-Air Unit

The required dehumidification capacity of a dedicated OA unit (Figure 3, Equation 3) is dictated by the design airflow (Vot), humidity ratio of the entering outdoor air (Woa), and humidity ratio of the conditioned air leaving the unit (Wca):

[FIGURE 3 OMITTED]

[q.sub.L] = 4.5 x Vot x (Woa - Wca)/7000 gr/lb ([q.sub.L] = 4.32 x Vot x [Woa - Wca]/1000 g/kg) (3)

where,

[q.sub.L] = required dehumidification capacity, lb/hr (kg/hr)

Vot = design outdoor airflow, cfm (L/s)

Woa = humidity ratio of the entering outdoor air, grains/lb (g/kg)

Wca = humidity ratio of the leaving conditioned air, grains/lb (g/kg)

Note: In Equation 3, 4.5 (4.32) is not a constant, but is derived from the density of air at "standard" conditions: 69[degrees]F (21[degrees]C) dry air at sea level has a density of 0.075 lb/[ft.sup.3] (0.0012kg/L). Air at other conditions and elevations will cause this factor to change.

0.075 lb/[ft.sup.3 x 60 min/hr = 4.5

(0.0012 kg/L x 3600 sec/hr) = 4.32

Since the previous section discussed the influence of safety factors on the design airflow (Vot) of the dedicated OA unit, this section will focus on the remaining two variables in Equation 3: the humidity ratio of the entering outdoor air (Woa) and the humidity ratio of the leaving conditioned air (Wca).

Safety Factor for Woa (Humidity Ratio of the Entering Outdoor Air)

The ASHRAE Handbook--Fundamentals (ASHRAE 2005) is a popular source for climatic data that represents the outdoor design conditions for various locations. Historically, design engineers have used the design dry-bulb temperature and mean coincident wet-bulb temperature when calculating the required capacity of cooling systems. However, the highest outdoor humidity ratio does not occur at the same time as the highest outdoor dry-bulb temperature.

Since 1997, the ASHRAE Handbook has included the design dew-point temperature and mean coincident dry-bulb temperature for use when calculating the required capacity of dehumidification systems. Table 2 lists the 0.4% outdoor design conditions for Jacksonville, Florida (ASHRAE 2005). Notice that the outdoor humidity ratio is 32% higher at the design dew-point condition than at the design dry-bulb condition. For this reason, it is very important to use the design dew-point condition when determining the required dehumidification capacity of a dedicated outdoor-air unit.
Table 2. Comparison of 0.4% Outdoor Design Conditions (Jacksonville,
Florida)

             Design Dry Bulb,    Design Dew Point,     Design Wet Bulb,
           Mean Coincident Wet  Mean Coincident Dry  Mean Coincident Dry
                   Bulb                 Bulb                Bulb

0.4%        95[degrees]F DBT,    78[degrees]F DPT,    80[degrees]F WBT,
            78[degrees]F WBT     85[degrees]F DBT     90[degrees]F DBT

Design     (35[degrees]C DBT,   (25[degrees]C DPT,   (27[degrees]C WBT,
condition   25[degrees]C WBT)    29[degrees]C DBT)    32[degrees]C DBT)

Humidity      116 grains/lb        144 grains/lb        141 grains/lb
ratioc         (16.6 g/kg)          (20.6 g/kg)          (20.2 g/kg)

Enthalpy       41.0 Btu/lb          42.9 Btu/lb          43.8 Btu/lb
              (95.5 kJ/kg)         (99.8 kJ/kg)        (101.9 kJ/kg)


The first common use of a safety factor when determining the required dehumidification capacity of a dedicated OA unit occurs when selecting the design humidity ratio of the entering outdoor air. The ASHRAE Handbook includes three sets of design conditions: 0.4%, 1%, and 2%. These indicate the percentage of hours during a year when with outdoor conditions are expected to exceed the tabulated design value. Table 3 lists the 0.4%, 1%, and 2% design dew point conditions for Jacksonville, Florida (ASHRAE 2005).
Table 3. Comparison of Design Dew-Point Conditions (Jacksonville,
Florida)

                   0.4%                   1%                 2%

Design      78[degrees]F DPT,    77[degrees]F DPT,    76[degrees]F DPT,
condition   85[degrees]F DBT     84[degrees]F DBT      83[degrees]F DBT
            (25.4 [degrees]C     (24.9 [degrees]C      (24.5 [degrees]C
                DPT, 29.2            DPT, 28.7            DPT, 28.3
             [degrees]C DBT)     [degrees]C DBT)       [degrees]C DBT)

Humidity   144 grains/lb (20.6  140 grains/lb (20.0  137 grains/lb (19.5
ratio             g/kg)                g/kg)                g/kg)


Many engineers tend to use the most conservative (0.4%) design condition, but this often results in the selection of a larger dedicated outdoor-air unit. The impact of this decision on equipment capacity will be demonstrated and discussed later in this paper.

Calculating the Humidity Ratio of the Conditioned Air Leaving the Unit (Wca)

The process for determining the required humidity ratio of the conditioned air delivered by a dedicated OA unit involves the following steps (Morris 2003):

1. Define the target humidity level for the occupied space (Wsp). This is the maximum indoor humidity that is considered acceptable. Some engineers consider 60% RH an acceptable upper limit, while others choose 50%, 55%, 65%, or some other limit.

2. Determine the design latent load for each zone ([q.sub.Lspace]). This is the latent load that occurs within the boundaries of the zone. It typically consists of moisture generated by people or other sources within the zone, and well as infiltration or diffusion of humid air from outdoors or from adjacent zones (Harriman, Brundrett, and Kittler 2001). These loads are often determined with the help of load calculation software.

3. Calculate the required humidity ratio of the conditioned air (Wca) delivered by the dedicated OA unit. If the dedicated OA system is being designed to offset the entire indoor latent load in addition to the ventilation latent load, then the conditioned outdoor air must be dry enough to offset the latent load in each zone, such that the humidity ratio in every zone is maintained at or below the desired upper limit. Equation 4 is used to determine the required Wca for each zone. Then, the humidity ratio of the conditioned outdoor air delivered by the unit must be the lowest (or worst-case) of all the zones it serves.

Wca = Wsp - [[q.sub.Lspace]/(0.69 x Voz)] (Wca = Wsp - [[q.sub.Lspace]/(3.0 x Voz)]) (4)

where,

Wca = required humidity ratio of the conditioned air leaving the unit, grains/lb (g/kg)

Wsp = desired upper limit for humidity ratio of the occupied space, grains/lb (g/kg)

[q.sub.Lspace] = design latent load in the zone, Btu/hr (W)

Voz = design zone outdoor airflow, cfm (L/s)

Note: In Equation 4, 0.69 (3.0) is not a constant, but is derived from the properties of air at "standard" conditions: density = 0.075 lb/[ft.sup.3] (0.0012 kg/L) and latent heat of water vapor = 1076 Btu/lb (2503 kJ/kg). Air at other conditions and elevations will cause this factor to change.

0.0075 lb/[ft.sup.3] x 1076 Btu/lb x 60 min/hr/7000 grains/lb = 0.69

(0.0012 kg/L x 2503 kJ/kg x 1000 J/kJ/1000 g/kg = 3.0)

To demonstrate this process, consider an example elementary school located in Jacksonville, Florida, USA. A dedicated outdoor-air unit is being designed to deliver air directly to four classrooms, removing the entire ventilation latent load and the entire indoor latent load.

1. For this example, the zone dry-bulb temperature setpoint during the cooling season will be 74[degrees]F (23[degrees]C) and the acceptable upper limit for indoor humidity is 60% RH. This equates to an indoor humidity ratio (Wsp) of 75.2 grains/lb (10.8 g/kg).

2. Table 4 lists the design latent loads ([q.sub.Lspace]) and required outdoor airflows (Voz) for each zone.
Table 4. Example of a DOAS Serving Four Classrooms

                       Classroom  Classroom  Classroom  Classroom
                          101        102        103        104

Wsp                      75.2       75.2       75.2       75.2
                       grains/lb  grains/lb  grains/lb  grains/lb
                         (10.8      (10.8      (10.8      (10.8
                         g/kg)      g/kg)      g/kg)      g/kg)


Space latent cooling     5250       5465       5697       5250
load, [q.sub.Lspace]     Btu/h      Btu/h      Btu/h      Btu/h
                       (1540 W)   (1600 W)   (1670 W)   (1540 W)

Zone outdoor airflow,   450 cfm    450 cfm    480 cfm    435 cfm
Voz                    (210 L/s)  (210 L/s)  (230 L/s)  (200 L/s)


Wca                      58.3       57.6       58.0       57.7
                       grains/lb  grains/lb  grains/lb  grains/lb
                         (8.34      (8.24      (8.30      (8.25
                         g/kg)      g/kg)      g/kg)      g/kg)


3. Equation 4 is used to calculate the required conditionedair humidity ratio (Wca) for each zone. For example, the required humidity ratio of the conditioned outdoor air delivered to Classroom 102 is 57.6 grains/lb (8.24 g/kg).

Wca = 75.2 grains/lb - [5465 Btu/hr / (0.69 x 450 cfm)] = 57.6 grains/lb

(Wca = 10.8 g/kg - [1600 W / (3.0 x 210 L/s)] = 8.24 g/kg)

As the results in Table 4 show, Classroom 102 requires the driest air. Therefore, for this example, the dedicated outdoor-air unit must be sized to dehumidify the outdoor air to 57.6 grains/lb (8.24 g/kg), which equates to a 52[degrees]F (11[degrees]C) dew point (drier than the space), in order to remove the entire indoor latent load (Figure 3).

Safety Factor for Wca (Humidity Ratio of the Conditioned Air Leaving the Unit)

Target Space Humidity Level (Wsp). The first influence of safety factors in determining the required humidity ratio of the conditioned air (Wca) is in the selection of the target humidity level for the occupied space (Wsp). For most comfort-cooling applications, engineers will typically design HVAC systems using a target relative humidity (RH) of 50%. However, the allowable upper limit (worst case) for indoor humidity may be higher. Some engineers consider 60% RH an acceptable upper limit, while others use 55%, 65%, or something different. The new ASHRAE Guide for Buildings in Hot and Humid Climates (Harriman and Lstiburek 2009) recommends limiting indoor humidity to no greater than 55[degrees]F (13[degrees]C) dew point, which equates to about 51% RH at 74[degrees]F (23[degrees]C) dry bulb.

This decision has a huge impact on the sizing of the dedicated outdoor-air unit. Using the same four-classroom example, the humidity ratio of the conditioned air (Wca) and the resulting dehumidification capacity of the dedicated OA unit were determined for various RH limits. As Table 5 demonstrates, in order to maintain drier air in the zones, Wca must be lower, resulting in greater dehumidification capacity required.
Table 5. Impact of Safety Factors on Wca and [q.sub.L]

                       60% RH           55% RH           50% RH

Wsp (1)            75.2 grains/lb   68.8 grains/lb   62.5 grains/lb
                    (10.8 g/kg)       (9.84 g/kg)      (8.94 g/kg)

Wca                57.6 grains/lb   51.2 grains/lb   44.9 grains/lb
                    (8.24 g/kg)       (7.33 g/kg)      (6.42 g/kg)

TDPca              52.0[degrees]F   48.9[degrees]F   45.4[degrees]F
                  (11.1[degrees]C)  (9.4[degrees]C)  (7.5[degrees]C)

Dedicated            100.8 lb/h       108.2 lb/h       115.6 lb/h
OA Unit             (45.7 kg/h)       (49.1 kg/h)      (52.4 kg/h)
Dehumidification
Capacity
([q.sub.L]) (2)

(1) Assumes the same 74[degrees]F (23[degrees]C) dry-bulb temperature
setpoint for the zones.

(2) Calculated per Equation 3 using the 0.4% design dew-point condition
for Jacksonville, Florida (included in Table 3), the design latent loads
and ventilation airflows for each zone listed in Table 4), and Vot =
[[SIGMA].sub.all zones] Voz = 1815 cfm (850 L/s).


For this example, designing the dedicated outdoor-air system to limit indoor humidity to 50% RH increases the required dehumidification capacity of the dedicated OA unit by 15% compared to allowing indoor humidity to rise to 60% RH at worst-case conditions.

However, be sure to analyze the impact on the entire system. When a chilled-water or DX dedicated OA unit delivers the conditioned air cold (that is, not reheated to "neutral"), a dedicated OA unit that is selected to deliver drier air (lower Wca) delivers the air at a colder dry-bulb temperature. This colder air offsets more of the space sensible cooling load, allowing for the local HVAC units to be smaller and use less energy (Murphy 2006).

Design Indoor Latent Loads ([q.sub.Lspace]). The second influence of safety factors in determining the required conditioned-air humidity ratio (Wca) is in the calculation of the design latent loads for each zone. As discussed earlier, the estimated zone population (Pz) not only impacts ventilation calculations, it also affects indoor latent load calculations.

However, another source of latent load in a zone is infiltration of humid air from outside. Calculating loads due to infiltration is often more of an art than a science because it is highly dependent on the quality of the building envelope construction (Harriman, Brundrett, and Kittler 2001). Since precise calculations are often difficult, many engineers choose to select the dedicated OA unit with extra (or reserve) dehumidification capacity, rather than attempt to apply safety factors to the calculation of the latent load due to infiltration.

SUMMARY AND DISCUSSION

Excessive use of safety factors in the design of a dedicated outdoor-air system (DOAS) can result in significant over-ventilation and larger-than-necessary equipment. This may inflate installed costs and can result in excessive energy use.

However, design decisions are often based on incomplete information and assumptions that may or may not be valid. Therefore, safety factors are important tools for a design engineer, allowing the system to be designed with "reserve capacity" to accommodate unexpected loads and the need for increased airflow or dehumidification capacity. To minimize the impact of safety factors on installed cost and energy use, the recommendations of this author are as follows:

1. Avoid applying safety factors to zone population (Pz) when calculating the required outdoor airflow (Voz) and design latent load ([q.sub.Lspace]) for each zone. Estimate population and loads as accurately as possible, then select a dedicated outdoor-air unit that has some extra (or reserve) airflow and/or dehumidification capacity. This may result in the selection of larger equipment, which can increase installed cost, but there are often energy-related benefits associated with a larger unit. The example in Table 6 shows two dedicated OA units selected to deliver the same airflow and provide the same dehumidification capacity. The larger casing size has more "reserve capacity" to accommodate unexpected loads and the need for increased airflow or dehumidification. The larger unit, when operating at the same design airflow, has a lower air-side pressure drop across both the cooling coil and the total-energy wheel, resulting is less fan power. In addition, the lower face velocity across the wheel results in higher effectiveness and more energy-recovery capacity.
Table 6. Energy-Related Impact of Equipment Sizing

                                  Smaller Unit         Larger Unit

Cooling coil face velocity          508 fpm              401 fpm
                                   (2.6 m/s)            (2.0 m/s)

Cooling coil pressure drop    0.96 in. [H.sub.2]O  0.60 in. [H.sub.2]O
                                    (240 Pa)             (150 Pa)

Wheel pressure drop           1.1 in. [H.sub.2]O   0.79 in. [H.sub.2]O
                                    (270 Pa)             (200 Pa)

Wheel total effectiveness             71%                  76%
(cooling)
Wheel capacity (cooling)            180 MBh              190 MBh
                                    (53 kW)              (56 kW)

Supply fan brake horsepower         4.8 hp               3.3 hp
                                    (3.6 kW)             (2.5 kW)

Exhaust fan brake horsepower        3.8 hp               2.5 hp
                                    (2.8 kW)             (1.9 kW)


2. When defining the design outdoor conditions (Woa) and the target space humidity level (Wsp) analyze the impact of these decisions up front and share that information with the owner to assist in making the final decisions. Using the same four-classroom example in Jacksonville (Table 4), Figures 4 and 5 demonstrate the impact of the selected ASHRAE design dew point condition (0.4%, 1%, or 2%) and the desired upper humidity limit for the occupied space (Wsp) on the required dehumidification capacity ([q.sub.L]) of the dedicated outdoor-air unit. This is the capacity needed to offset the entire ventilation latent load plus all of the indoor latent loads. (Figure 4 shows the required capacity if the unit has no exhaust-air energy recovery, while Figure 5 shows the required capacity if the unit includes a total-energy recovery device with 70% latent effectiveness.) These figures emphasize several points:

* Designing the DOAS for a more severe outdoor condition, or for a lower indoor humidity level, both increase the required dehumidification capacity of the dedicated OA unit. In this example, the result of these decisions can impact capacity by as much as 45%.

* Using total-energy recovery can significantly reduce the required capacity of the chilled-water coil, DX refrigeration circuit, or desiccant-based dehumidifier. But, it also dampens the impact of these more-conservative design decisions. Without exhaust-air energy recovery, designing the system for the 0.4% ASHRAE design dew point condition and 45% indoor RH increases the required dehumidification capacity by about 38.3 lb/h (17.4 kg/h) compared to using the 2% designs condition and 65% indoor RH (Figure 4). However, with total-energy recovery, the more-conservative design decisions increase the required dehumidification capacity by only 11.5 lb/h (5.2 kg/h) (Figure 5).

The use of safety factors is influenced by the type of local HVAC equipment that will be installed. For example, chilled ceiling panels and chilled beams must operate dry and avoid condensation. This typically results in designing the dedicated OA unit to maintain a lower indoor humidity level (Wsp), typically 50[degrees]F to 55[degrees]F (10[degrees]C to 13[degrees]C) dew point, and adding a safety factor to ensure that the unit has sufficient dehumidification capacity. On the other hand, fan-coils, water-source heat pumps, PTACs, or air handlers are usually capable of providing some dehumidification if needed, so the sizing of the dedicated OA unit may not need to be as conservative.

The use of safety factors is also influenced by the type of dedicated outdoor-air equipment being selected. For example, DX dehumidification equipment may be less tolerant of being selected with "reserve capacity" than chilled-water or desiccant-based equipment, because of the need to properly match airflow range with the operating envelope of the refrigeration equipment. Adding an exhaust-air energy recovery device helps overcome some of these challenges because there will be less variation to the condition of the air entering the dehumidifying coil.

REFERENCES

ASHRAE. 2005. 2005 ASHRAE Handbook--Fundamentals. Atlanta: American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc.

ASHRAE. 2007. ANSI/ASHRAE Standard 62.1-2007, Ventilation for Acceptable Indoor Air Quality. Atlanta: American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc.

ASHRAE. 2007. ANSI/IESNA/ASHRAE Standard 90.1-2007, Energy Standard for Buildings Except Low-Rise Residential Buildings. Atlanta: American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc.

Coad, W. 1999. Conditioning ventilation air for improved performance and air quality. HPAC Engineering (September): 49-56.

Harriman, L., G. Brundrett, and R. Kittler. 2001. Humidity Control Design Guide for Commercial and Institutional Buildings. Atlanta: American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc.

Harriman, L., and J. Ltstiburek. 2009. The ASHRAE Guide for Buildings in Hot and Humid Climates. Atlanta: American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc.

ICC. 2006. 2006 International Fire Code. Washington, DC: International Code Council, Inc.

Morris, W. 2003. The ABCs of DOAS: Dedicated outdoor air systems. ASHRAE Journal 45(5):24-29.

Mumma, S. 2008. Terminal equipment with DOAS: series vs. parallel. Engineered Systems 45(5):86-89.

Murphy, J. 2006. Smart dedicated outdoor air systems. ASHRAE Journal 44(3):23-31.

Shank, K. and S. Mumma. 2001. Selecting the supply air conditions for a dedicated outdoor air system working in parallel with distributed sensible cooling terminal equipment. ASHRAE Transactions 107(1):562-571.

Stanke, D. 2004. Standard 62.1: Single-zone and dedicated OA systems. ASHRAE Journal 46(10):12-20.

Stanke, D. 2005. Standard 62.1: Single-path, multiple-zone systems. ASHRAE Journal 47(1):28-35.

Stanke, D. 2005. Standard 62.1: Dual-path, multiple-zone systems. ASHRAE Journal 47(5):20-30.

USGBC. 2009. Leadership in Energy and Environmental Design (LEED) Green Building Rating System. Washington, DC: U.S. Green Building Council.

BIBLIOGRAPHY

Cummings, J. and D. Shirey. 2001. Separating the "V" from HVAC. ASHRAE IAQ Applications 2(3):20.

Dieckmann, J., K. Roth, and J. Brodrick. 2007. Emerging technologies: Dedicated outdoor air systems revisited. ASHRAE Journal 49(12):127-129.

Mumma, S. 2001. Designing dedicated outdoor air systems. ASHRAE Journal 43(5): 28-31.

Mumma, S. 2003. Decoupling OA and space thermal control. ASHRAE IAQ Applications 4(1):12-15.

Mumma, S. 2007. DOAS and desiccants. Engineered Systems 44(8):37-49.

Mumma, S. 2008. DOAS supply air conditions. ASHRAE IAQ Applications 9(2):18-20.

Murphy, J. and B. Bradley. 2002. Dehumidification in HVAC Systems (SYS-APM004-EN). La Crosse, WI: Trane.

Murphy, J. and B. Bradley. 2005. Advances in desiccantbased dehumidification. Engineers Newsletter 34(4). La Crosse, WI: Trane.

John Murphy

Member ASHRAE

John Murphy is an applications engineer with Trane Commercial Systems, La Crosse, WI.
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Geographic Code:1USA
Date:Jul 1, 2009
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