Energy Efficient BuildingsOutdoor Air Control

IntroductionDuring the design phase of a HVAC system, the volume of outdoor air that must be supplied by an air handler during peak occupancy conditions is calculated. Buildings with large ventilation demands, such as locker rooms and laboratories, may require 100% outside air. However, the peak fraction of outdoor to supply air in many buildings is typically between 15% and 30% outside air. This peak value is used when calculating peak cooling and heating loads. Although some of the cooling and heating supplied to a building is to meet sensible and latent zone loads, a significant fraction is to condition outside air. Thus, reducing the outdoor air ventilation requirement below peak values has the potential for significant energy savings.

To control the fraction outside air, foa, introduced into a building the outside air, exhaust air and mixed air dampers (shown below) are modulated in a coordinated fashion between the extremes shown in the table below.

Foa Exhaust Air Damper Mixed Air Damper Outside Air Damper0.0 0% open 100% open 0% open0.3 30% open 70% open 30% open1.0 100% open 0% open 100% open

In constant-air-volume air handlers without active damper control, the damper positions are fixed during installation of the air handler. Thus, the quantity of outdoor air introduced into the building remains constant throughout the year. Active damper control allows the

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quantity of outdoor air introduced into the building to vary. This creates the opportunity for significant energy savings.

Varying the quantity of outside air in accordance with the occupancy of a building is called demand control ventilation. Varying the quantity of outside air to take advantage of outside air conditions is called economizer control. Both types of control can reduce the load on the cooling coil. In addition, energy recovery ventilators can exchange energy between exhaust and intake air streams to further reduce coil loads.

Advanced control systems employ both demand and economizer control of outdoor air to minimize energy use while meeting ventilation requirements. As with many energy systems, the potential savings are large since the peak conditions used for sizing systems rarely occur.

Demand Control VentilationThe ASHRAE standard for ventilation, ASHRAE 62.1, permits an HVAC system to vary outdoor air flow into a space as long as the air handler delivers the required amount of breathing air to all occupied zones. According to ASHRAE 62.1-2007, the required quantity of outdoor, Vbz, air is:

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

WhereRp = people outdoor air rate (typically 5 to 10 cfm/person depending on space type)Pz = number people in zoneRa = area outdoor air rate (typically 0.06 to 0.12 cfm/ft2 depending on space type)Az = floor area of zone

The area component of the requirement is to remove contaminants from outgassing of carpets, furniture, etc. It also represents the minimum required outdoor air.

Example

Calculate the required volume of air in a 1,000 ft2 space occupied by 10 people if Rp = 10 cfm/person and Ra = 0.12 cfm/ft2.

Vbz = (Rp x Pz) + (Ra x Az) Vbz = (10 cfm/person x 10 people) + (0.12 cfm/ft2 x 600 ft2) Vbz = 1,000 cfm + 120 cfm = 1,120 cfm

The three primary strategies for varying ventilation due to demand are:

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Occupancy schedules: supply design outdoor air during occupied hours as scheduled, with minimum or no outside air during unoccupied hours.

Occupancy sensors: supply design outdoor air during occupied times as sensed, with minimum or no outside air when zone is unoccupied.

CO2 sensors: supply sufficient outdoor air to keep CO2 concentration within bounds.

The first two strategies rely on estimates and approximations to increase outdoor air ventilation during occupied periods and reduce it during unoccupied periods. Although these strategies are significant improvements over supplying a fixed quantity of outdoor air, they do not minimize energy use or guarantee adequate ventilation. The use of CO2 sensors allows much better control over ventilation and is further developed below.

CO2 Sensor ControlTo understand how CO2 sensors can be used to control ventilation, consider a zone ventilated with outdoor air. CO2 enters the zone in the ventilation air. The quantity of CO2 in the ventilation air is the product of the outdoor air flow rate Voa and the concentration of CO2 in the outside air Coa. CO2 also enters the room from the occupants at a rate Vco2. CO2 leaves the room in the exhaust air at concentration Cz.

A steady-state mass balance on the CO2 entering and leaving the zone gives:

VOA COA + VCO2 – VOA CZ = Ø (SS)

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This relation can be rearranged to give the difference in CO2 concentration between indoor and outdoor air:

VCO2 = VOA CZ – VOA COA VCO2 = VOA (CZ – COA) (CZ – COA) = VCO2 / VOA

The CO2 emitted from a typical person at light work VCO2 is about 0.0105 cfm/person. Depending on the type of space, ASHRAE 62.1 – 2007 specifies an outdoor air ventilation rate VOA of about 10 cfm/person. Thus, to meet this code, the volume of outdoor air can be controlled to a CO2 concentration difference between indoor and outdoor air of about: (Cz – COA) = VCO2 / VOA (Cz – COA) = 0.0105 cfm CO2/person / 10 cfm OA /person (Cz – COA) = .00105 ft3 CO2 / ft3 OA (Cz – COA) = 1.05/1,000 = 10.5/10,000 = 105/100,000 = 1,050/1,000,000 = 1,050 ppm

The concentration of outdoor air, COA, is generally about 400 ppm. To reduce cost and control complexity, many systems use only one CO2 sensor in the zone and assume that the concentration of CO2 in the outdoor air is constant. If so, then outdoor air can be controlled to about:

CZ – COA = 1,050 ppm CZ = COA + 1,050 ppm CZ ≈ (400 + 1,050) ppm CZ ≈ 1,450 ppm

A key metric for characterizing the quantity of outdoor air introduced into a building is the fraction of outdoor to supply air, foa.

foa = Voa / Vsa

To control the fraction outside air introduced into a building the outside air, exhaust air and mixed air dampers are modulated in a coordinated fashion to maintain the concentration of CO2 at the desired level.

Demand Control Ventilation Using CO2 Sensors in Single-Zone SystemsIn single zone systems, the quantity of ventilation air can be controlled based on a single CO2 sensor in the zone by modulating the discharge, mixed air and outdoor air dampers.

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Demand Control Ventilation Using CO2 Sensors in Multi-zone SystemsTwo strategies are frequently used to control outdoor air ventilation in multi-zone systems: the first uses a single CO2 sensor in the return air duct and the second employs multiple CO2 sensors in the zones.

Single CO2 Sensor in Return Air DuctWhen ventilation air in a multi-zone system is controlled by a single CO2 sensor, the sensor is generally located in the return air duct. The sensor controls the dampers to maintain ~ 1,100 ppm CO2. The sensor monitors the average concentration; thus some zones are over ventilated and some zones under ventilated. This type of system successfully reduces energy use by reducing foa below foa at design (peak) conditions at low first cost, but doesn’t strictly meet code or guarantee occupant comfort.

Multiple CO2 Sensors in Each ZoneA better strategy is to place one CO2 sensor in each zone. The building automation system (BAS) then determines how much OA needed in each zone, and sets the building outdoor air to meet critical zone. Thus, some zones are over ventilated, but no zone is under-ventilated. The system is often modified to save initial and control costs by placing CO2 sensors only in zones likely to set the maximum demand for outdoor air.

To understand how multi-zone demand ventilation control with multiple CO2 sensors works, consider the example below for a VAV system. The total volume flow rate Vsa and outdoor air flow rate Voa at design conditions for three zones are shown below. Based on these requirements, the fraction outdoor air, foa, at design conditions is 0.35.

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Design Z1 Z2 Z3 TotalVsa(cfm) 2,000 4,000 4,000 10,000VOA (cfm) 500 1,000 2,000 3,500

foa 0.35

Next, consider some a typical operating condition with the following supply air and outdoor air requirements. Zone 3 requires a higher fraction of outdoor air than the other zones; thus zone 3 is the critical zone and the fraction outdoor air for the entire building is set at 0.27. Although the actual outdoor air supplied to zones 1 and 2 exceeds the minimum requirement, all three zones meet the outdoor air requirement. In addition, heating and cooling energy use is reduced because the quantity of outside air introduced into the building (1,890 cfm) is less than the quantity of outside air if the dampers were fixed at the design fraction outdoor air (0.35 x 7,000 cfm = 2,450 cfm).

Design Z1 Z2 Z3 TotalVsa(cfm) 1,000 3,000 3,000 7,000

VOA,REQUIRED (cfm) 200 600 800 1,600Foa 0.20 0.20 0.27

VOA,ACTUAL (cfm) 1,000 x .27 = 270 3,000 x .27 = 810 3,000 x .27 = 800 1,890

Temperature-Based Economizer Control Demand control ventilation strategies determine the minimum quantity of outside air required to properly ventilate a building. However, in some cases it may be advantageous to use more than the minimum amount of outside air. Consider the single-duct reheat system shown below. The mixed air is cooled to the supply air temperature by the cooling coil. Thus, the cooling energy use can be minimized by varying the fraction of outside air so that the mixed air temperature is as close to the supply air temperature as possible.

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Single-duct reheat/VAV system.

The optimal strategy for minimizing cooling coil energy use by adjusting damper positions is:

If Toa > Tra, use minimum outside air. If Tsa < Toa<Tra, use 100% outside air to minimize cooling If Toa < Tsa, blend outside air with return air to the mixed air temperature equals the

supply air temperature, but maintain outside air above the minimum requirement

This strategy for controlling fraction outdoor air is shown graphically below for the case when cooling supply air temperature is 50 F and the return air temperature is 72 F.

When TOA < Tsa , the fraction outdoor air required to maintain the mixed air temperature at the supply air temperature can be determined from an energy balance on the mixing box. TMA = fOA TOA + (1-fOA) TRA

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10 20 30 40 50 60 70 80

Toa (F)

foa

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TMA = fOATOA + TRA – fOA TRA

fOA = TMA-TRA / TOA – TRA

Economizers vary foa so that the mixed air temperature equals supply air temperature. Setting TMA = TSA in the preceding equation gives:

fOA = TSA-TRA / TOA-TRA

Thus, optimal temperature-based economizer control can be summarized as:

If TOA > TRA thenfOA = fOA min

Else if TOA > TSA and TOA < TRA then fOA = 100%

Else if TOA <TSA then fOA = (TSA – TRA) / (TOA - TRA)

End if

The energy savings from economizer control depends on the outdoor air conditions. In hot, humid weather, economizer control generates no savings compared to simply setting fraction outdoor air to the minimum required for ventilation. As the weather moderates, economizer control generates increasing savings. Thus, economizer control generates the largest savings in buildings that require cooling all year round and run air conditioning equipment to meet that load. In these buildings, economizer control can be so effective that chillers or compressors do not run at all during winter months. Economizer control generates smaller savings in buildings that require cooling only during summer months.

Enthalpy-Based Economizer Control Although temperature-based enthalpy control leads to reduced cooling coil energy use, it can actually increase the load on the cooling coil in when the outdoor air is cool and humid. For example consider the case shown below when the outside air has a higher enthalpy than the return air, even though it is at a lower temperature.:

Return air: TRA = 72 F, ǾRA = 60% > hRA = 28.3 Btu/lba Outside air: TOA = 70 F, ǾOA = 90 % > hOA= 32.4 Btu/lba

Supply air: TOA = 55 F, ǾOA = 100 % > hSA= 23.3 Btu/lba

To account for these conditions, enthalpy-based economizer controls measure both temperature and humidity of the return and outside air and calculate the enthalpy of the air streams. Dampers are then controlled using the temperature-based strategy, except that

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temperature is replaced by enthalpy as the primary control variable. An optimal enthalpy-based economizer control strategy would be:

If hOA > hRA thenfOA = fOA min

Else if hOA > hSA and hOA < hRA then fOA = 100%

Else if hOA < hSA then fOA = (hSA – hRA) / (hOA - hRA)

End if

The following relations can be used to calculate air enthalpy.

w (lbw/lba) = exp (-7.1283 + 0.5074 Tdp(F) - 0.0001120 Tdp(F)2) h (Btu/lba) = .24 T(F) + w (lbw/lba) (1061 + .444 T (F))

In many climates, enthalpy-based control results in only a modest improvement over temperature-based control. In addition, enthalpy control also increases first and control costs, and can become unreliable if a humidity sensor fails. For these reasons, the use of enthalpy control over temperature control should be weighed against the energy savings for a given climate.

Energy Recovery Ventilators Some buildings or zones, such as locker rooms and labs, may require 100% outside air for ventilation. In these applications, energy use to heat and cool the outside air can be significant. This energy use can be reduced by installing an air to air heat exchanger between the incoming and exhaust air streams. The heat exchanger will pre-heat cold incoming air during winter and pre-cool hot incoming air during summer, reducing both heating and cooling costs. Some designs, such as rotating heat recovery wheels, may actually reduce latent loads in addition to sensible loads.

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Public domain images of rotating heat recovery wheels. Sources: 1: http://buildingdata.energy.gov/sites/buildingdata.energy.gov/

2: http://homeenergypros.lbl.gov/

Energy savings are determined by the volume of exhaust and intake air streams and by the effectiveness of the heat exchanger, , the product of the mass flow rate and specific heat, mcp, and the temperatures of the hot, Th, and cold, Tc, streams. Using the heat exchanger effectiveness method, actual heat transfer, Q, is:

Q = mcpmin (Th – Tc)

When the volume flow rates of the exhaust and intake air streams are identical, actual heat transfer reduces to:

Q = mcp (Th – Tc) = V pcp (Th – Tc)

where V is the volume flow rate and pcp is the product of the air density and specific heat.

Performance data for a typical heat wheel are shown below. Heat exchanger effectiveness increases as face velocity decreases. Heat exchanger effectiveness also increases as the flow rates between the two air steams become less equivalent. Pressure drop increases as face velocity increases. Thus, decreasing face velocity by specifying large heat exchangers relative to the flow increases thermal energy savings while decreasing fan power requirements.

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ExampleConsider a building maintained at 72 F that intakes and exhausts 10,000 cfm of outside air. The building uses an air-to-air heat exchanger between the intake and exhaust air streams with face velocity = 500 ft/min. Determine the cooling coil energy savings (Btu/hr) if the outside air temperature is 90 F and the heating coil energy savings (Btu/hr) when the outside air temperature is 30 F. Determine the additional electricity power required by the supply and return air fans due to the increased pressure drop across the heat wheel if the fans are 70% efficient and the fan motors are 90% efficient.

From the performance data given above, the effectiveness is 0.77.

Q = V pcp (Th – Tc)

Summer: Q = 0.77 10,000 (cfm) 60 (min/hr) 0.018 (Btu/ft3-F) (90 – 72) (F) = 149,688 (Btu/hr)Winter: Q = 0.77 10,000 (cfm) 60 (min/hr) 0.018 (Btu/ft3-F) (72 - 30) (F) = 349,272 (Btu/hr)

From the performance data given above, the additional pressure drop is 0.37 in H20. Thus, the additional fan electrical power, FP, is:

FP = 2 (fans) x 10,000 (cfm) x 0.37 (inH20) / [6,356 (cfm-inH20/hp) x 0.70 x 0.90] x 0.75 kW/hpFP = 1.39 kW

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