CO 2 -Based DCV Using

Transcription

1 2006 American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. ( Published in ASHRAE Journal (Vol. 48, May 2006). For personal use only. Additional distribution in either paper or digital form is not permitted without ASHRAE s permission. Demand-Controlled Ventilation CO 2 -Based DCV Using By Steven T. Taylor, P.E., Fellow ASHRAE Control of carbon dioxide (CO 2 ) concentration has been used for many years as an energy conservation measure in buildings to reduce outdoor air rates, and the energy required to condition the outdoor air, when spaces are not fully occupied. In fact, CO 2 -based demandcontrolled ventilation (CO 2 -DCV) is required for most densely occupied spaces by energy conservation standards such as ANSI/ASHRAE/IESNA Standard , Energy Standard for Buildings Except Low-Rise Residential Buildings, and California s Title However, revisions to the way ventilation rates are calculated in ANSI/ASHRAE Standard , Ventilation for Acceptable Indoor Air Quality, necessitate a change to the way CO 2 concentration is used in DCV control systems. This article summarizes how to use CO 2 -DCV with Standard from a theoretical standpoint and discusses in detail one CO 2 -DCV approach for single zone systems. Additional details on CO 2 -DCV, including applying DCV to multiple zone systems, are outlined in Appendix A of the new Standard User s Manual. Standard 62.1 Ventilation Rates To understand how to use CO 2 -DCV, it is first necessary to understand how ventilation rates are calculated in Standard 62.1 using the prescriptive Ventilation Rate Procedure. In previous versions of the standard, ventilation rates generally were determined by multiplying the number of people in the space times an outdoor airflow rate per person that varied by occupancy type. In the latest version of the standard, rates are calculated by summing two components, one intended to dilute the contaminants generated by occupants (bioeffluents) and their activities, and one intended to dilute contaminants emitted from building About the Author Steven T. Taylor, P.E., is a principal at Taylor Engineering, a consulting engineering fi rm in Alameda, Calif. He is a past chair of SSPC 62.1, Ventilation for Acceptable Indoor Air Quality. M a y A S H R A E J o u r n a l 6 7

2 ...revisions to the way ventilation rates are calculated in ANSI/ASHRAE Standard , Ventilation for Acceptable Indoor Air Quality, necessitate a change to the way CO 2 concentration is used in DCV control systems. materials, furnishings, and other non-occupant related sources. Mathematically, this is expressed in Equation 1: V bz = Rp P z + R a A z (1) where V bz is the minimum breathing (occupied) zone outdoor air rate, R p is the occupant ventilation rate component, P z is the number of occupants, R a is the building ventilation rate component, and A z is the occupiable floor area of the space. R p and R a are determined from Table 6-1 in Standard 62.1 based on occupancy type. The amount of outdoor air required at the supply air diffuser serving the space must be adjusted to account for how effectively outdoor air is delivered from the diffuser into the breathing zone. This is expressed in Equation 2: V oz = V bz E z (2) where V oz is the outdoor air rate required to be supplied to the zone and E z is the zone air distribution effectiveness, which is determined from Table 6-2 in Standard 62.1 showing values of E z for various supply air configurations. For single zone systems, the outdoor air rate required at system air intake, V ot, is equal to the zone outdoor air rate, so: V ot = V oz = V bz E z (3) For methods to calculate V ot for multiple-zone systems, consult the standard and the 62.1 User s Manual. This article is limited to single zone systems. CO 2 DCV Fundamentals CO 2 is a bioeffluent generated by people at a rate determined by their size, age, fitness, and activity level. At the same time people are generating CO 2, they also are producing odorous bioeffluents. These odorous bioeffluents are generated proportionally to the rate of CO 2 production, 2 although diet and personal hygiene also play a role. Nevertheless, CO 2 concentration is a fairly dependable indicator of the concentration of the odorous bioeffluents that the occupant component of the ventilation rate in Equation 1 attempts to control. Hence, we can use CO 2 concentration to dynamically adjust the ventilation rate, reducing outdoor air intake rates when zones are not occupied at their design occupancy. This CO 2 based demand-controlled ventilation is specifically allowed by Standard Figure 1 shows airflow rates and CO 2 concentrations for a single-zone ventilation system, where: V = the primary supply flow rate to the zone V pz ot = the outdoor air rate at the air handler (V ot is equal to the design outdoor air rate V ot when the zone is at full occupancy) v = the zone volume C s = the concentration of CO 2 in the supply air C R = the concentration of CO 2 in the room at breathing C RA level = the concentration of CO 2 in the return air = the generation rate of CO 2 in the zone The zone is assumed to be pressurized and thus exfiltrating the outdoor air supplied. The exfiltrated air is assumed to have CO 2 concentration equal to the room air concentration since space leakage is likely to occur at windows and doors that are located within the breathing zone. Using a control volume around the room and balancing CO 2 : C r + V pz C s (V V pz V ot ) C RA V ot C R = v t (4) Assuming steady-state (this assumption is justified below), the equation simplifies to: = V pz (CRA C s ) + V ot (C C R C RA ) (5) 6 8 A S H R A E J o u r n a l a s h r a e. o r g M a y

3 At the air handler, mass balance equations yield: V ot ( C RA C OA ) = V pz (C ra C s ) (6) Combining these two equations results in: = V ot (C C R C OA ) (7) From Equation 3 for single zone systems, the system outdoor air rate can be converted to the breathing zone outdoor air rate: V bz V ot = Ez E z where V bz is the breathing zone ventilation rate calculated from Equation 1 when the number of occupants is below the design occupancy. V ot and V bz in Equation 8 become design values V ot and V bz when the occupancy is at design conditions. If people are the only sources of CO 2 in the zone, then the source strength of CO 2 is = k m P z (9) where k is the generation rate of CO2 and m is the activity level of the people in the zone (in met). Generation rate k (8) averages about cfm/met/person over the general adult population. 3 Table 1 shows typical met levels for a variety of activities. Combining the last three equations and Equation 1, the outdoor air intake rate can be determined as a function of CO 2 concentration as: R aaz V ot = Rp R (C C R C OA ) E z km (10) With typical adult CO 2 generation rates and the units of concentration in parts per million, this equation can be written: R aaz V ot = Rp R (C C R C OA ) E z 8400m (11) Equation 11 can be combined with Equations 1 through 3 and used to determine the CO 2 concentration that will occur in the breathing zone at design occupancy and minimum outdoor air rate under steady-state conditions: Advertisement formerly in this space. 7 0 A S H R A E J o u r n a l a s h r a e. o r g M a y

4 R p P z + R aaz R aaz V ot = = E z R p (C C R C OA ) E z 8400m Solving for room CO 2 concentration we get: (12) C R = C OA + R p E z m R aaz P z (13) Table 2 shows steady-state CO 2 concentrations for several common occupancy types calculated using Equation 13 assuming default occupant density, estimates of activity levels, zone air distribution effectiveness E z equal to 1.0, and an ambient CO 2 concentration of 400 ppm. For offices, the steady-state CO 2 concentration (990 ppm) that results from Standard 62.1 ventilation requirements is about the same as the concentration (904 ppm) that results using ventilation requirements from the 1989 to 2001 versions of Standard 62 (20 cfm/person [10 L/s per person]). Concentrations for the other occupancy types listed, which are all densely occupied zones, are significantly higher than those resulting from prior Standard 62 ventilation rates. This is due to the less conservative code intended philosophy used to determined ventilation rates in Standard (For a more detailed rationale for the revised ventilation rates, see the 62.1 User s Manual.) Figure 1: Single zone airflows and CO 2 concentrations. Assumption of Steady-State Conditions Both Equation 11 and Equation 1, from which it was derived, are based on an assumption of steady-state conditions. In non-steady-state conditions, typical of most realworld applications, CO 2 concentration will generally lag behind changes in the actual number of occupants in the zone and changes in ventilation airfl ow rates. However, using Equation 11 to control outdoor air rates is still valid because the rate of generation of CO 2 by occupants should be nearly proportional to the rate of bioeffl uent generation; both are generated at a rate proportional to the number of people and their activity level. It is bioeffl uent (odor) concentration we are trying to control, and if the source strengths of CO 2 and bioeffl uents are proportional, CO 2 concentration may be used as an indicator of bioeffl uent concentration. Thus the steady-state assumption in Equation 11 is made not because the actual system is at steady-state but because the ventilation rate equation, Equation 1, is based on steady-state conditions. This steady-state relationship is simply being used to establish the relationship between CO 2 (odor) concentration and airfl ow setpoint in Equation 11. Therefore, while the rate of air supplied using Equation 11 will not exactly track the source strength of bioeffl uents due to transient effects, it should maintain an acceptable bioeffl uent concentration. To help understand this argument better, consider a cooling system analogy with variables of cooling load (analogous to bioeffl uent source strength), cooling capacity (ventilation rate), thermal mass (space volume), and temperature (CO 2 concentration). A thermal analysis of a space would result in a differential equation analogous to Equation 4. Temperature is used as an indicator of thermal comfort just as CO 2 is used as an indicator of bioeffl uent concentration, and we know from chamber studies (and common practice) approximately what temperature setpoint range to maintain to maintain a sense of comfort for most occupants. If we control cooling capacity (ventilation rate) to maintain temperature (CO 2 concentration) at the desired setpoint, then we have achieved our goal without any need to calculate the cooling load (bioeffl uent source strength) using differential equations. This is not to say that the control logic should not look at the time-rate-of-change of the controlled variable (temperature, CO 2 concentration) to improve the control system s ability to maintain the desired setpoint; just as with temperature controls, CO 2 -DCV controls can use proportional or proportional plus integral logic depending on the system design. 7 2 A S H R A E J o u r n a l a s h r a e. o r g M a y

5 Activity met Seated, Quiet 1.0 Reading and Writing, Seated 1.0 Typing 1.1 Filing, Seated 1.2 Filing, Standing 1.4 Walking at 0.89 m/s 2.0 House Cleaning Exercise Table 1: Typical met levels for various activities. 6 Occupancy Activity Steady State CO 2 Category Level Concentration Classrooms (age 9 plus) 1.0 met 1025 ppm Restaurant Dining Rooms 1.4 met 1570 ppm Conference/Meeting 1.0 met 1755 ppm Lobbies/Prefunction 1.5 met 1725 ppm Offi ce Space 1.2 met 990 ppm Sales 1.5 met 1210 ppm Advertisement formerly in this space. Table 2: Steady-state CO 2 concentrations at 400 ppm ambient. Dynamic Reset of Outdoor Air IntakeConstant Volume Systems With Airflow Measurement For single zone systems with outdoor airflow measuring and control devices, Equation 11 can be used to dynamically reset outdoor air intake rate, V ot. The control system, such as a direct digital control system, would have to be capable of using Equation 11 to dynamically calculate the required outdoor air rate setpoint and then modulate dampers (or provide some other means) to adjust the outdoor air rate to the new setpoint. The variables in Equation 11 can be determined from Table 3. Dynamic Reset of Outdoor Air IntakeConstant Volume Systems Without Airflow Measurement In the previous example, DCV was implemented by solving Equation 11 dynamically for the outdoor air intake airflow setpoint and adjusting dampers to maintain this setpoint. This requires a sophisticated control system (such as a direct digital control system) to make the setpoint calculation and an airflow measuring device (such as a pitot array) to measure outdoor airflow, neither of which are commonly found on typical single-zone HVAC systems such as packaged air conditioners. DCV could be implemented with prior versions of Standard 62 using only a CO 2 sensor and controller in the zone and a standard outdoor air economizer (mixing damper) assembly. Unfortunately, no precise way exists to do this with the current Standard due to the addition of the occupant and building components in Equation 1, which complicates the mathematics since the effective ventilation rate per person and the CO 2 concentration both vary with population. But a very simple approach to DCV is still possible: Calculate the required V ot at design occupancy using Equation 1 to Equation 3. Using the same equations, calculate the outdoor air rate with no occupants (P P z = 0). Call this value V at (the area building based component adjusted for distribution effectiveness). Using Equation 13, determine the steady-state CO 2 concentration when the zone is fully occupied and at its design outdoor airflow rate, V ot. See examples in Table 2. Call this value CO 2max. M a y A S H R A E J o u r n a l 7 3

6 Parameter Ventilation Parameters (R p, R a, and A z ) Determining Parameter Value These parameters are determined in accordance with Standard See the Standard 62.1 or the 62.1 User s Manual. Zone Air Distribution Effectiveness (E E z ) Activity Level Indoor Concentration E z is determined from Table 6-2 in Standard and may be adjusted dynamically in some cases. For instance, if the system uses overhead supply and return, E z could be 0.8 when the unit is heating, and 1.0 when the unit is cooling. Activity level is estimated from Table 2. The lower the estimate, the more conservative the resulting outdoor airfl ow rates will be. Indoor concentrations of CO 2 are measured with a CO 2 sensor located within the breathing zone of the zone (usually located adjacent to the zone thermostat). Typical sensors are of the infrared type with accuracies on the order of ±75 ppm. Many commonly used commercial sensors are factory calibrated and guaranteed not to require recalibration for as long as fi ve years. Outdoor CO 2 Concentration Outdoor CO 2 concentration is commonly determined in one of two ways: Conservatively assumed constant value typical of the area where the intake is located: outdoor air CO 2 concentration remains fairly constant unless the intake is located near roads where vehicle exhaust can raise levels during traffi c conditions. But even where such spikes can occur, assuming a typical background (non-traffi c affected) concentration (e.g., 400 ppm) will still be effective even if it results in somewhat higher outdoor air rates when spikes occur. Assuming a typical value is of course a very reliable approach since there are no additional sensors to get out of calibration. Dynamic measurement using a CO 2 sensor located outdoors, typically a duct-mounted type located in the outdoor air intake plenum or duct: although arguably the most accurate approach since it results in actual differential CO 2 concentration measurement, this approach can also be the least accurate and reliable due to sensor inaccuracy. For example, one sensor could read low while the other reads high, creating a doubly inaccurate differential reading. On the other hand, in areas where outdoor CO 2 concentration varies considerably (e.g. from 400 ppm to 600 ppm due to local automobile traffi c), using an outdoor CO 2 sensor can improve energy savings even with sensor inaccuracy. The accuracy of differential CO 2 concentration measurement can be improved by using a single CO 2 sensor with a sampling pump, sequenced valves, and tubes piped to the zone and to the outdoorsbut fi rst costs will be higher. Table 3: Determining parameters used in CO 2 DCV equations. Provide a CO 2 sensor/controller that is adjusted to send a maximum output signal when the room CO 2 is at the CO 2max and a minimum output signal when the room CO 2 is at the ambient conditions (e.g., 400 ppm). Adjust the outdoor air damper so that at the maximum controller output signal the system delivers design outdoor air rate V ot. This is generally done by or in conjunction with the test and balance contractor. Adjust the outdoor air damper so that at the minimum controller output signal the system delivers the building area outdoor air rate V at. Again, this is generally done by or in conjunction with the test and balance contractor. In this way, the minimum outdoor air intake rate varies from the building area rate (V V at ) to the design rate ( V ot ) as the CO2 concentration varies from ambient to the steady-state maximum rate Figure 2: CO 2 DCV for packaged air-conditioning unit. 7 4 A S H R A E J o u r n a l a s h r a e. o r g M a y

7 CO 2max. Note that this control varies the minimum outdoor air rate; the economizer controller can override this DCV minimum rate to provide additional outdoor air to reduce cooling energy usage when weather and load conditions are favorable. The design is shown in Figure 2 for a packaged single zone AC unit. A CO 2 sensor provides an analog signal proportional to CO 2 concentration. It is wired to a signal converter (transducer) that scales the output and converts it to be compatible with the damper actuator. For instance, the damper actuator may operate using a variable resistance signal with a potentiometer to maintain a minimum damper position when the AC unit is on. The signal converter replaces the potentiometerthe CO 2 DCV controls will maintain minimum outdoor air rates instead. The signal converter is adjusted to send a resistance that provides V at (as measured at the AC unit outdoor air intake by the test & balance contractor) when the CO 2 sensor output corresponds to ambient CO 2 concentration and to send a resistance that provides V ot when the CO 2 sensor output corresponds to CO 2max. Implementing CO 2 -DCV under previous versions of Standard 62 would require essentially the same components except that a controller (e.g., proportional, two-position) would also be required to maintain space CO 2 at the desired setpoint. The controller must be tuned in the field to ensure stable control. The cost of the controller and controller tuning is not required with the latest version of Standard 62.1, so in a sense CO 2 -DCV may be simpler and less expensive to implement than in the past. References 1. California Code of Regulations Title 24, Part 6, California s Energy Efficiency Standards for Residential and Nonresidential Buildings. California Energy Commission. 2. Rasmussen, C., et al The influence of human activity on ventilation requirements for the control of body odor. Proceedings of the CLIMA 2000 World Congress on Heating, Ventilating and Air- Conditioning 4: ASHRAE HandbookFundamentals. Chapter 8. Advertisement formerly in this space. Conclusions CO 2 -based demand-controlled ventilation has been a commonly used energy conservation strategy for many years. However, changes to the way outdoor air rates are calculated in the latest version of Standard 62.1 requires that DCV control system designs be modified. This article shows that using CO 2 -DCV in single zone systems based on the latest version of Standard 62.1 is no more complicated, and in fact may be simpler, to implement compared to CO 2 -DCV system designs based on prior versions of the standard. M a y A S H R A E J o u r n a l 7 5

8 2005, American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. ( org). Reprinted by permission from ASHRAE Journal, (Vol. 47, No. 5, May 2005). This article may not be copied nor distributed in either paper or digital form without ASHRAE s permission. Standard Designing Dual-Path, Multiple-Zone Systems By Dennis Stanke, Member ASHRAE ANSI/ASHRAE Standard , Ventilation for Acceptable Indoor Air Quality, 1 prescribes new minimum breathing zone ventilation rates and a new calculation procedure to fi nd the minimum intake airfl ow for different ventilation systems. Previous articles discussed the new ventilation design requirements for single-zone, 100% outdoor air, 2 changeover-bypass VAV, 3 and single-path VAV 4 systems. The detailed, step-by-step examples in these earlier articles proved to be a twoedged sword. Many readers were thankful for the calculation details. Others were daunted by the number of steps and apparently concluded that the calculations are too complex for ventilation system designers. However, in this author s opinion, Standard spells out more clearly what must be calculated and neither lengthens nor complicates the procedure substantially, compared to previous versions (since 1989). Although some readers may find it offputting, this article again includes design details to aid learningthis time, for an important set of ventilation systems wherein one or more zones receive ventilation air via two separate paths. Readers who persevere will find that Standard 62.1 offers designers the opportunity to account for secondary ventilation and thereby design energy-efficient systems that are less costly to install and operate. Dual-Path, Multiple-Zone Systems Many HVAC systems are configured as dual-path, multiple-zone, recirculating ventilation systems, which Appendix A in Standard describes as systems that provide all or part of their ventilation by recirculating air from other zones without directly mixing it with outdoor air. Dual-path systems (Figure 1) include a primary ventilation path (which supplies a mixture of first-pass outdoor air and centrally recirculated air) and a secondary ventilation path (which supplies only recirculated air). Secondary ventilation may be provided by either central recir- About the Author Dennis Stanke is a staff applications engineer with Trane, La Crosse, Wis. He is vice chair of Standing Standards Project Committee A S H R A E J o u r n a l a s h r a e. o r g M a y

9 culation of return air from all zones, or local recirculation of return air from one, several, or all zones. Dual-fan, dual-duct systems are dual-path ventilation systems with central secondary recirculation (Figure 1). One central fan supplies primary ventilation, and another central fan supplies secondary ventilation. Series fan-powered (SFP) VAV systems, on the other hand, are dual-path ventilation systems with local secondary recirculation (Figure 1). A central fan supplies primary ventilation and local fans in the VAV boxes supply secondary ventilation. Although any of these systems may be applied in various building types, the example in this article focuses on the design details for a series fan-powered VAV system applied in an office building. (Systems with central secondary recirculation are often easier to design, but systems with local secondary recirculation are more common and allow us to more completely illustrate the required calculations.) Dual-path ventilation systems offer a unique benefit. Ventilation air is delivered not only in the primary airstream from the central system but also in the local secondary recirculation air from local air sources, such as the return plenum. As the example will show, secondary recirculation greatly improves system ventilation efficiency, which reduces outdoor air intake flow compared to singlepath ventilation systems. Proper accounting for secondary recirculation in compliance with Standard requires that the designer use the dual-path system equation in Appendix A. Of course, the use of this equation (and the higher system ventilation efficiency that results) as an alternative compliance approach is entirely voluntary. Dual-path systems designed using simpler approaches, such as the Table 6-3 defaults or the single-path system equation in Appendix A (similar to the multiple-space system equation in Standard 62 since 1989), also comply. While these approaches entail simpler calculations, they do not result in the higher system efficiency and lower intake airflow that result from proper accounting of secondary recirculation. The series fan-powered VAV system (Figure 2) in our example includes a central air-handling unit (with a modulating outdoor air damper), a variable-volume primary air fan, series fan-powered VAV boxes (with hot water reheat in perimeter zones), and a central relief fan to control building pressure. Zone temperature setpoint is maintained by adjusting the fraction of primary airflow to the VAV box. Plenum airflow increases as primary airflow decreases (and vice versa), so each series fan provides constantvolume, variable-temperature discharge airflow to its zone. Zone Calculations Our example office building (Figure 3) includes eight HVAC zones, each with a thermostat controlling one or more VAV boxes. Each VAV zone in this case is also a separate ventilation zone, which Section 3 defines as one occupied space or several occupied spaces with similar occupancy category, occupant density, zone air-distribution effectiveness, and zone primary airflow per unit area. As shown in earlier articles, design calculations usually begin with zone ventilation requirements and proceed to system intake airflow requirements. Following the prescribed steps for zone ventilation calculations in Section 6.2.2, we first find each zone s outdoor airfl ow (V V oz ) (Table 1): 1. Use Equation 6-1 to find the minimum required breathing zone outdoor airflow (V bz = R p P z + R a A z ). 2. Look up zone air-distribution effectiveness (E E z ) in Table 6-2 based on the zone air-distribution configuration. 3. Use Equation 6-2 to find the minimum required zone outdoor airflow (V V oz = V bz /E/ z ) for each zone. These calculations (detailed in a previous article 4 ) are straightforward, once the zones are identified and a design population level (see inset) is established for each zone. Series Fan-Powered System Calculations Similar to previous versions, Standard acknowledges that multiple-zone recirculating ventilation systems deliver excess outdoor airflow to many zones, but that recirculation recovers some of that excess outdoor air. When the primary airstream contains sufficient outdoor air to properly ventilate the critical zone,* the same primary air overventilates all other zones to some degree. When recirculated, the unused outdoor air from overventilated zones reduces the required intake airflow. However, unused outdoor air that leaves the system (in relief air, for instance) without diluting contaminants reduces system ventilation efficiency.** The standard defines two design approaches to find and correct for system ventilation efficiency (E v ). The default approach uses prescribed values for E v (listed in Table 6-3 but not shown here), which depend upon the required fraction of outdoor air in the primary air supplied to the critical zone. The calculated approach determines E v for the system using equations (in Appendix A). A previous article 4 applied each approach to the design of a multiple-zone, single-path ventilation system. Either approach may be used to establish E v for dual-path systems. However, the default efficiency values in Table 6-3 are based only on primary recirculation airflow to each zone. * The critical zone, in terms of system ventilation, requires the highest fraction of outdoor air in the primary airstream; that is, it results in the lowest zone ventilation effi ciency (E vz ). ** To avoid multiple-zone system ineffi ciency, some designers advocate dedicated/100% outdoor air systems, which typically deliver constant volume outdoor airfl ow directly to each ventilation zone. While this approach simplifi es ventilation calculations, the design population must be assumed to be present in every zone. With no correction for system occupant diversity, many zones receive excess outdoor air during normal operation. Total excess outdoor air may be less than in recirculating systems, but outdoor air intake fl ow actually may be higher in some cases because no opportunity exists to recover and recirculate unused outdoor air. M a y A S H R A E J o u r n a l 2 1

10 Relief Fan Dual-Fan, Dual-Duct System RA Series Fan-Powered VAV System Local Secondary Recirculation Path RA Central Secondary Recirculation Path htg Relief Fan Supply Fan Central Primary Recirculation Path OA Supply Fan clg Dual-Duct VAV Box OA Central Primary Recirculation Path Series Fan- Powered VAV Box Ventilation Zone Ventilation Zone Ventilation Zone Ventilation Zone Figure 1: Dual-path, multiple-zone, recirculating ventilation systems. Since the default approach takes no credit for ventilation contributed by secondary recirculation, it significantly overventilates dual-path systems. With that in mind, this article covers only the calculated approach presented in Appendix A. It results in proper minimum ventilation for the critical zone and accounts for unused outdoor air leaving the system (E v < 1.0). It also takes credit for unused outdoor air recirculating from all other overventilated zones, both at the central air handler and at each local fan-powered VAV box. Relief Damper RELIEF INTAKE Cooling Design From our zone calculations, we now know how much outdoor airflow each zone must receive. The next step is to figure out the minimum system level outdoor air intake fl ow (V V ot ) that will deliver the required zone outdoor airflow. Initially we will use the cooling design condition (when system heat gainand, therefore, system primary airflowis greatest), even though the worst-case ventilation condition (when the required outdoor air intake flow is greatest) may occur at the heating design condition for some systems. In most dual-path ventilation systems, the highest outdoor air intake flow during mechanical cooling is likely to occur when most non-critical zones receive design primary airflow; the critical zone is at (or near) its minimum primary airflow setting; and the primary fan delivers design airflow. The procedure described here is similar to the one in Section for multiple-zone recirculating systems (Table 2), but we used the definitions and equations from Appendix A to calculate system ventilation efficiency (E E v ) rather than look up default values in Table 6.3. For this example, we used typical values for zone primary airflow ( V pz ) at cooling design conditions and selected series fans that deliver the required primary airflow when the damper in each VAV box is open-to-primary/closed-to-plenum airflow. (Note: Step 4 and Steps 9 through 12 may be applied to each zone as we did in our example, but experienced designers may P Flow-Measuring Intake (OA) Damper Relief Fan RRA Return Air Plenum Recirculating Damper choose to apply these steps only to zones deemed to be potentially critical, as described in Appendix A.) 4. For each zone (or selected zones), find the zone discharge outdoor air fraction (Z d ), according to the definition in Appendix A (Z = d V oz /V dz ), where V dz is the minimum expected zone discharge airfl ow. In most series fan-powered systems, diffuser airflow is the same at all operating conditions. If we only consider diffuser airflow (no transfer airflow), then minimum V dz equals design V dz. Note: For dual-path ventilation systems, the required fraction of outdoor air in the primary supply (Z Z p ) air delivered to the VAV box is not the same as the fraction needed in air delivered to the zone (Z d ) due to the influence of secondary recirculation. In our example, the south offices require 210 cfm of outdoor airflow (V V oz - clg ), of which about 11% must be outdoor air (Z = 210 d /1,900 = 0.111). 5. For the system, solve Equation 6-7 (D = P s /ΣPΣ P z ) to find occupant diversity using the expected peak system population (P P s ) and design zone population for all zones. This step is optional, but it reduces overventilation by accounting for variations in occupancy among all zones. Estimating population (both zone and system) is key to the design process because of 2 2 A S H R A E J o u r n a l a s h r a e. o r g M a y MA Mixed Air Filters Coils Plenum Supply Fan Figure 2: Series fan-powered variable-air-volume system. H C SA RA (Return Plenum) P Interior Zones T P VAV Box T Perimeter Zones Local Exhaust Fan VAV Box With Reheat Although it is commonly considered as only supply diffuser airfl ow, discharge airfl ow may include any controlled airstream that discharges into the ventilation zone, such as transfer air from the return plenum or from adjacent zones.

11 outdoor air fraction s (or perhaps more accurately, the used outdoor air fraction) according to the definition in Appendix A ( = s V ou /V V ps ). In our example, = s 2,800/18,600 = 0.15 at the cooling deits impact on system ventilation requirements. We estimated a maximum system population of 164 people in our example, so D = 164/224 = For the system, find the uncorrected outdoor air intake flow using Equation 6-6 ( V ou = D Σ(RR p P z ) + Σ(R a A z )). This value represents the rate at which outdoor air (found in both first-pass intake air and unused recirculated air) is used up in the process of diluting indoor North contaminants generated within the system. (In fact, a more appropriate name Offi ces for it may be used outdoor air rate.) Any outdoor air introduced in excess of this value is unused outdoor air that helps provide dilution ventilation if recirculated and reduces system ventilation Offi ces West efficiency if exhausted. Our example system needs at least V ou = 2,800 cfm of outdoor air in the breathing zones for proper dilution. 7. For the system, establish the system South Conf. Room Warehouse Northeast Interior Offi ces Southwest Interior Offi ces primary airfl ow (V V ps = ΣV V pz ) at the cooling design condition. This equation uses the sum of the zone primary airflows at the condition analyzed, not the sum-of-peak zone primary airflows. At the cooling design condition, many zones need peak primary airflow while others need less than peak. That s because the sun moves and not all zone sensible loads peak simultaneously. It s reasonable to simply use the maximum primary fan airflow for V ps at cooling design. Usually, the central primary VAV fan is selected to deliver block airflow, rather than sum-of-peak airflow, based on a system load diversity factor (LDF = system block load divided by sum-of-peak zone load). In our example office, we used a system load diversity factor of 0.70 based on load calculations, so the central fan delivers V ps = 0.70 North 26,600 = 18,600 cfm at cooling design. Conf. (During cooling operation, most dualpath systems exhibit their lowest system Room ventilation eff iciency when system primary airflow is high and critical zone primary airflow is low. Since all non-critical zones are overventilated to a greater East Offi ces extent at this condition, excess outdoor air recirculates, but a large portion of it is lost in the relief air leaving the system.) 8. For the system, find the average South Offi ces Figure 3: Multiple-zone office building. The defi nition in Appendix A states that s s represents the fraction of outdoor air intake fl ow in the system primary airfl ow. That wording does not strictly match the equation for s, which shows the outdoor air usage rate as a fraction of system primary airfl ow. However, the equation is correct according to the derivation of these equations. Perhaps the defi nition in the standard can be clarifi ed quickly via the continuous maintenance process so that it more closely matches the defi ning equation. Supplemental Recirculating Fan Conference rooms are notorious for fouling up multiple-zone system ventilation calculations, especially in VAV systems. Why? Conference rooms have high design population density, so the fraction of outdoor air needed in the air supplied to the zone is also high, especially when cooling load is low. Typically, in single-path systems, the minimum primary airfl ow must be set to a very high value to limit Z p (= V oz /V V pz ) and ensure proper ventilation without requiring 100% outdoor air at the intake. In dual-path systems, a high minimum primary setting (which increases E p = V pz /V dz ), may not be necessary, since the secondary recirculation path can keep zone ventilation effi ciency (E vz ) high, even with low primary recirculation to a zone. In either system, designers may want to include intermittent (or continuous) supplemental recirculating fans, which draw air from the return plenum and discharge it into the conference room. How does this improve system ventilation effi ciency? Fan operation adds another local recirculation path to the conference room. The air-conditioning diffusers provide some outdoor air (via primary and perhaps secondary recirculation) while the supplemental fan diffusers add a local recirculation path from the return plenum. To solve the equations in Appendix A, discharge airfl ow into the zone (V V includes diffuser airfl ow and any supplemental dz ) airflow from the return plenum. Increasing V dz changes several of the parameters, but most importantly, it decreases Z (i.e., d V oz /V V signifi cantly. Zone ventilation effi ciency dz ) E vz rises. If it s the critical zone, system ventilation effi ciency E v rises too, decreasing outdoor air intake fl ow, V ot. Supplemental recirculating fans work well and the equations in Appendix A can handle them, but such fans may or may not be allowed in all jurisdictions. Designers should check with code authorities before pursuing this design approach. (The equations probably don t apply for zones with supplemental return fans, which result in increased transfer air from adjacent zones. This actually introduces a tertiary ventilation path so it s likely to improve ventilation, but it s too diffi cult to predict the fraction of unused outdoor air transferred from adjacent zones.) 2 4 A S H R A E J o u r n a l a s h r a e. o r g M a y

12 Cooling Heating Procedural Step Variable R p P z R a A z V bz E z V oz E z * V oz Ventilation Zone cfm/p P cfm/ft 2 ft 2 cfm cfm cfm South Offi ces , West Offi ces , South Conference Room , East Offi ces , Southwest Interior Offi ces , Northeast Interior Offi ces , North Offi ces , North Conference Room , * For zones with reheat at the VAV box, discharge air temperature changes from cool to warm, so Ez changes from 1.0 to 0.8 at design heating conditions. Table 1: Zone ventilation calculations. sign conditions, that is, 15% of the outdoor air in the primary airstream is used to dilute contaminants. 9. For each zone (or selected zones), find the lowest fraction (E p ) of primary air in all air delivered to the zone according to the definition in Appendix A (E E p = V pz /V dz ). Since proper ventilation airflow must be delivered to all zones at all load conditions, the lowest E p in a zone is likely to result in the lowest zone ventilation efficiency (E vz ) in that zone. The lowest possible E p for a zone occurs when zone primary airflow (V V pz ) is at the minimum primary airflow setting. For our example, we arbitrarily assumed a 25% minimum primary airflow, so E p = V pz /V dz = 0.25 V dz /V dz = 0.25 for each zone. (This does not mean that all zones are at 25% primary airflow at the cooling design condition. Rather, it means that the lowest possible E p value for a zone is 0.25.) Note: Although dual-path systems often can be ventilated properly with much lower minimum primary airflow settings, we arbitrarily used 25% minimums for our example calculations to allow easy comparison with the ventilation systems considered in earlier articles. 10. For each zone (or selected zones), establish the ventilation quality of locally recirculated return air (secondary air) using the return air mixing effi ciency (E r ), defined in Appendix A. This efficiency varies from 0.0 to 1.0, depending on the location of the zone secondary air source with respect to the central system return air. When zone secondary air has the same ventilation qualitythat is, the same fraction of unused outdoor airas the central return air, then E r = 1.0. (In dual-fan, dual-duct VAV systems, the return air supplied by the heating fan is the same as that recirculated by the main supply fan, so E r = 1.0.) When zone secondary air has the same ventilation quality as the zone served, then E r = 0.0. (In systems where return air from the zone is ducted to the inlet of the fan-powered VAV box serving the zone, then E r = 0.0.) In actual systems with fan-powered boxes, E r is closer to 1.0 if the VAV box is located near the central return air inlet, and closer to 0.0 if located directly over the return grille of the zone served. In our example, the fan-powered boxes are located centrally, so we assumed E r = 0.8 for all zones. (The standard does not require a calculation method for E r, so the values used must be based on designer judgment. However, a current ASHRAE research project [RP-1276] will measure E r values in a real building with fan-powered VAV boxes in various locations. This data should provide the information needed to help designers establish E r values in the future.) 11. For each zone (or selected zones), according to the definitions in Appendix A, find: Design Population Standard allows the designer to use either the highest expected zone population or the average zone population when calculating breathing zone ventilation airfl ow using Equation 6-1. This design population option introduces design fl exibility, but it also introduces a potential source of inconsistency among designers. Peak population is usually available, but if used, it increases zone ventilation requirements. Average population, on the other hand, decreases zone ventilation requirements, but it s not always easy to fi nd. Or, more accurately, it s not always easy to predict the zone population profi le. For some occupancy categoriesclassrooms or perhaps churches and theaters, for examplezone population profi le may be readily predictable based on a time-of-day schedule. However, for other categoriesoffi ces, conference rooms, retail areas and perhaps theatersoccupant level fl uctuations over the averaging period T (found using Equation 6-9) are not easily established. In these areas, one designer s profi le estimate may be considerably different than that of another designer. So, the defi nition for design population P is fl exible z but can lead to inconsistency among designers. Time will tell if a stricter, less fl exible defi nition is needed. 2 6 A S H R A E J o u r n a l a s h r a e. o r g M a y

13 From Above Procedural Step a 11b 11c Ventilation Box Type V pz (clg V fan Zone design) V dz cfm cfm cfm cfm V pz min * V oz clg Z d E p E r F a F b F c E vz South Offi ces SFP Reheat 1,900 1,900 1, West Offi ces SFP Reheat 2,000 2,000 2, South Conference Room SFP Reheat 3,300 3,300 3, East Offi ces SFP Reheat 2,000 2,000 2, Southwest Interior Offi ces SFP VAV 7,000 7,000 7,000 1, Northeast Interior Offi ces SFP VAV 7,000 7,000 7,000 1, North Offi ces SFP VAV 1,600 1,600 1, North Conference Room SFP VAV 1,800 1,800 1, System V V pz 26,600 (Step 5) D = P s / PP z (Step 6) Vou (Step 7) Vps V (Step 8) s (Step 13) Ev (Step 14) Vot 3,120 * Set at 25% of design primary airfl ow. Ventilation critical zone. P s (system population) = 164 people, and PP z (sum of zone peak population) = 224 people. Table 2: System ventilation calculations at cooling design ,800 18, a. The fraction of supply air to the zone from locations outside the zone ( F a = E p + (1 E p ) E r ). For our example, F a = (1 0.25) 0.8 = 0.85 for all zones. b. The fraction of supply air to the zone from the primary airstream (F = b E p ). For our example, F b = 0.25 for all zones. c. The fraction of outdoor air to the zone from outside the zone (F( = 1 (1 c E z ) (1 E ) (1 r E p )). For our example, F = 1.0 for all zones when cooling, because c E z =1.0 for all zones. 12. For each zone (or selected zones), find zone ventilation effectiveness using Equation A-2 [E vz = F a + s F Z b d F c ) / F ] for dual-path systems. In our example, E a vz = ( )/0.85 = for the south offices. 13. For the system, find system ventilation effi ciency using Equation A-3 (E = minimum E v vz ), the lowest zone ventilation effi ciency among the zones served by the system. In our example, we identify the north offices as the ventilation critical zone at the cooling design condition. It has the lowest zone ventilation efficiency, so E v = (We used three decimal places in our example because our numbers result in only slight differences in E vz values.) 14. Finally, find outdoor air intake fl ow for the system by solving Equation 6-8 (V ot = V ou /E/ v in the body of the standard). In our example, V ot = 2,800/0.897 = 3,120 cfm at the design cooling condition. In other words, about 17% of the 18,600 cfm primary airflow needed at the cooling design condition must be first-pass outdoor air. When compared to the 4,310 cfm of first-pass outdoor air needed for a single-path system applied to the same building, 4 it s clear that dual-path systems can significantly reduce the required outdoor air intake flow. Heating Design We repeated the previous calculation steps (Table 3) for the heating design conditions, when system heat loss to outdoors is greatest, to see if worst-case ventilation occurs when primary airflow is very low. (We also could use the minimum outdoor air intake flow at the heating design condition to save energy by establishing separate minimum intake flow settings for summer and winter operation.) Step 4 results in a minimum discharge outdoor air fraction (Z d htg ) that s slightly higher than Z d-clg in the perimeter zones with reheat0.138 vs in the south offices, for example. (Zone air-distribution effectiveness is somewhat lower when delivering heat from overhead diffusers, which makes V oz - htg somewhat higher than V oz - clg.) Since the zone and system populations don t change, Steps 5 and 6 yield the same values for both D and V ou. Step 7 is a little tricky: What is system primary airflow ( V ps ) at the heating design condition? That depends on the system and the 2 8 A S H R A E J o u r n a l a s h r a e. o r g M a y

14 Equations and Variables from Addendum 62n [6-1] V bz = R p z a A z [6-2] V oz = V bz z/ z [6-3] V ot = V oz single-zone systems [6-4] V ot = V V oz 100% outdoor-air systems [6-5] Z p = V oz /V pz [6-6] ou R allzones p z a A z = D allzones V bzp + allzones V bza [6-7] D = P s / allzones P z [6-8] V ot = V ou /Ev multiple-zone recirculating systems [6-9a] T = 3v/V bz IP version [6-9b] T = 50v/V bz SI version where A z is zone floor area, the net occupiable floor area of the zone, ft² (m²) D is occupant diversity, the ratio of system population to the sum of zone populations E v is ventilation efficiency of the system E z is air-distribution effectiveness within the zone P s is system population, the maximum simultaneous number of occupants in the area served by the ventilation system P is zone population, the largest expected number of people z to occupy the ventilation zone during typical usage (See caveats in Addendum 62n Section ) R a is area outdoor air rate, the required airflow per unit area of the ventilation zone determined from Addendum 62n Table 6.1, cfm/ft² (L/s m²) R p is people outdoor air rate, the required airflow per person determined from Addendum 62n Table 6.1, in cfm/person (L/s person) T is averaging time period, minutes v is ventilation-zone volume, ft³ (m³) V bz is breathing-zone outdoor airflow, the outdoor airflow required in the breathing zone of the occupiable space(s) of the ventilation zone, cfm (L/s) V ot is outdoor air intake flow, adjusted for occupant diversity and corrected for ventilation efficiency, cfm (L/s) V ou is the uncorrected outdoor air intake flow, cfm (L/s) V oz is zone outdoor airflow, the outdoor airflow that must be provided to the zone by the supply-air-distribution system at design conditions, cfm (L/s) V pz is zone primary airflow, the primary airflow that the air handler delivers to the ventilation zone; includes both outdoor air and recirculated return air Z p is zone primary outdoor air fraction, the fraction of outdoor air in the primary airflow delivered to the ventilation zone for VAV systems, Z p for design purposes is based on the minimum expected primary airflow, V pzm. weather. For proper system ventilation, we want to find the highest volume of outdoor air needed. Does this occur when all zones receive minimum primary airflow (an easy condition to check)? Or when perimeter zones receive minimum primary airflow and interior zones receive more than minimum (a more difficult condition to check)? For our example, we first assumed that all zones receive minimum primary airflow (V V ps = 6,650 cfm). Step 8 results in average outdoor air fraction ( s ) of 0.42, compared to 0.15 at cooling design. In other words, the used outdoor air fraction increases when primary air decreases. Steps 9, 10 and 11 result in E p, E r, F a, and F b values that are identical to those at the cooling design condition. F, on c Multiple-Zone Systems Does the designer always need to calculate ventilation requirements at both heating and cooling design? Not necessarily. Some systemsseries-fan powered and dual-fan, dual-duct VAV systems with constant volume discharge, for instancealmost surely need less intake air as primary airfl ow drops (as it does at part-cooling load) than when primary airfl ow peaks (as it does at cooling design condition). The same is true for parallel fan-powered and dualfan, dual-duct VAV systems with VAV discharge airfl ow, since any zone in the heating mode has two ventilation the other hand, is lower in the reheat zones because zone airdistribution effectiveness (E z ) is lower when delivering warm air. Step 12 reveals that the south offices are now the critical zone, rather than the north offices, while Step 13 indicates that system ventilation efficiency (E v ) rises to Higher efficiency means that outdoor air intake flow (V V ot ; Step 14) drops to 2,900 cfm from the 3,120 cfm needed at the cooling design condition making cooling design the worst-case ventilation condition. This result is expected to be typical of dual-path recirculating systems because the system contains less excess outdoor air at reduced primary airflow. paths. On the other hand, single-path systems, such as single-duct VAV and constant volume reheat systems, with low zone air-distribution effectiveness during heating, may need higher outdoor air intake fl ow heating design conditions. Until we have more experience with these systems, or until more analytical research results tell us the answer, we can t draw absolute conclusions. Our advice to designers is to use an ASHRAE spreadsheet (you will fi nd one at www. ashrae.org) or develop your own spreadsheet. Once you have entered the equations, it s easy to experiment and draw your own conclusions for your ventilation system design. M a y A S H R A E J o u r n a l 2 9

15 From Above Procedural Step a 11b 11c Ventilation Box Type V pz (htg V fan Zone design) V dz cfm cfm cfm cfm cfm V pz min * V oz htg Z d E p E r F a F b F c E vz South Offi ces SFP Reheat 475 1,900 1, West Offi ces SFP Reheat 500 2,000 2, South Conference Room SFP Reheat 825 3,300 3, East Offi ces SFP Reheat 500 2,000 2, Southwest Interior Offi ces SFP VAV 1,750 7,000 7,000 1, Northeast Interior Offi ces SFP VAV 1,750 7,000 7,000 1, North Offi ces SFP VAV 400 1,600 1, North Conference Room SFP VAV 450 1,800 1, System Vpz V 6,650 (Step 5) D (Step 6) Vou (Step 7) Vps V (Step 8) s (Step 13) Ev (Step 14) Vot 2,900 * Set at 25% of design primary airfl ow. Ventilation critical zone. Table 3: System ventilation calculations at heating design ,800 6, What happens if our simple assumption about primary airflow (all zones at the minimum primary setting) at the heating design condition is too low? To find out, we recalculated system ventilation efficiency for our example with the perimeter zones at minimum primary airflow and the interior zones arbitrarily at 50% of design primary airflow. This is based on the assumption that these zones need more than minimum primary airflow (probably the norm in most systems) even when it s really cold outdoors. Without going into detail, we found that system primary airfl ow (V V ps ) rises to 11,000 cfm, system ventilation effi ciency (E ) drops to 0.917, and minimum outdoor air intake fl ow (V v V ot ) rises to 3,050 cfm. In this example system, the cooling design condition still requires the highest outdoor airflow. For an actual ventilation system, however, the designer should take care to base the design on worst-case intake airflow. A load analysis (for instance) can help determine the lowest primary airflow needed in each zone at all design conditions. Other Dual-Path Ventilation Systems We looked at ventilation system design for a series fanpowered VAV system in detail. What about other dual-path ventilation systems? Parallel Fan-Powered VAV The local (secondary) ventilation path in a parallel fanpowered VAV system only functions when the fans in the VAV boxes operate, which is during heating. The local ventilation path and the benefits of secondary recirculation disappear during cooling, when the local parallel fans are off. If we applied such a system to our example building, we d find that zone ventilation requirements don t change and that worstcase ventilation (highest required outdoor air intake flow) again occurs at the cooling design condition. At the heating design condition, system ventilation efficiency rises, and outdoor air intake flow is much less than that required at cooling design. Dual-Fan, Dual-Duct VAV Two central air handlersone that delivers a cool mixture of first-pass outdoor air and centrally recirculated return air (primary ventilation), and another that delivers warm, centrally recirculated return air (secondary ventilation)supply air to the dual-duct VAV boxes. The boxes can be controlled to deliver either constant-volume variable-temperature or variable-volume constant-temperature air to each zone. When configured with constant-volume variable-temperature VAV boxes, dual-fan, dual-duct VAV systems usually have the highest system ventilation efficiency among all multiple-zone recirculating systems. Applying this dual-fan, dual duct system to our example, again zone ventilation requirements do not change. Worst-case ventila- tion may occur at either the cooling design or heating design condition, depending on the zone minimum primary airflow settings. 3 0 A S H R A E J o u r n a l a s h r a e. o r g M a y

16 The key difference between dual-fan, dual duct systems with central secondary recirculation and series fan-powered systems with local secondary recirculation is that E (the fraction of average r system return air in the secondary air) always equals 1.0, because return air from all zones mixes at the central secondary recirculation fan. This means that system efficiency can be slightly higher and outdoor air intake flow can be slightly lower than would be the case when using a series fan-powered system. Of course, actual results may differ, depending on the control configuration (constant volume vs. variable volume) and minimum settings of the VAV boxes, but detailed design analysis is left for future articles. Summary Standard not only clarifies the calculation procedures that were always required by Standard 62, but it also adds an important new tool for designers. Earlier versions required that the designer use the multiple space equation to account for the inherent inefficiency of multiple-zone systems. However, that equation addressed only single-path ventilation systems and, therefore, did not reward dual-path systems for ventilating more efficiently. Standard , through the dual-path equations presented in Appendix A, incorporates a more generalized multiple-space equation, giving ventilation credit where it is due. This article demonstrated, in detail, how to calculate the worst-case ventilation for one type of dual-path ventilation system. We also noted that some types of dual-path systems ventilate more efficiently than others, and identified the tradeoffs that designers can consider when selecting and designing these systems to comply with Standard References 1. ANSI/ASHRAE Standard , Ventilation for Acceptable Indoor Air Quality. 2. Stanke, D Addendum 62n: Single-zone and dedicated-oa systems. ASHRAE Journal 46(10): Stanke, D Standard Addendum 62n: Ventilation for changeover-bypass VAV systems. ASHRAE Journal 46(11): Stanke, D Standard Addendum 62n: Single-path multiple-zone system design. ASHRAE Journal 47(1):28 35.

17 2005, American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. ( org). Reprinted by permission from ASHRAE Journal, (Vol. 47, No. 1, January 2005). This article may not be copied nor distributed in either paper or digital form without ASHRAE s permission. Standard Addendum 62n Single-Path Multiple-Zone System Design By Dennis Stanke, Member ASHRAE A NSI/ASHRAE Standard , Ventilation for Acceptable Indoor Air Quality, 1 as modifi ed by Addendum 62n, 2 prescribes new minimum breathing-zone ventilation rates and new calculation procedures to fi nd intake airfl ow for different ventilation systems. Previous articles 3,4 discussed the design of simple ventilation systems (singlezone, 100% outdoor-air, and changeover-bypass VAV) in compliance with Addendum 62n requirements. Here, we examine the design of a more complex set of ventilation systems, namely single-path, multiplezone recirculating systems. Although the Ventilation Rate Procedure in Standard 62 has required specific calculations (Equation 6-1) for multiple-zone systems since 1989, the calculation procedure was sketchy at best; consequently, it was widely misunderstood and largely ignored by designers. Addendum 62n includes a detailed calculation procedure for multiple-zone system design. Use of this procedure is expected to increase consistency among designers and reduce the tendency to design multiple-zone systemsespecially VAV systemsthat provide inadequate ventilation for some fully occupied zones. Addendum 62n also includes operational control options that can be used to modulate ventilation capacity as ventilation load and/or efficiency varies, but these options are left to a future article. The following discussion covers only design calculations. Many HVAC systems are configured as single-supply or single-path, multiple- zone, recirculating ventilation systems. For instance, constant-volume systems with terminal reheat, traditional constantvolume multizone systems, single-duct VAV systems, and single-fan dual-duct VAV systems all provide ventilation from a single source or path. (A single-fan, dual-duct system supplies air to each space using two different ducts, but the air in each duct contains the same fraction of outdoor air, because one fana single sourcedelivers the same air mixture to each duct.) Other systems have multiple ventilation paths, including dual-fan, dual-duct VAV systems and VAV systems with fan-powered or induction terminal units. Single-duct VAV systems with series fan-powered boxes are always dual-path ventilation systems, but those with parallel fan-powered boxes are single-path with the local fan off and dual-path with it on. Although any of these HVAC systems may be used in vari- About the Author Dennis Stanke is a staff applications engineer with Trane, La Crosse, Wis. He is vice chair of SSPC A S H R A E J o u r n a l a s h r a e. o r g J a n u a r y

18 ous building types, we narrow our discussion to a single-duct VAV system, with throttling VAV boxes for interior zones and reheat VAV boxes in perimeter zones, applied in an example office building. Demonstrating Compliance by Example Our example system (Figure 1) includes a central air handler, with a modulating outdoor-air damper that may be controlled as an economizer; a variable-volume supply fan to deliver primary air; cooling-only, throttling VAV boxes in the interior zones; throttling VAV boxes with electric reheat in the perimeter zones; a central return fan; and a central relief damper for building pressure control. Although we won t discuss system control details here, it s important that we share the same mental picture of the VAV system we re designing: Intake airflow is sensed and maintained by adjusting the intake damper position. (Often, the return- and outdoor-air dampers are linked such that closing the outdoor-air damper opens the return-air damper proportionately. Alternately, these dampers can be controlled separately to reduce fan energy while maintaining proper intake airflow, but this has no impact on ventilation requirements at design conditions.) Primary air temperature is sensed and maintained by sequentially adjusting the heating-coil control valve, economizer dampers, and coolingcoil control valve. Duct pressure is sensed and maintained at setpoint by adjusting the primary fan capacity (via fan speed, for instance, or inlet guide vane position). Zone temperature is sensed and maintained at the cooling setpoint by adjusting the setpoint for VAV-box primary airflow. VAV-box airflow is sensed and maintained at setpoint by adjusting the position of the VAV-box damper. For zones that need reheat, zone temperature is sensed and maintained at the heating setpoint by adjusting reheat capacity (electric reheat or a hot water valve) and, thereby, discharge air temperature. Return air plenum pressure (at the central air handler) is sensed and maintained by adjusting return fan capacity. Building pressure is sensed and maintained between set limits by adjusting the relief (central exhaust) damper position. Since multiple-zone systems provide the same primary air mixture to all zones, the fraction of outdoor air in the primary airstream must be sufficient to deliver the outdoor airflow needed by the critical zonethe zone needing the greatest fraction of outdoor air in its primary airstream. In the past, many designers simply added the zone outdoor airflow requirements and set the intake airflow to match this sum, which resulted in a very low outdoor-air fraction and many underventilated zones. Some designers went to the other extreme, finding the highest fraction of outdoor air needed by any zone in the system and setting the intake airflow to provide this fraction at all times. This approach considers only first-pass outdoor air, giving no credit for unused recirculated outdoor air, and results in a very high outdoor-air fraction and overventilation in all zones. Proper design in compliance with Addendum 62n calculation procedures strikes a balance between these extremes, appropriately accounting for both critical-zone needs and unused, recirculated outdoor air. Let s look at an example office building (Figure 2). We assumed that thermal comfort can be achieved using only eight VAV thermostats, with each thermostat controlling one or more VAV boxes. We considered each of these comfort zones (or HVAC zones per ASHRAE Standard ) as a separate ventilation zone. According to Addendum 62n, a ventilation zone is one occupied space or several occupied spaces with similar occupancy category, occupant density, zone air-distribution effectiveness, and zone primary airflow per unit area. Most (but not all) HVAC zones qualify as ventilation zones. The area and population for each zone in this example were selected to help illustrate the calculations rather than to reflect typical zone sizes or population densities. To comply with Addendum 62n, our design calculations begin by finding the ventilation needs at the zone level and conclude by determining the required intake airflow at the system level. Zone Ventilation Calculations Following the procedure under zone calculations in Section 6.2.1, we found zone outdoor airflow (V oz ) for each zone (Figure 3): 1. Referring to Addendum 62n, Table 6.1 (not shown), look up the prescribed minimum people outdoor-air rate (R p ) and the prescribed minimum building area outdoor-air rate (R a ). In our example office building, each zone needs 5 cfm/person and 0.06 cfm/ft². Using these values, along with the design zone population (P z ) and zone fl oor area (A z ), find the minimum breathingzone outdoor airfl ow by solving Equation 6-1 (V bz = R P + p z R a A z ). Either peak or average expected occupancy may be J a n u a r y A S H R A E J o u r n a l 2 9

19 used to establish P z ; we used peak population in all zones. (An earlier article 3 covered population-averaging calculations in detail. See for the most current version.) For our example, the west offices need V bz = ,000 = = 220 cfm for proper ventilation in the breathing zone. 2. Look up zone air-distribution effectiveness (E z ), based on the air-distribution configuration and the default values presented in Addendum 62n, Table 6.2 (not shown). All of our example zones use overhead diffusers and ceiling returns, and they all receive 55 F primary air, so E z = 1.0 when cooling. If the thermostat calls for heat in any of the perimeter zones, primary air is reheated and discharged at 95 F; so, E z = 0.8 when heating. 3. Find the minimum zone outdoor airflow by solving Equation 6-2 ( V oz = V bz / E ) for both cooling and heating operation. For example, the west offices need z V oz = 220/1.0 = 220 cfm at the diffusers when cooling, and V oz = 220/0.8 = 275 cfm when heating. Relief Damper Relief Intake Return Air Plenum RRA System Ventilation Calculations As in Standard , -1999, and -2001, Addendum 62n recognizes that multiple-zone recirculating systems must overventilate some zones to properly ventilate all zones. It also recognizes that unused outdoor air recirculated from overventilated zones reduces the required intake airflow, but that unused outdoor air that leaves the building (by exhaust or exfiltration) increases the required intake airflow. Proper accounting results in a ventilation credit for recirculated outdoor air and a ventilation debit for exhausted outdoor air. Addendum 62n makes this accounting straightforward by requiring a specific calculation procedure to determine the minimum outdoor-air intake fl ow based on the system ventilation effi ciency inherent in every multiple-zone recirculating system. Earlier versions of the standard required use of the multiple-space equation, Y = / (1 + Z), to find the fraction of intake air needed. This approach resulted in about the same intake airflow as Addendum 62n for single-path systems; but without a clear procedural explanation, the equation was widely misunderstood and largely ignored by designers. Recirculating Damper MA Flow-Measuring Mixed Intake (OA) Air Filters Coils Damper Plenum Variable-Speed Return Fan Figure 1: Variable air volume reheat system. Variable-Speed Supply Fan SA RA Interior Zones (Return Plenum) VAV Box Perimeter Zones Local Exhaust Fan VAV Box With Reheat Designs based on the 62n procedure result in proper ventilation for the critical zone at worst-case design conditions while allowing credit for good outdoor air that recirculates from all other overventilated zones. From the zone calculations that we completed earlier, we know how much outdoor airflow must reach the diffusers in each zone. Now, let s figure out the minimum required intake airflow for the system at design conditions. Before we start, we should recognize something that Addendum 62n implies but doesn t explain: The worst-case or highest required intake airfl ow may or may not occur at the design cooling condition (when system primary airflow is highest). In some cases, it may actually occur at the design heating condition (when zone primary airflow values are very low). With Averaging Zone Population for Ventilation System Design T v/ / In earlier versions of the standard, only intermittent occupancy zones (at peak population for three hours or less) could be designed for ventilation at the average population (but not less than one-half of the peak population). Now, any zone may be designed for average population. According to the short-term conditions section of Addendum 62n, the system must be designed to deliver the required outdoor airflow to each occupied breathing zone. However, if occupancy or intake airflow varies, the ventilation system design may be based on average conditions over a specifi c time period rather than on peak conditions. The averaging time T for a given zone is determined according to Equation 6-9 (T = 3 v V oz ) using zone volume and the breathing-zone outdoor airflow that would be needed at peak population. T equals three zone time constants, the time it takes for contaminant concentration to achieve a nearly steady-state value in response to a step change in contaminant source. When applied to population, this averaging approach replaces the population-averaging option for intermittent occupancy spaces, found in previous versions of the standard, Averaging time may be applied to make design adjustments when changing conditions in the zone can be predicted. For instance, if zone population fl uctuations are predictable, then the design breathing-zone outdoor airfl ow may be calculated based on the highest average population over any T-minute - period. 3 0 A S H R A E J o u r n a l a s h r a e. o r g J a n u a r y

20 this in mind, we ll need to check the required intake airflow at both design cooling and design heating because it s ultimately the worst-case outdoor-air intake flow that will establish the required capacities for the heating and cooling coils. For our example, we tried to use reasonable values for zone primary airfl ow (V pz ) at design cooling load. We arbitrarily set all minimum primary airflow settings (V pz -min) to 25% of design cooling airflow. We assumed that each reheat box enters reheat mode after its primary airflow decreases to the minimum setting and the zone temperature drops below the heating setpoint. Reheat operation continues until the zone temperature exceeds the heating setpoint. Case 1: Ventilation Calculations for De- West Offi ces North Offi ces Warehouse Northeast Interior Offi ces Southwest Interior Offi ces fault Cooling Design Building on our earlier zone-level calculations (Figure 3), we followed the step-by-step, multiple-zone recirculating systems procedure to find the minimum, system-level, outdoor-air intake fl ow (V V ot ) at the design cooling condition (Figure 4): 4. For each zone, find the zone primary outdoor-air fraction by solving Equation 6-5 (Z p = V oz /V pz ) using the zone outdoor airflow (V oz ) values for cooling from Step 3 and the minimum primary airflow setting. As an example, at minimum primary airflow, the south offices need Z p = 210/475 = 0.44 when delivering cool air. 5. Addendum 62n allows the designer to use a either default value for system ventilation effi ciency (E v ) using Table 6.3 (not shown) or a calculated value (found using equations in Appendix G). In this case, we used Table 6.3 and the highest zone primary outdoor-air fraction among the zones served ( max Z p = 0.50 for the north offices) to look up the corresponding default system ventilation effi ciency (E v ). From that value, we can interpolate to find E v = Find occupant diversity according to Equation 6-7 (D = P s / P z ) by using the expected peak system population (P s ) and the sum of design zone populations. For our example, we expect a maximum system population of 164 people, so D = 164/224 = Find the uncorrected outdoor-air intake fl ow for the system by solving Equation 6-6 (V ou = D (R( p P z ) + (R( a A z )). Without correcting for zone ventilation effectiveness and system ventilation efficiency, we find that the system needs V ou = 2,800 cfm of outdoor air at the breathing zones. 8. Finally, find outdoor-air intake North fl ow for the system by solving Equation Conf. 6-8 (V V ot = V ou/e / ). In our example, V v ot = Room 2,800/ 0.65 = 4,310 cfm at the design cooling condition. But, is this really the worst-case (highest volume) intake airflow? What happens at design heating conditions? East Offi ces Case 2: Ventilation Calculations for Default Heating Design Let s find the minimum system-level outdoor-air intake fl ow (V V ot ) for the design heating condition. The procedure is South Conf. Room South Offi ces the same one that was just described for default cooling design in Case 1. It builds Figure 2: Multiple-zone office building. on the zone-level calculations that were completed earlier (Figure 3), but in this case, we assume that each space receives minimum primary airflow at the design outdoor heating condition (Figure 5).* 4. For each zone, find the zone primary outdoor-air fraction by solving Equation 6-5 (Z p = V oz /V pz ) with the zone outdoor * Some readers might deem this to be a radical assumption because interior zones typically need more than minimum cooling airfl ow, even on the coldest day. But, it s an assumption that is likely to require a high intake airfl ow, which is useful for this demonstration. Design Cooling Condition For single-path VAV systems, the worst-case condition for ventilation (that is, the lowest system ventilation effi ciency and the highest required intake airfl ow) in the cooling mode usually occurs when the VAV primary airfl ow for the system is at its highest value. Since almost all VAV systems exhibit load diversity (all zones don t require peak cooling airfl ow simultaneously), the critical zone can be assumed to be delivering minimum primary airfl ow with the central fan at cooling-design or block primary airfl ow. In some cases, worst-case ventilation in the cooling mode may actually occur at a central fan airfl ow that s slightly lower than block airfl ow. If a system doesn t have much load diversity (all interior zones, for example)and if the critical zone requires a lot of primary airfl owthen the central fan may or may not be at block airfl ow when the critical zone is at minimum primary airfl ow. How can you fi nd out the system primary airfl ow at the worst-case ventilation condition? Simply assume that primary airfl ow at the fan is the sum of all noncritical-zone peak airfl ow values plus the minimum primary airfl ow for the critical zone. At this condition, the difference between s and Z p will be greatest, so system ventilation efficiency will be at its lowest value and outdoor-air intake fl ow will be at its highest valuesthe worst-case condition. (Operationally, this worstcase condition may not actually occur, since it assumes that the critical zone requires minimum primary airflow even when fully occupied; this might be the case for some perimeter zones, for example, during cold weather.) J a n u a r y A S H R A E J o u r n a l 3 1

21 Cooling Heating Procedural Step Variable R p P z R a A z V bz E z * V oz E z ** V oz Ventilation Zone Box Type cfm/p p cfm/ft 2 ft 2 cfm cfm cfm South Offi ces Reheat , West Offi ces Reheat , South Conference Room Reheat , East Offi ces Reheat , Southwest Interior Offi ces VAV , Northeast Interior Offi ces VAV , North Offi ces VAV , North Conference Room VAV , * For zones with a throttling VAV box, discharge air is usually cool whenever the zone is occupied (morning warmup usually occurs before occupancy). ** For zones with terminal reheat, discharge air temperature can be either cool or warm when the zone is occupied, so E z drops from 1.0 when cooling to 0.8 when heating. Figure 3: Zone ventilation calculations. airfl ow (V oz ) values for heating from Step 3 and the minimum primary airflow setting. At minimum primary airflow, the south office needs Z p = 260/475 = 0.55 when delivering warm air. 5. Using Table 6.3 (not shown) and the highest zone primary outdoor-air fraction among the zones served ( max Z p = 0.55 for the south, west, and east offices) to look up the corresponding default system ventilation effi ciency (E v ), we find that E v = Find occupant diversity according to Equation 6-7 (D = P s / P z ), as shown previously. In our example, D = 164/224 = Find the uncorrected outdoor-air intake flow for the system from Equation 6-6 (V ou = D (R( p P z ) + (R( a A z )). Once again, without correcting for zone air-distribution effectiveness and system ventilation efficiency, our system needs V ou = 2,800 cfm of outdoor air. 8. Finally, find outdoor-air intake flow for the system by solv- Multiple-Zone Systems In multiple-zone recirculating systems, such as constant-volume reheat systems and all varieties of VAV systems, one air handler supplies a mixture of outdoor air and recirculated return air to two or more ventilation zones. The required outdoor-air intake fl ow only can be determined by properly accounting for system ventilation effi ciency. Why? These ventilation systems include an unavoidable built-in ineffi ciency. This ineffi ciency exists because the intake airfl ow must be suffi cient to ventilate the critical zonethe zone that requires the highest fraction of outdoor air in its primary airstream. Since a multiple-zone system delivers the same primary air mixture to each ventilation zone, proper minimum ventilation in the critical zone overventilates all other zones. As a result, some outdoor air leaves the building via the relief, exhaust, and exfi ltration airstreams without performing useful dilution. This ineffi ciency isn t necessarily bad; it simply must be recognized and accounted for in system ventilation calculations. ing Equation 6-8 (V V ot = V ou /E/ ). In our example, V v ot = 2,800/0.60 = 4,670 cfm at design heating conditions. The system is less efficient at this heating condition than it was at the design cooling condition (system ventilation efficiency of 0.60 in heating vs in cooling). So, using the default approach (Table 6.3), worst-case/highest outdoor-air intake fl ow occurs at the design heating condition (V V ot = 4,670 cfm), assuming that all zones receive minimum primary airflow. Case 3: Ventilation Calculations for Calculated Cooling Design As mentioned previously, Addendum 62n allows the designer to use either a default or calculated value for system ventilation effi ciency (E v ). We used the default approach in Cases 1 and 2. Now, let s look at the calculated approach, which uses the equations found in Appendix G. Again, we build on the zone-level calculations (Figure 3) to find the minimum system-level outdoor-air intake fl ow (V V ot ) needed at the design cooling condition (Figure 6): 4. Find the minimum discharge outdoor-air fraction (Z = d V oz /V dz ) for each zone, using the zone outdoor airfl ow (V oz ) for cooling operation. Notice that this fraction differs from the primary outdoor-air fraction (Z p = V oz /V pz ) in the default approach. In this case, we re interested in the fraction of outdoor air in the airstream that discharges into the zonenot in the primary airstream from the air handler. 5. Find occupant diversity according to Equation 6-7 (D = P s / P z ) using expected peak system population (P s ) and design zone population; as in the default approach (Case 1), D = 164/224 = Find the uncorrected outdoor-air intake flow for the system by solving Equation 6-6 (V ou = D R p P z )+ (R( a A z )). Again, without correcting for zone air-distribution effectiveness and system ventilation efficiency, the system needs V ou = 2,800 cfm of outdoor air. 7. Establish the system primary airfl ow (V ps = LDF V pz - This nuance makes no difference for single-path systems (Vpz = V dz d ), but becomes an important distinction for dual-path systems with local recirculation, as we ll see in future articles. 3 2 A S H R A E J o u r n a l a s h r a e. o r g J a n u a r y

22 From Figure 3 From Table 6.3 Procedural Step Ventilation Zone V pz (Design) V pz-min V oz-clg Z p-clg E v cfm cfm cfm South Offi ces 1, West Offi ces 2, South Conference Room 3, East Offi ces 2, Southwest Interior Offi ces 7,000 1, Northeast Interior Offi ces 7,000 1, North Offi ces 1, * 0.65 North Conference Room 1, System (Step 6) D 0.73 (Step 7) V ou 2,800 (Step 8) V ot 4,310 * For ventilation-critical zones with a throttling VAV box, discharge air is usually cool whenever the zone is occupied (morning warmup usually occurs before occupancy). Figure 4: System ventilation calculations for default efficiency cooling design (Case 1). peak). In VAV systems, primary airflow to each zone varies with load. Of course, system primary airflow also varies but it never can be more than the central fan can deliver. (The system is always least efficient when primary airflow is high and critical-zone airflow is low because all noncritical zones are overventilated at this condition.) The central VAV fan usually is selected to deliver block, not sum-of-peak, airflow. In our example office, we assumed a system load diversity factor (LDF ) of 0.70, so the central fan delivers V ps = ,600 = 18,600 cfm at the design cooling load. 8. Find the average outdoor-air fraction ( s = V ou /V ps ) for the system. In our example, s = 2,800/18,600 = 0.15 at the design cooling condition. 9. For each zone, find zone ventilation effectiveness using Equation G-1 (E vz = 1 + s Z d ) for single-path systems. 10. Find system ventilation effi ciency using Equation G-3 (E v = minimum E vz ). In our example, E v = 0.65 at the design cooling condition. As in the default approach (Case 1), the north offices are the ventilation-critical zone. 11. Finally, find outdoor-air intake fl ow for the system by solving Equation 6-8 (V V ot = V /E ). In our example, V ou v ot = 2800/0.65 = 4310 cfm at the design cooling condition. This is identical to the intake requirement we found using the default approach. Why? The default approach is based on an assumed average outdoor-air fraction ( s ) of By coincidence, that value matches this example s average outdoor-air fraction at design cooling. In most cases, however, these numbers will differ. We refer to the zone that requires the highest fraction of outdoor air in its discharge (primary plus recirculated) airstream as the ventilation critical zone. From Figure 3 From Table 6.3 Procedural Step Ventilation Zone V pz (Design) V pz-min V oz-htg Z p-htg E v cfm cfm cfm South Offi ces 1, * 0.60 West Offi ces 2, * South Conference Room 3, East Offi ces 2, * Southwest Interior Offi ces 7,000 1, Northeast Interior Offi ces 7,000 1, North Offi ces 1, North Conference Room 1, System (Step 6) D 0.73 (Step 7) V ou 2,800 (Step 8) V ot 4,670 * For ventilation-critical zones with a throttling VAV box, discharge air is usually cool whenever the zone is occupied (morning warmup usually occurs before occupancy). Figure 5: System ventilation calculations for default efficiency heating design (Case 2). Now that we know the minimum intake at the design cooling condition, let s use the calculated approach to find the minimum intake for the design heating condition. The highest of these two intake values is the worst-case intake airflow. Case 4: System Ventilation Calculations for Calculated Heating Design As in the default approach for heating design (Case 2), assume that all spaces receive minimum primary airflow at the design heating condition. Building on the zone-level calculations (Figure 3), we ll follow the same steps that we used in Case 3 to calculate efficiency and intake airflow for cooling design (Figure 7). 4. For each zone, find the minimum discharge outdoor-air fraction (Z = V d oz /V dz ), using the appropriate V oz value for heating operation. For example, the south offices need Z = 260 d /475 = 0.55 when heating. 5. Find occupant diversity according to Equation 6-7 (D = P s / P z ), D = 164/224 = Find the uncorrected outdoor-air intake flow for the system by solving Equation 6-6 (V ou = D (R( p P z ) + (R( a A z )); as before, V ou = 2,800 cfm. 7. Establish the system primary airflow ( V ps ). For design heating calculations, we assume that all zones receive minimum primary airflow at worst case, so V ps = 6,650 cfm in our example. 8. Find the average outdoor-air fraction ( s = V ou /V ps ) for the system. In our example, s = 2,800/6,650 = 0.42 at the design heating condition. 9. For each zone, find zone ventilation effectiveness using Equation G-1 (E vz = 1 + s Z ). d J a n u a r y A S H R A E J o u r n a l 3 3

23 From Figure 3 Procedural Step Ventilation Zone V pz V pz- V oz- Z d- E vz (Design) min clg clg clg cfm cfm cfm South Offi ces 1, West Offi ces 2, South Conference Room 3, East Offi ces 2, Southwest Interior Offi ces 7,000 1, Northeast Interior Offi ces 7,000 1, North Offi ces 1, * North Conference Room 1, System (Step 5) D 0.73 (Step 6) V ou 2 2,800 (Step 7) V ps 18,600 (Step 8) s 0.15 (Step 10) E v 0.65 (Step 11) V ot 4,310 * For ventilation-critical zones with a throttling VAV box, discharge air is usually cool whenever the zone is occupied (morning warmup usually occurs before occupancy). Figure 6: System ventilation calculations for calculated efficiency cooling design (Case 3). From Figure 3 Procedural Step Ventilation Zone V pz V pz- V oz- Z d- E vz (Design) min htg htg htg cfm cfm cfm South Offi ces 1, * West Offi ces 2, * South Conference Room 3, East Offi ces 2, * Southwest Interior Offi ces 7,000 1, Northeast Interior Offi ces 7,000 1, North Offi ces 1, North Conference Room 1, System (Step 5) D 0.73 (Step 6) V ou 2 2,800 (Step 7) V ps 6,650 (Step 8) s 0.42 (Step 10) E v 0.87 (Step 11) V ot 3,220 * For ventilation-critical zones with a throttling VAV box, discharge air is usually cool whenever the zone is occupied (morning warmup usually occurs before occupancy). Figure 7: System ventilation for calculated efficiency heating design (Case 4). 10. Find system ventilation efficiency using Equation G-3 (E v = minimum E vz ). In our example, E v = 0.87 at the design heating condition. As before, the south, west, and east offices are equally critical for design heating calculations. Notice, too, that the ventilation system is much more efficient at this condition. When the average outdoor-air fraction ( s ) approaches the critical zone s outdoor-air fraction (Z d ), less unused air is exhausted; consequently, system ventilation efficiency rises. 11. Finally, find outdoor-air intake fl ow for the system by solving Equation 6-8 (V V ot = V ou /E/ ). In our example, V v ot = 2,800/0.87 = 3,230 cfm at the design heating condition. The system is more efficient at the design heating condition than it was at the design cooling condition (system ventilation effi ciency of 0.87 in heating vs in cooling). So, using the calculated approach (Appendix G), worst-case/highest outdoor-air intake fl ow occurs at the design cooling condition (V V ot = 4,310 cfm). Reviewing our previous calculations, if we simply use the default table to find system ventilation efficiency (Cases 1 and 2), our example design needs outdoor-air intake fl ow of 4,670 cfm, which occurred at the design heating condition. If we use the more complicated but more accurate calculations in Appendix G (Cases 3 and 4), our example design needs outdoor-air intake fl ow of 4,360 cfm, which occurred at the design cooling condition. Since either approach is allowed, the designer can comply using either of these intake airflow values. Assuming that our system controls can maintain the minimum required intake airflow, we can now size both the cooling coil and the heating coil for worst-case outdoor-air intake flow. What About Part-Load Operation? To comply with Addendum 62n, we need to find the highest minimum outdoor-air intake fl ow (V V ot ), which we ve called worst-case intake airflow. We could apply optional adjustments (averaging) for short-term conditions in our worstcase calculations, but we chose not to do so in the preceding discussion. In some cases, averaging adjustments can lower the worst-case intake value. In others, averaging can be used to assure proper ventilation when either supply-fan capacity or outdoor-air intake flow varies. Adjustments for short-term conditions can help the designer find the appropriate worst-case minimum intake flow. Having found this value, the system can be designed to maintain this intake airflow during all occupied hours. In VAV systems, where both primary airflow and mixing-box pressure change in response to zone demands for cooling, this usually requires some means for sensing intake airflow and modulating the outdoor-air damper to maintain the minimum airflow setting. But, do we really need to treat the worst-case outdoor airflow at all operating conditions, without regard to current ventilation needs? No. 3 4 A S H R A E J o u r n a l a s h r a e. o r g J a n u a r y

24 Equations and Variables from Addendum 62n [6-1] V bz = R p P z + R a A z [6-2] V oz = V bz z [6-3] V ot = V oz [6-4] V ot = V oz [6-5] Z p = V oz /V pz single-zone systems 100% outdoor-air systems [6-6] V ou = D allzones R p P z + allzones R a A z = D allzones V bzp + allzones V bza [6-7] D = P s / allzones P z [6-8] V ot = V ou /E v multiple-zone recirculating systems [6-9a] T = 3v/V bz IP version [6-9b] T = 50v/V bz SI version where A is zone floor area, the net occupiable floor area of the z zone, ft² (m²) D is occupant diversity, the ratio of system population to the sum of zone populations E v is ventilation efficiency of the system E z is air-distribution effectiveness within the zone P s is system population, the maximum simultaneous number of occupants in the area served by the ventilation system P is zone population, the largest expected number of people z to occupy the ventilation zone during typical usage (See caveats in Addendum 62n Section ) R a is area outdoor air rate, the required airflow per unit area of the ventilation zone determined from Addendum 62n Table 6.1, cfm/ft² (L/s m²) R p is people outdoor air rate, the required airflow per person determined from Addendum 62n Table 6.1, in cfm/person (L/s person) T is averaging time period, minutes v is ventilation-zone volume, ft³ (m³) V bz is breathing-zone outdoor airflow, the outdoor airflow required in the breathing zone of the occupiable space(s) of the ventilation zone, cfm (L/s) V ot is outdoor air intake flow, adjusted for occupant diversity and corrected for ventilation efficiency, cfm (L/s) V ou is the uncorrected outdoor air intake flow, cfm (L/s) V oz is zone outdoor airflow, the outdoor airflow that must be provided to the zone by the supply-air-distribution system at design conditions, cfm (L/s) V pz is zone primary airflow, the primary airflow that the air handler delivers to the ventilation zone; includes both outdoor air and recirculated return air Z p is zone primary outdoor air fraction, the fraction of outdoor air in the primary airflow delivered to the ventilation zone for VAV systems, Z p for design purposes is based on the minimum expected primary airflow, V pzm. In multiple-zone recirculating systems, system ventilation efficiency almost always increases as primary fan airflow decreasesprovided, of course, that design efficiency is properly calculated at the worst-case condition (that is, with low primary airflow to the critical zone). Although we must design the system with sufficient capacity for worst-case intake airflow, we could operate it at many conditions with less-than-worst-case intake and still comply with Addendum 62n. To do so, our design could incorporate one of the optional dynamic reset approaches presented in Addendum 62n, using a control approach that resets intake airflow to match current requirements at part-cooling load. In a future article, we ll examine partload operation and optional dynamic reset in detail. For now, we simply note we always must design for worst-case intake flow (as discussed earlier), regardless of any dynamic reset control options we may choose to implement. In other words, dynamic reset does not alter the worst-case outdoor-air intake flow needed to comply with the standard. Summary Historically, Standard 62 required both zone- and system-level calculations for the design of single-path, multiple-zone ventilation systems (like throttling VAV systems). Unfortunately, the calculation procedures were unclear and frequently misinterpreted or ignored by designers. As a result, many multiple-zone systems were improperly ventilated. Addendum 62n clarifies the multiplezone system calculations to reduce both underventilation and unnecessary overventilation. It allows a simple default approach, as well as a more accurate calculated approach for determining system ventilation efficiency. As shown here, either calculation procedure can be readily applied to single-path VAV systems at the design conditions for both cooling and heating, to provide a compliant determination of worst-case minimum outdoor-air intake flow. References 1. ANSI/ASHRAE Standard , Ventilation for Acceptable Indoor Air Quality. 2. ANSI/ASHRAE Addendum n to ANSI/ASHRAE Standard Stanke, D Addendum 62n: single-zone and dedicated-oa systems. ASHRAE Journal 46(10): Stanke, D Standard Addendum 62n: ventilation for changeover-bypass VAV systems. ASHRAE Journal 46(11): J a n u a r y A S H R A E J o u r n a l 3 5

25 2004, American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. ( org). Reprinted by permission from ASHRAE Journal, (Vol. 46, No. 11, November 2004). This article may not be copied nor distributed in either paper or digital form without ASHRAE s permission. Standard Addendum 62n Ventilation for Changeover-Bypass VAV Systems By Dennis Stanke, Member ASHRAE A NSI/ASHRAE Standard , Ventilation for Acceptable Indoor Air Quality, 1 prescribes ventilation rates for commercial and institutional buildings. Historically, Standard 62 required both zone- and system-level calculations for the design of multiple-zone ventilation systems, such as single-path constant volume and VAV systems. Unfortunately, without clear calculation procedures, system-level calculations frequently were misinterpreted or ignored by designers. As a result, many multiple-zone systems were underventilated. In an effort to avoid underventilation and increase calculation consistency among designers, Addendum 62n 2 updates prescribed ventilation rates and the calculation procedure for zone ventilation airflow and for system intake airflow for different ventilation systems. A previous article 3 introduced Addendum 62n rates and described the design of single-zone and 100%- outdoor-air ventilation systems. This article discusses the 62n-compliant design of a very specific multiple-zone ventilation system: changeover-bypass VAV, also called variable volume and temperature (VVT). A changeover-bypass VAV system (Figure 1) uses a constant-volume air handler (often a packaged rooftop unit or split D system) to ventilate and cool or heat a large area within a building. The area served comprises many comfort zones or HVAC zones, def ined by ANSI/ASHRAE/IESNA Although Addendum 62n shows ventilation rates in both IP and SI units, this paper uses IP, except in selected specifi c calculations. This is because 62n uses rational conversions, not mathematical. Standard as areas with heating and cooling requirements that are sufficiently similar so that desired conditions can be maintained throughout using a single sensor. Each comfort zone has its own thermostatically controlled VAV damper and ideally, all comfort zones associated with an air handler have similar thermal loads (that is, all zones need some level of heating or some level of cooling, like interior spaces or southern perimeter spaces). In practice, however, many systems include comfort zones with dissimilar loads; some zones need heating at the same time that other zones need cooling. In either case, each comfort zone usually is considered as a ventilation zone, defined by Addendum 62n as an area with similar occupancy category, occupant density, zone air distribution effectiveness, and zone primary airflow per unit area. About the Author Dennis Stanke is a staff applications engineer with Trane, La Crosse, Wis. He is vice chair of Standing Standards Project Committee A S H R A E J o u r n a l a s h r a e. o r g N o v e m b e r

26 The changeover-bypass VAV system includes a central bypass damper that modulates open as the zone VAV dampers modulate closed. The air handler delivers approximately constant primary airflow, while each comfort zone receives a variable air volume to match its thermal load. Air-handler mode (heating or cooling) is usually based on zone-by-zone polling to determine the number of zones calling for heating or cooling, for instance, and the strength of the calls. Air-handler capacity may be controlled by the largest load, or by simply maintaining primary air temperature at a heating or cooling setpoint. As the zone dampers close and zone airflow drops, the bypass damper opens to maintain a constant differential pressure between the primary duct and the return plenum, and consequently, relatively constant airflow through the fan and coils. This results in relatively constant outdoor-air intake flow (ignoring wind effects). Note: Depending on building construction and the operation of relief (exhaust) fans and/or dampers, mixed-air pressure may actually rise relative to outdoor pressure as the bypass damper opens. If the minimum outdoor-air-damper position is fixed, it should be set to ensure at least minimum outdoor-air intake fl ow with the bypass damper at both extremes (open to maximum and fully closed). A single duct delivers air for thermal comfort and ventilation to all comfort zones. When the system is occupied, the outdoor air damper at the air handler opens to provide minimum outdoor-air intake flow. This outdoor airflow helps determine design cooling and heating capacity. Designing to Comply with Addendum 62n Changeover-bypass VAV systems can be designed (and operated) to comply with Addendum 62n,* thereby ensuring proper ventilation in each zone. To comply, the system must deliver sufficient outdoor-air intake fl ow at the air handler as well as sufficient outdoor air to the breathing zone within each ventilation zone. These flows, which are inextricably linked, must be determined for design cooling and design heating conditions. Because this is a multiple-zone recirculating system, wherein the air handler supplies a mixture of outdoor air and recirculated return air to more than one ventilation zone, Equations 6.5 through 6.8 of Addendum 62n must be used. In the following paragraphs, we consider two approaches to ventilation system design for an area within an example office building (Figure 2). In the first approach, we work from the zones to the air handler, selecting minimum zone primary airfl ow settings and then calculating the outdoor-air intake fl ow needed for proper ventilation of all zones. In the second approach, we work from the air handler to the zones, selecting * Along with many other Web-published addenda, Addendum 62n will be incorporated into an updated version of Standard and published as ASHRAE Standard later this year. the outdoor-air intake fl ow and then calculating the required settings for minimum zone primary airfl ow. Based on thermal loads, our example building area is divided into eight comfort zones, each with its own thermostat and VAV damper. We consider each of these comfort zones as a separate ventilation zone. The following zone-level calculations must be performed, regardless of which approach to system design is used (Figure 3): 1. Look up the minimum people outdoor air rate (R p ) and the minimum area outdoor air rate (R a ) in Addendum 62n, Table 6.1, for each zone. Using peak zone population (P z ) and zone fl oor area (A z ), find the design breathing-zone outdoor airflow using Equation 6-1 ( V bz = R p P z + R a A z ). (We could have used average zone population for P z, but then we probably could not have taken credit for occupant diversity. See Step 6.) For example, the south offices in our example building require V bz = = = 220 cfm. 2. For each zone, look up zone air-distribution effectiveness (E z ) based on the air-distribution configuration and the default values in Table 6.2 (see excerpt in Figure 4). In our example, all offices use overhead diffusers and ceiling returns, and receive 58 F (14 C) primary air during cooling and 95 F (35 C) during heating. Effectiveness changes from 1.0 when cooling to 0.8 when heating, demonstrating how operating mode can affect zone ventilation performance. 3.Find zone outdoor airflow using Equation 6-2 (V oz = V bz / E ) for both cooling and heating operation. z For our example south offices, V oz = 220/1.0 = 220 cfm during cooling, and V oz = 220/0.8 = 275 cfm during heating. Therefore, the system must be designed to deliver at least V oz = 275 cfm, either continuously or on average, to ensure adequate ventilation in both modes. (During operation, the example VAV dampers may be allowed to close to deliver less than 275 cfm when the primary air temperature doesn t match the zone thermal requirements, provided system controls are designed to ensure that average minimum primary airflow is at least 275 cfm. See Design for Short-Term Conditions. ) Addendum 62n doesn t explicitly require calculation of V oz at both cooling and heating design, but it is implicit in the requirements for system ventilation efficiency and in the variable load conditions section, which states that required ventilation rate must be delivered under all load conditions. Each ventilation zone in our example is assigned to one of two air handlers, forming two multiple-zone ventilation systems (Figure 2). The interior system includes all interior zones and the north zones, which are largely adjacent to a conditioned warehouse space; these zones usually need some level of cooling, regardless of outdoor conditions. The perimeter system includes zones that need some level of cooling in warm weather and some level of heating in cold weather. N o v e m b e r A S H R A E J o u r n a l 2 3

27 EA RA Warehouse Bypass Damper OA V ot Supply Fan SA V ps North Offi ces North Conf. Room Constant-volume supply fan variable air volume to spaces VAV Damper V pz V pz Northeast Interior Offi ces Interior RTU Perimeter RTU Figure 1: Changeover-bypass system. The following procedures describe the two approaches to ventilation system design mentioned previously. Each approach can result in systems that comply with Addendum 62n. Design Procedure 1: Choose Minimum Zone Settings, Find Intake Airflow Using this approach, we first select minimum settings for zone primary airfl ow and then determine the minimum outdoor-air intake fl ow at the air handler. We assume continuous primary airflow (non-zero minimums) to illustrate the calculations. In some cases (as in our example perimeter system), continuous primary airflow can quickly overcool or overheat some zones, making them uncomfortable or causing frequent system changeover (heating to cooling or vice versa). While continuous primary airflow works well for ventilation, it may not provide acceptable comfort and/or energy use. Lower or even intermittent primary airflow (zero minimums) can be used to design changeover-bypass systemsbut only if operation results in an average primary airflow that equals or exceeds the minimum chosen for continuous primary airflow. Averaging is discussed in more detail later (see Design for Short-Term Conditions ). The ventilation rate procedure in Addendum 62n prescribes the steps for calculating both zone-level ventilation airflows (described earlier) and systemlevel intake airflow. Before we examine these steps, though, we must take care of some preliminary calculations and settings. First, based on design cooling load and primary air temperature, determine design zone primary airfl ow (V pz ) for each zone. We use cooling airflow because it s usually much West Offi ces South Conference Room Southwest Interior Offi ces North Offi ces East Offi ces Figure 2: Changeover-bypass office building. Advertisement formerly in this space. N o v e m b e r A S H R A E J o u r n a l 2 5

28 Cooling Heating Procedural Step Variable R p P z R a A z V bz E z V oz E z V oz Ventilation Zone cfm/p p cfm/sf sf cfm cfm cfm South Offi ces , West Offi ces , South Conference Room , East Offi ces , Southwest Interior Offi ces Northeast Interior Offi ces , , , ,060 North Offi ces , North Conference Room , Figure 3: Zone ventilation calculations, which must be performed regardless of chosen approach. higher than heating airflow. (This example ignores zone-tozone load diversity for simplicity, but as in other VAV systems, primary airflow at the changeover-bypass air handler often is less than the sum of the peak zone airflows.) Next, establish a minimum primary airflow setting (V V pz-min ) for each zone. In this case, we set the zone minimums to 30% of design airflow. We know that Standard 90.1 allows reheat (if we need it) with 30% minimum settings, and although we might prefer lower settings, 30% minimums allow us to use default values for system ventilation efficiency (see Step 5). (With lower minimum settings, Standard 62 may have required us to calculate E v, rather than use default values.) Depending on the VAV damper controls, it may be possible to set two minimumsone for cooling and another for heatingbut we considered only the simple case of a single minimum setting for each zone. For design, we assumed that these are average values for minimum primary airflow. For operation, the system controls must enforce these average minimums. Zero minimums are acceptable only if the system controls are designed to deliver, on average, the design minimum primary flow. (See Design for Short-Term Conditions. ) The following system-level steps (Figure 5) build on the zone-level steps covered previously, so we begin with Step Using the highest zone outdoor airflow (V V oz-clg or V oz-htg ) and the lowest primary airflow (V pz = V pz-min ), find the minimum primary outdoor-air fraction for each zone using Equation 6-5 (Z p = V oz /V pz ). For example, for our south offices, Z p = 275/540 = Addendum 62n doesn t explicitly differentiate between heating-mode and cooling-mode Z p values, but Table 6.3 (next step) implies that the highest V oz must be used because we must find the largest value of Z p among all the zones served by the system. Note: To avoid a design that requires 100% intake airflow, minimum (or average) V pz settings must exceed zone Advertisement formerly in this space. N o v e m b e r A S H R A E J o u r n a l 2 7

29 outdoor airflow ( V oz ). As a result, some systems may require duct reheat or zone heat to avoid overcooling at low load; depending on thermal zoning, systems without local heat may be less comfortable and may change over more frequently. 5. Look up system ventilation effi ciency (E v ) in Table 6.3 (Figure 6). Use the largest primary outdoor-air fraction ( max Z p ) because it represents the critical zonethe zone that needs the highest percentage of outdoor air in its primary airstream. Advertisement formerly in this space. For our example, max Z p occurs in the south offices (Z p = 0.51) and the north offices (Z p = 0.52). Interpolating in Table 6.3, we find that minimum system ventilation effi ciency E v = 0.64 and 0.63 for each system, respectively. Alternatively (or if max Z p exceeds 0.55), we could determine Ev using the equations in Addendum 62n, Appendix G,** but most changeover-bypass system designers likely prefer the simpler default-table approach. Note: Lower minimum settings, though perhaps preferred for zone operation, would result in max Z p greater than 0.55 and require use of the equations rather than the default values from Table 6.3; reduced E v values and increased outdoorair intake fl ow would result. 6. Find occupant diversity for the system using Equation 6-7 (D = P s /ΣP z ). Use a reasonable estimate for the expected system population (P s ), along with the sum-of-peak zone population (P z ). In this example, we assumed that the conference rooms would only be populated by people from the nearby office spaces. So, system occupant diversity is D = 60/90 = 0.67 for the perimeter system, and D = 100/136 = 0.74 for the interior system. 7. Find the uncorrected outdoor-air intake flow for the system using Equation 6-6 (V ou = D Σ(R( p P z )+Σ(R a A z )). For our perimeter system, V ou = = 840 cfm; for the interior system, V ou = ,440 = 1,950 cfm. 8. Finally, find the design (minimum) outdoor-air intake fl ow using Equation 6-8 (V V ot = V ou /E ). In our example, the v perimeter system needs V ot = 840/0.64 = 1,310 cfm (about 14% outdoor air), while the interior system needs V ot = 1950/0.63 = 3,100 cfm (about 18% outdoor air). As is always the case in multiple-zone recirculating systems, high zone minimum settings lead to low intake airflow requirements. This design approach (zone-to-system) results in a relatively low outdoor airflow percentage, but it requires high minimum settings for zone primary airflow. This may mean reheat in many zones or frequent heating/cooling changeover. ** In some cases, calculated E v values may be lower than default values, but Appendix G calculations are left to future articles. 2 8 A S H R A E J o u r n a l a s h r a e. o r g N o v e m b e r

30 Design Procedure 2: Choose Intake Minimum, Find Zone Minimum Settings Using this second approach, we select the minimum outdoor-air intake flow at the air handler and then determine the required minimum zone-primary-airflow setting for each space. In this case, we start with more intake airflow than we found using the first approach. As we will see, more intake airflow results in lower zone minimum settings (or averages). Lower minimums mean less overcooling and/or overheating and less frequent changeover, which in turn may mean increased occupant comfort and lower energy use. Air Distribution Confi guration Although you won t find the following calculation steps (Figure 7) detailed there, this approach is based on the principles of Addendum 62n. The calculations build on the zone-level calculations (Figure 3) covered earlier, so we start with Step 4: 4. Establish the design (minimum) outdoor-air intake fl ow (V V ot ) within the constraints of the air handler (such as primary airflow per ton of cooling capacity or unit heating capacity). We assume that each packaged unit in our example can handle 30% outdoor air, so V ot = 9, = 2,730 cfm for the Ceiling Supply of Cool Air 1.0 Ceiling Supply of Warm Air and Floor Return Ceiling Supply of Warm Air at Least 8 C (15 F) Above Space Temp. and Ceiling Return Ceiling Supply of Warm Air less than 8 C (15 F) above Space Temperature and Ceiling Return, Provided That the 0.8 m/s (150 fpm) Supply Air Jet Reaches to Within 1.4 m (4.5 ft) of Floor Level Figure 4: Zone air-distribution effectiveness for several ventilation-zone configurations. Not all Table 6.2 configurations are listed. perimeter unit and V ot = 17, = E z 5,220 cfm for the interior unit. Again, we ignore load diversity in this example by 1.0 assuming that primary airflow is simply the sum of the peak zone airflows Find occupant diversity using Equation 6-7 (D = P s /ΣP z ). For the perimeter and interior systems, respectively, D = 60/ = 0.67 and D = 100/136 = Find the occupant portion of breathing-zone outdoor airflow (R p P z ) for each zone using the people outdoor-air rate (R p ) from Addendum 62n, Table 6.1, and the design zone population (P z ). 7. Using the area outdoor-air rate (R a ) from Table 6.1 and zone floor area (A z ), find the building portion of breathing-zone outdoor airflow (R a A z ) for each zone. 8. Find the uncorrected outdoor-air intake fl ow for each system using Equation 6-6 (V ou = D Σ(R( p P z ) + Σ(R( a A z )). In our example, the perimeter system needs V ou = = 840 cfm, while the interior system needs V ou = ,440 = 1940 cfm. 9. Using Equation 6-8 (E v = V ou /V V ot ) and the V ou and V ot values already established, find the lowest allowable system ventilation Equations and Variables from Addendum 62n [6-1] V bz = R p P z + R a A z [6-2] V oz = V bz / E z [6-3] V ot = V oz [6-4] V ot = V oz [6-5] Z p = V oz /V pz [6-6] V ou = D allzones R p P z + allzones R a A z = D allzones V bzp + allzones V bza single-zone systems 100% outdoor-air systems [6-7] D = P s / allzones P z [6-8] V ot = V ou /E v multiple-zone recirculating systems [6-9a] T = 3v/V bz IP version [6-9b] T = 50v/V bz SI version where A z is zone fl oor area, the net occupiable fl oor area of the zone, ft² (m²) D is occupant diversity, the ratio of system population to the sum of zone populations E v is ventilation effi ciency of the system E z is air-distribution effectiveness within the zone P s is system population, the maximum simultaneous number of occupants in the area served by the ventilation system P z is zone population, the largest expected number of people to occupy the ventilation zone during typical usage (See caveats in Addendum 62n Section ) R a is area outdoor air rate, the required airfl ow per unit area of the ventilation zone determined from Addendum 62nTable 6.1, cfm/ft² (L/s m²) R p is people outdoor air rate, the required airfl ow per person determined from Addendum 62nTable 6.1, in cfm/person (L/s person) T is averaging time period, minutes v is ventilation-zone volume, ft³ (m³) V bz is breathing-zone outdoor airfl ow, the outdoor airfl ow required in the breathing zone of the occupiable space(s) of the ventilation zone, cfm (L/s) V ot is outdoor air intake fl ow, adjusted for occupant diversity and corrected for ventilation effi ciency, cfm (L/s) V ou is the uncorrected outdoor air intake fl ow, cfm (L/s) V oz is zone outdoor airfl ow, the outdoor airfl ow that must be provided to the zone by the supply-air-distribution system at design conditions, cfm (L/s) V pz is zone primary airfl ow, the primary airfl ow that the air handler delivers to the ventilation zone; includes both outdoor air and recirculated return air Z p is zone primary outdoor air fraction, the fraction of outdoor air in the primary airfl ow delivered to the ventilation zone for VAV systems, Z p for design purposes is based on the minimum expected primary airfl ow, V pzm. N o v e m b e r A S H R A E J o u r n a l 2 9

31 Cooling Design Min. From (SAT = 58 F) Setting Fig. 3 Procedural Step Outdoor Ventilation Zone Clg Load V pz Figure 5: System ventilation calculations for Design Procedure 1 (fixed minimum zone settings). efficiency for proper ventilation. For the perimeter system, lowest allowable E v = 840/2730 = 0.31, and for the interior system lowest allowable E v = 1,940/5,220 = Use Table 6.3 (Figure 6 ) to find the maximum primary outdoor-air fraction ( max Z p ) for any zone based on the lowest allowable system ventilation efficiency. Our minimum E v values are off the chart, but because we found the lowest allowable E for each system, we can use any value above 0.31 or 0.37, v respectively, to find max Z p. For simplicity (that is, to stay in the default table), we used the lowest E v value in Table 6.3, v = 0.60, for both systems; that value corresponds to max Z p In other words, no zone can require a primary outdoor air fraction (Z p ) greater than Given V oz ( Figure 3) and using Equation 6-5 (V pz = V oz /Z p ), find each zone s minimum setting for zone primary airfl ow. For the south offices in our example, minimum V pz = 275/0.55 = 500 cfm. This design approach yields somewhat lower minimum settings (or averages) for zone primary airflow than the first procedure. Raising outdoor-air intake flow allows lower minimum primary airflow settings and therefore, less reheat or local heat (if reheat is used) and less frequent changeover, which may increase comfort and reduce energy use. Maximum Z p * Many changeover-bypass system designers are likely to use the default table to find max Z p because it s relatively simple, even though it yields conservative results; that s why we used it in our example. But, the equations in Appendix G can result in a much lower max Z p value when using a high intake-airflow value, so minimum zone primary airfl ow settings can be much Min. Primary V pz -min V oz Z p E v D V ou V ot cfm/sf cfm % cfm cfm (Tbl. 6.3) cfm cfm % South Offi ces 0.9 1, * West Offi ces 1.0 2, South Conference Room 1.1 3, East Offi ces 1.0 2, Perimeter System 9, ,310 14% Southwest Interior Offi ces 0.7 7, ,100 1, Northeast Interior Offi ces 0.7 7, ,100 1, North Offi ces 0.8 1, * North Conference Room 0.9 1, Interior System 17, ,950 3,100 18% * Ventilation-critical zone. E v lower (almost one half in some zones) with only a little more work. Designers may want to consider using the equations to comply with the standard, rather than the default table; calculations are not as simple, but results are more accurate and less conservative. Design for Short-Term Conditions Regardless of the design approach used, Addendum 62n recognizes that it may be reasonable to base design airflow values on average conditions over a finite period, rather than peak conditions. Addendum 62n list three acceptable design adjustments : It specifically allows use of an average zone population (P z ) in Equation 6-1 for calculation of the design breathing-zone outdoor airfl ow (V bz ). (This seems to System Ventilation preclude the use of occupant diversity Effi ciency, E > 0.55 Use Appendix G * Largest zone primary outdoor airfl ow fraction, calculated using Equation 6-5, among all of the ventilation zones that the system serves. Figure 6: E v defaults from Addendum n to ANSI/ASHRAE Standard , Table 6.3. v Air Intake (D ) at the system level, to avoid double credit at both zone and system population; we didn t use average zone population in our examples.) It allows interruptions in delivered breathing-zone outdoor airflow (V bz )provided that the average airflow over averaging time period T equals or exceeds the minimum V bz value found using Equation 6-1. (This short-term condition design option may be key for changeover-bypass systems with minimum primary airflow settings of zero, but calculating average V bz requires detailed knowledge of system control operation, which varies greatly by manufacturer and designer. Due to this variation in system operation, detailed averaging calculations are beyond the scope of this article.) It allows interruption of outdoor-air intake fl ow (V V ot ), pro- 3 0 A S H R A E J o u r n a l a s h r a e. o r g N o v e m b e r

32 Setting Zone Minimums For Design Procedure 1, our rule of thumb set the minimum value for zone primary airfl ow (V pz ) at about twice the minimum zone outdoor airfl ow (V oz ) needed for ventilation. This results in a zone-air-distribution effectiveness (Z p ) of about 50%, so we can use Table 6.3 default values rather than Appendix G calculated values. Lower minimum settings (like zero) could be selected, provided system controls are designed to deliver at least V oz on average. Minimum settings also may be limited by diffuser selection and location. When the damper is open, minimum primary airfl ow should result in a diffuser velocity that will ensure good air distribution in the space (without diffuser dumping ) when cooling, and good air mixing in the space (without stratifi cation) when heating. If electric reheat is used, minimum primary airfl ow may be based on the minimum airfl ow needed for safe heater operation. Standard 90.1 requirements also may impact minimum primary airflow settings. To comply using primary air reheat, the volume of air reheated must be no more than the highest of: The minimum volume required for ventilation (V oz ). An airflow rate of 0.4 cfm/ft². An airflow equal to 30% of design peak airflow (V pz ). An airflow of 300 cfm, for small zones. Any higher rate that reduces overall energy use by reducing intake airfl ow (V V ot ), even though it increases local reheat energy. In our example zones, all minimum primary airflow settings are 30% of peak design airflow and all are lower than 0.4 cfm/ ft², so reheat would be allowed in any of the zones. vided that the average intake airflow equals or exceeds the value found using Equations 6-5 through 6-8. C h a n g e o v e r - b y p a s s VAV s y s t e m s w i t h ve r y l ow o r z e r o - m i n i m u m s e t t i n g s f o r z o n e primary airfl ow tend to interrupt zone ventilation for short periods, making them candidates for design based on primary airflow averaging. Before we average anything, we need an averaging time. For a given ventilation zone, Equation 6-9 (T = 3v/V bz ) from Addendum 62n defines required averaging time as the ratio of three space volumes to the continuous breathingzone outdoor airflow. This equates to three space time constants. In response to a step change in airflow, space conditions approach steady-state values after three time constants, so this seems to be a reasonable averaging time. Assuming a 10 ft ceiling, our south office, for example, has an averaging time of T = 3 2,000 10/220 = 270 minutes in cooling mode and T = 3 2,000 10/275 = 220 minutes in heating mode. Depending on load, any zone in a changeover-bypass VAV system may require either heating or cooling at any time. For a given zone, heating or cooling mode either matches the system mode or opposes it. An opposing zone will be more comfortable with very low primary airflow. A match- ing zone must have enough airflow to ensure that, averaged over time T,, the zone receives at least the minimum required breathing-zone outdoor airfl ow. A matching zone must be overventilated to compensate for those times when it becomes an opposing zone. A system could be designed to integrate primary airflow Advertisement formerly in this space. We didn t consider the zero-minimum case in the preceding design calculations because actual design depends on the control operation specified for a system. N o v e m b e r A S H R A E J o u r n a l 3 1

33 Procedural Step From 11 Minimum V pz Fig. 3 Setting* Ventilation Zone V ot D V bz V bz V Procedure Procedure ou oz pz (At 30%) (Occ.) (Area) E v Z p (Min.) 1 2 cfm cfm cfm cfm cfm cfm % % South Offi ces West Offi ces South Conference Room East Offi ces Perimeter System 2, Southwest Interior Offi ces ,060 1, Northeast Interior Offi ces ,060 1, North Offi ces North Conference Room Interior System 5, ,440 1, * To aid comparison, minimum primary airfl ow settings (V pz ) are shown as percentages of design primary airfl ow. Figure 7: System ventilation calculations for Design Procedure 2 (fixed minimum intake setting). over the averaging time (T) for each zone, allowing zero opposing minimums and resetting the matching minimums as high as necessary to ensure that the required continuous breathing-zone outdoor airfl ow is provided on average. In any case, Addendum 62n specifically allows designs based on averaging for short-term conditions, but leaves implementation details to the designer. Operational Considerations The preceding design procedures result in proper matching of minimum zone airflow settings with minimum intake airflow settings, allowing proper equipment selection. But Addendum 62n addresses ventilation system operation as well as system design. It allows the system controls to reset both minimum zone primary airflow and/or outdoor-air intake fl ow, matching current ventilation capacity to current ventilation requirements. As we saw when we considered short-term conditions, control design can be crucial for both the energy-conscious, comfortable operation of changeoverbypass systems and for proper design. The dynamic reset section of Addendum 62n allows the use of system controls that reset the design outdoor air intake fl ow (V V ot ) and/or space or zone airflow as operating conditions change. It lists three specific operating conditions, allowing a system-level reset control strategy that: Uses CO 2 - or schedule-based, demand-controlled ventilation (DCV) to reset zone-level V oz and system-level V ot in response to zone population changes. Several DCV approaches indirectly can determine or estimate the actual airflow per person and reset either V pz or V ot (or both) in response to the current requirement. Uses zone airflow to find the currently required intake airflow (V V ot ) based on recalculation of system ventilation effi ciency (E v )a dynamic version of Design Procedure 1. This includes resetting V ot as zone air-distribution effectiveness (E z ) varies due to heating/cooling operation. Lowers the minimum zone primary airfl ow during economizer operation, when V ot exceeds the minimum required intakea dynamic version of Design Procedure 2. Addendum 62n allows these dynamic reset approaches. It gives no implementation details, but the requirement is clear: Any system control approach that responds to varying conditions must be capable of providing the required breathing-zone ventilation airflow whenever the zones served by the system are occupied. Dynamic reset approaches and results can differ widely. Regrettably, they are outside the scope of this article. Summary Addendum 62n calculation procedures can be applied to changeover-bypass systems with both zero and non-zero minimum settings for zone primary airflow. Design calculations can proceed from the zone minimums to find the air-handler intake flow, or from intake flow to find the zone minimums. As we ve shown, the default table for system ventilation efficiency can be used (and is likely to be common practicesimple design for simple systems). As with other VAV systems, dynamic reset approaches are allowed by Addendum 62n, but operational controls for changeover bypass systems are beyond our present scope. References 1. ANSI/ASHRAE Standard , Ventilation for Acceptable Indoor Air Quality. 2. ANSI/ASHRAE Addendum n to ANSI/ASHRAE Standard Stanke, D Addendum 62n: single-zone dedicated-oa systems ASHRAE Journal 46(10): ANSI/ASHRAE/IESNA Standard , Energy Standard for Buildings Except Low-Rise Residential Buildings. 3 2 A S H R A E J o u r n a l a s h r a e. o r g N o v e m b e r

34 The following article was published in ASHRAE Journal, October Copyright 2004 American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. It is presented for educational purposes only. This article may not be copied and/or distributed electronically or in paper form without permission of ASHRAE. Supply Air CO 2 Control of minimum outdoor air for multiple space systems By David Warden, P.Eng., Member ASHRAE Most ventilating systems have shortcomings in the control of their minimum outdoor air intake. They do not always deliver the intended ventilation to the occupants, their real performance is hard to check and they do not adjust minimum outdoor air intake as the ventilation demand varies. Demand control ventilation (DCV), based on sensing the rise in space CO 2 concentration, can address these problems 8,9 but is difficult to apply to multiple space systems. Supply air CO 2 control (SACO 2 ) effectively applies CO 2 technology to recirculating systems. It is inexpensive, saves energy, and helps ensure good indoor air quality. Supply Air CO 2 Control Description SACO 2 is a technique for measuring the outdoor air fraction in the supply air and controlling the outdoor air intake, so the supply air always contains a high enough fraction of outdoor air to ventilate any space served by the system. It is applicable to recirculating systems serving multiple spaces where ventilation targets are based on outdoor airflow rate per person. Figure 1 illustrates application of SACO 2 to a typical VAV system. A single CO 2 sensor alternately senses CO 2 concentration in the supply duct and outdoors. A valve switches between sources, and fan suction draws air through the sensor. The outdoor air intake is controlled so the rise in CO 2 concentration between outdoors and the supply air does not exceed a value that corresponds to the required minimum outdoor air fraction in the supply air. Calculation of this minimum outdoor air fraction and the maximum rise in CO 2 concentration and the outdoor air intake fraction are illustrated in Sidebars Space Ventilation Calculations, CO2 Calculations and Minimum Outdoor Air Intake Calculation. (Other values in Figure 1 were calculated in a similar manner.) Outdoor Air in Multiple Space Recirculating Systems As the supply air must contain a high enough outdoor air fraction to satisfy the space with the greatest need, most other spaces receive a surplus of outdoor air (a) because they need more supply air to control temperature than to provide ventilation, and (b) because of population diversity between spaces. This surplus unused outdoor air is About the Author David Warden, P.Eng., is the principal at Warden Engineering in Victoria, BC, Canada. 2 6 A S H R A E J o u r n a l a s h r a e. o r g O c t o b e r

35 Relief 43% OA C o ppm A A A OA Intake C o ppm 18% OA 82% Recirc. Containing 43% OA CO 2 S NO NC DDC OA REF. Supply Air 53% OA C o ppm Relief 51% OA C o ppm A A A OA Intake C o ppm 5% OA 95% Recirc. Containing 51% OA CO 2 S NO DDC OA REF. Supply Air 53% OA C o ppm to satisfy a fully occupied space System peak overall occupancy 200 cfm/p SA 107 cfm/p OA C o ppm Individual office min. supply airflow and max. occupant density 37.5 cfm/p SA 20 cfm/p OA C o ppm System 25% overall occupancy 800 cfm/p SA 426 cfm/p OA C o ppm Individual office min. supply airflow and max. occupant density 37.5 cfm/p SA 20 cfm/p OA C o ppm Figure 1: SACO 2 system at peak occupancy. Figure 2: SACO 2 system at partial occupancy. recirculated and often contributes more ventilation than first pass air directly from the outdoor air intake (twice as much in Figure 1). Calculations to account for the effect of unused outdoor air were developed by Kowalczewski 3 more than 30 years ago and ASHRAE Standard 62, Ventilation for Acceptable Indoor Air Quality, has required similar calculations for multiple space systems since (see Equation 6-1 in Standard and system ventilation efficiency E v from Table 6-3, or Appendix G in Addendum 62n. 2 ) Derivations that may be most useful and easiest to find are Appendix G in Standard and ASHRAE Journal article Outdoor Air: Calculation and Delivery, 4 which expanded the method to handle fan-powered terminal systems and other situations with a secondary source of recirculated air. Supply Air CO 2 as Occupancy Varies As the total number of building occupants varies, the unused outdoor air content and the CO 2 concentration in recirculated air varies. The CO 2 sensor detects the effect this has on the supply air and, unless more outdoor air is needed for makeup or free cooling, the SACO 2 system adjusts the outdoor air intake to maintain the design minimum outdoor air fraction in the supply air. Storage in the air volume served by the system causes the concentration of contaminants directly emitted by occupants and the CO 2 concentration to lag the change in occupancy in a similar manner so even under changing conditions, CO 2 is an indicator of the outdoor air content available for ventilation. Figure 2 illustrates the reduction of outdoor air intake for the previous example under 25% occupancy conditions, assuming that the total supply flow is unchanged. The system still provides enough outdoor air for any fully occupied space but the necessary outdoor air intake fraction drops to 5%. Similarly, if the number of system occupants exceeds the design occupancy, CO 2 concentration rises, the sensor detects this, and SACO 2 increases the outdoor air intake to correct the rise in CO 2 concentration and the outdoor air content of the supply air. OA, Sources and Short-Circuits Outdoor air often enters through windows and doors, transfers from adjacent systems with a ventilation surplus or leaks from supply ducts into the return system. When such air is recirculated, the SACO 2 system detects the drop in supply air CO 2 concentration and reduces the minimum outdoor air intake accordingly. Similarly, if relief air short-circuits into the outdoor air intake, the SACO 2 system detects the rise in CO 2 concentration and increases outdoor air intake as needed. VAV Systems and Zone Flow Reset If total system supply airflow decreases (e.g., loaded filters or VAV), the return air contains less unused outdoor air and the SACO 2 system adjusts the O c t o b e r A S H R A E J o u r n a l 2 7

36 outdoor air intake to maintain the total outdoor air fraction in the supply air. (The outdoor air intake fraction rises, but the outdoor air intake flow falls slightly as the relief air contains less outdoor air and less is carried out of the building before it can be used by occupants.) If the outdoor air content in the supply air is higher than needed (e.g., makeup air or economizer cycle), the SACO 2 system will detect and measure this. If desired, this information can be used to dynamically calculate and adjust minimum zone supply flow rates to reduce reheat and fan energy. Empty and Nearly Empty Buildings Some contaminants affecting odor, irritation and/or health come from building surfaces and contents, or are temporarily absorbed and reemitted from these surfaces. Although not all codes require it, it is highly desirable to ventilate the building before scheduled occupancy to reduce contaminant buildup after any sustained period without ventilation (e.g., overnight or on the weekend). Similarly, a limit exists on reducing outdoor air intake during periods of very low occupancy. 62n s area outdoor air rates (0.06 to 0.18 cfm/ft 2 [0.3 to 0.9 L/s per m 2 ]) apply where 62n has been adopted into code. Elsewhere, they could be used as a benchmark, but it is a lot of air for a single occupant and less ventilation may be practical under low occupancy conditions in clean buildings with low emitting materials. In some cases, enough air may enter through general infiltration, makeup for exhausts, suction through closed outdoor air dampers and transfer of air from adjacent spaces with a ventilation surplus. As some of these sources are hard to quantify, to maintain building pressure control it is usually desirable to supplement demand ventilation control with an absolute minimum intake airflow whenever the building is occupied. A simple method of providing this absolute minimum intake, such as a minimum signal to the outdoor air damper may suffice, but for large or special systems, more sophisticated methods may be appropriate. New and Existing Buildings Installation of SACO 2 is simple in either new or existing buildings. In existing buildings, however, establishing the setpoint requires more work because the relationship between existing space supply rates and the design outdoor air intake needs to be checked. As most existing ventilation systems were not designed based on multiple spaces calculations, this analysis often will reveal spaces with ventilation problems. To solve these existing problems, it may be necessary to increase primary supply to these spaces, provide them with a secondary source of recirculated air or increase the outdoor air fraction for the whole system. Compared to an existing ventilation system that is properly designed and operating correctly, SACO 2 saves energy to the extent that the area it serves is less than fully occupied over the system s operating hours. In practice, existing systems bring in anything from no outdoor air, to a large excess. When this is corrected by the SACO 2 system, energy use will change accordingly. Cost of Supply Air CO 2 Control Purchase, installation, programming, calibration and commissioning cost of SACO 2 in U.S. dollars based on competitive pricing and local installation by contractors familiar with the work is likely to be about $500 plus $1,000/sensor, plus general contractor s margins and taxes. Engineering costs can be expected to be relatively minor on new projects. On existing projects, time will be spent to review existing drawings and reports, visit the site, analyze the existing system, prepare separate documents and check the installation. This will result in costs likely to be on the order of $2,000 plus $600/sensor if the work is done by an Meeting Rooms Fully occupied meeting rooms supplied from an office system require a higher minimum supply flow rate per unit area than office spaces due to their high occupant density. In 62-89, this is exacerbated by an outdoor air per person rate that is slightly high compared to similar spaces. (See the table in sidebar Space Ventilation Calculations. ) This supply rate usually is more than is needed for cooling and unless reheat is provided, the space will be overcooled. Increasing the outdoor air intake for the whole system increases the outdoor air fraction in the supply air, thus reducing the minimum supply rate for the meeting room and saving reheat but increasing initial cost and energy used to condition outdoor air. A more efficient approach is to add direct recirculation of unused outdoor air from spaces that have a surplus of outdoor air. Ideally, this second ventilation source is from a central point to take full advantage of occupant diversity across a wide area (e.g., dual fan dual duct). Other alternatives are fan-powered VAV terminals or transfer fans arranged to draw secondary recirculation from a location where unused outdoor air will be available (e.g., adjacent to a central return air collection point). An exhaust fan from a meeting room is not an efficient solution because it only increases ventilation to the extent that it is larger than the supply and it is difficult to control the source and quality of any air it draws into the space. The required flow of secondary air can be calculated by rearranging Equation 6 from Outdoor Air, Calculation and Delivery 4 or equation G-2 from Addendum 62n Appendix G 2 and back-calculating based on the outdoor air fraction in the primary supply. The following additional measures can also save energy: 1. A button for occupants to change the supply rate or activate a transfer fan. 2. Reduce the minimum ventilation supply rate to the meeting room when the outdoor air fraction in the supply air is high (e.g., outdoor air economizer cycle). This combines well with SACO Sense space CO 2 to reset minimum supply flow based on demand (attractive in large meeting rooms with highly variable occupancy). 2 8 A S H R A E J o u r n a l a s h r a e. o r g O c t o b e r

37 engineer set up for this work. On the first project, about two extra weeks of engineering time is required to adequately learn the concept and develop appropriate methods, spreadsheets, detail drawings, technical specifications, and check sheets. These costs do not include fixing existing space ventilation problems, faulty dampers or expanding the DDC system capacity (one extra analog input and one extra digital output). Comparison With Other Minimum OA Control Strategies A new control method is only relevant if it offers benefits compared to existing alternatives. This section is intended to help readers assess this. First, it discusses significant criteria and identifies typical problems that may occur. Second (Table 1), it lists alternative methods and identifies criteria that I expect they may have problems satisfying. Full discussion of these opinions is not possible in this article, but I expect that once alerted to a potential problem, readers will be able to assess this for themselves. Third, other CO 2 -based methods are discussed in more detail. Function: If a method cannot be counted on to perform the desired function when it is needed, then the method is of little value. One example would be control from temperature balance in warm weather when there is inadequate temperature difference between outdoor and return air. Temperature balance also can be problematic at other times due to inability to accurately read the mixed air temperature. Other examples discussed below are multisensor space CO 2 control and average return air CO 2 control. The function of SACO 2 is discussed previously. Setup: Methods that depend on setting a specific minimum flow, but have no provision to permit accurate direct measurement of that flow by a balancing agent, are difficult to set up. Sometimes they can be set up based on the difference between supply and return flows, but it is often not possible to do this with sufficient accuracy. In this case, they are usually set up by measuring the total flow and the temperatures of return, outdoor and supply air, then calculating the minimum outdoor airflow based on temperature balance. Difficulties in getting accurate temperature measurements sometimes make the results more an expression of hope rather than an accurate representation of what is happening. Assuming that the airsupply system is (or will be) balanced, SACO 2 can be set up in a just few minutes without requiring balancing equipment or waiting for special weather conditions (see CO 2 Sensing ). Reliability: Control based on using a fixed signal to place dampers precisely in a slightly open position deteriorates over time as wear increases play in linkages and when the setting is lost during damper or actuator maintenance. My experience is that control from direct flow sensing in the outdoor airstream is subject to dirt blocking the upstream pitot openings or affecting electronic sensing elements in the airstream. Multisensor space CO 2 control is a concern due to its many potential failure points. SACO 2 has relatively little that can go wrong. One sensor and one switching valve serve the whole system. SACO 2 can alarm a fault, and as it is an active method of control, it compensates for play in damper linkages. Energy: Systems that do not reduce the minimum outdoor intake during partial occupancy can waste a lot of energy as most buildings have far less than the peak number of occupants most of the time. CO 2 -based DCV systems can deal with this issue, the other (non-dcv) systems cannot. SACO 2 also saves by adjusting for recirculation of outdoor air that has entered through doors or windows or has spilled over from adjoining systems with a surplus. Another energy benefit with SACO 2 for VAV systems is the ability to save reheat energy by resetting the space flow minimums down based on sensing any rise in outdoor air content of the supply air due to makeup air or an airside economizer cycle. Cost: Control from a CO 2 sensor in each space is extraordinarily expensive. Multipoint space sensors generally will be less costly but are not cheap. A separate minimum outdoor air damper and outdoor air duct arranged for flow measurement are more expensive than most other methods and a separate outdoor air fan or direct flow sensing increases this cost. The cost of the remaining methods likely is acceptable for most multiple space systems but, in some cases, may be a burden for very small systems. The cost of SACO 2 is reasonable as only one sensor is needed for the whole system. The one sensor potentially can handle more than one system, and no special mixing damper or intake arrangement is needed. Maintenance: Systems based upon placing dampers in a precise partially open position should be regularly checked and recalibrated by a balancing agent. Direct flow sensing in the outdoor air intake implies regular and careful cleaning. CO 2 sensors also need periodic checking and recalibration. With the right sensor, this is simple and takes five to 15 minutes. Some operators do it themselves and some use a specialist at about $100/sensor including a report. This is more than many owners currently pay because at present they are not doing the maintenance needed to ensure their existing system really works. VAV: Systems that mix in a fixed minimum percentage of flow from the outdoor air intake have serious problems with VAV. If the supply flow drops to 50%, so will the flow through the outdoor air intake. If the damper setting is based on the minimum supply flow then at maximum flow, outdoor air intake may be double what is needed, resulting in high energy use and the cost of larger equipment. If the dampers modulate through a range of minimum positions based on total supply flow, the result will be better, but not particularly accurate. Adaptability: If the actual number of occupants served by the system is greater or smaller than the number on which the design is based, SACO 2 and other CO 2 based systems will automatically adapt. Other systems will deliver the wrong outdoor air quantity unless the intake is recalulated and reset. If open plan office space is changed to a meeting room O c t o b e r A S H R A E J o u r n a l 2 9

38 Minimum Outdoor Air Control Method Description Function Setup Reliability Energy Cost Maintenance VAV Adaptability Verifiability Record Passive Damper Position Separate Damper Minimum opening position on modulating damper set by damper actuator. No flow measurement provisions. Two position minimum outdoor damper separate from any modulating OA damper. OA Duct Two position minimum OA damper plus separate minimum outdoor air duct arranged for flow measurement during system setup. OA Fan Separate constant flow, ducted, fan powered, minimum outdoor air to main system supply. DOAV Separate constant flow, ducted, fan powered, minimum outdoor air system directly serving each space. n/a Active Control Direct Flow Sensing Separate minimum outdoor air duct with flow sensing controlling a minimum OA damper. Temperature Balance Control OA & RA dampers to provide minimum outdoor air based on a temperature balance between OA, RA and SA (and supply airflow for VAV). Pressure Balance Separate two-position min. OA damper acts as a flow orifice. RA damper controls suction pressure in the mixing plenum to ensure the desired minimum OA intake airflow. Active Control with Occupancy Feedback Space CO 2 Ind. Sensors Control OA & RA dampers based on a separate CO 2 sensor for each space. Space CO 2 Multipoint Control OA & RA dampers using a limited number of sensors to sample every space. Average Return CO 2 Control OA & RA dampers based on the common return air from all spaces or groups of spaces. Supply CO 2 Control OA & RA dampers based on sensing CO 2 concentration in the supply duct. Note: This table is intended to provide a simple indication of issues to watch out for with different methods of outdoor air control. It is somewhat subjective (e.g., what cost is acceptable?) and does not provide absolute ratings that would apply in all circumstances. The darker shading indicates that functional issues are likely to make the method inappropriate in most cases. Table 1: Potential issues for minimum outdoor air control methods. without increase in system-wide population, most systems just require changing the minimum supply rate to the space, but with some systems, more work is required. With room CO 2 sensing, an extra CO 2 sensor or CO 2 sampling point is needed. With DOAV systems ductwork changes, rebalancing and increasing outdoor air intake are required. Verifiability: With most systems it is not practical for an operator to personally check if the system is delivering the intended minimum outdoor air. They can check that something is happening, but to check whether it achieves the desired result, they need to call in a balancing agent. This is relatively expensive and rarely done. With SACO 2 it is easy to check if the supply air CO 2 readings are as they should be, and checking sensor calibration only takes a few minutes with simple equipment (see CO 2 Sensing ). Record: Being able to log and chart important aspects of HVAC system operation is now the expected norm, yet few systems have provision to do this for ventilation, which is one of the most critical issues in terms of health, occupant satisfaction and protection from litigation. Many outdoor air-control systems have no really meaningful way to do this. With dedicated outdoor air fans, you can record that the fan was pressurizing its duct system. Inlet flow sensing can record intake of outdoor air. SACO 2 can record the outdoor air content in the supply to the spaces and space CO 2 can give an indication of outdoor air per person delivered to the space. Direct outdoor air ventilation (DOAV) systems that do not recirculate but directly supply each space with enough outdoor air to meet its peak demand are popular in Europe and gaining interest in North America. Their outdoor air intake is the sum of the peak space requirements; hence, ventilation calculation is simple. Some other differences are: 1) They do not provide airside free cooling or enhanced ventilation in mild weather. 2) They do not recirculate unused outdoor air back to the occupants (e.g., from empty spaces, partially occupied spaces and air that bypassed the breathing zone). 3) Ventilation of over-occupied spaces is less effective as recirculated air from 3 0 A S H R A E J o u r n a l a s h r a e. o r g O c t o b e r

39 CO 2 Sensing CO 2 sensors are sophisticated devices that are better and less costly than in the past. Still, they are not cheap, not precise and their calibration drifts over time. Special attention is needed to how they are applied, how they are calibrated and how accuracy is checked. Sensor Configuration. The CO 2 sensing confi guration shown in Figure 1 uses a single CO 2 sensor to alternately sense supply air and outdoor air via a three-way EP valve. Fan suction draws outdoor air samples through the sensor. This confi guration offers the following benefi ts: Outdoor air CO 2 concentration is measured rather than assumed. This compensates for the variations that occur through the day, during the year, and between sites. A single sensor provides far more accurate control than separate supply air and outdoor sensors as error is largely cancelled out when the difference in concentration is calculated. Sensor faults can be alarmed based on outdoor air readings outside normal range. The sensor can be located remotely in a well-lit location at a comfortable height to check the reading on a built in display, verify sensor accuracy and recalibrate the sensor. My experience is that using 0.25 in. (6 mm) pneumatic tube, runs from source to sensor of 33 ft (10 m) work well. The suction run from the sensor could be much longer as a larger tube diameter could be used without affecting the time to change from sensing one source to the other. A single sensor can control any system size. Where appropriate, valving can be arranged to permit a single sensor also to sense return air, the second supply duct of a dual fan system, supply ducts from more than one system, or individual spaces. Sensor Calibration. CO 2 sensors are calibrated using specially formulated test gases (typically nitrogen to set the zero and a CO 2 /nitrogen mix to set the top of the range). To minimize error, it is important to use calibration gas at a concentration close to that at which you wish to control. If the supply air is going to be at about 700 ppm, calibrating with special 700 ppm ±2% gas is far more accurate than with standard 2,500 ppm ±2% gas. It also makes it easy to check whether the sensor needs recalibrating. Verification and Recalibration. Checking sensor accuracy is simple, quick and can avoid unnecessary recalibration. Provided the sensor has a display, it only requires a bottle of gas at the concentration of interest, a regulator and a plastic tube. Sensors drift most when they are new, so checking three months after installation is a good idea. Thereafter, if the recommended sensor confi guration with outdoor air reference is used, an annual check should be fi ne. If the system has an outdoor air economizer, minimum outdoor air will be seasonal, and it is best to check the sensor at the start of the main season in which the system will operate on minimum outdoor air. Auto recalibration is an option on many sensors. It typically resets the sensor based on the assumption that the lowest reading of the week is at the typical outdoor concentration. If this assumption is not true, then auto recalibration can drive the sensor into serious error. The recommended method (Figure 1) avoids this risk, is more accurate and offers self-checking. Control Issues. The control system should read the supply air CO 2 concentration most of the time and periodically sample the outdoor air. The supply air concentration changes quickly as the dampers move but the outdoor air CO 2 concentration normally is comparatively stable. CO 2 sensors sample the air in their sensing chamber (e.g., every 10 seconds), and noticeable jitter exists between readings even if there is no change in the gas. For control purposes, the readings should be averaged to improve accuracy and stability. As fast changes are not desirable in minimum outdoor air control, averaging the signal over one or more minutes is not a problem. When switching between sources, the controls should ignore readings until the new sample flows through the sensor, readings stabilize and it has enough readings to average out jitter. Following any extended period when outdoor air ventilation was shut down or the space was unoccupied (night or weekend), a separate control method should be used to bring in suffi cient outdoor air to dilute contaminants that have built up while the system was off. other spaces is not available to help flush contaminants. 4) They must be physically modified and rebalanced if a change is made in the building (e.g., new meeting room). If all spaces are continuously at peak occupancy, a DOAV system requires less outdoor air intake than a recirculating system. However, in all buildings with a normal diversity of population that I have checked, DOAV systems have required more outdoor air intake than recirculating systems because much of their outdoor air is delivered to spaces that are empty or partially occupied, and is then exhausted. (In a partially occupied building, the DOAV outdoor air intake is unchanged and is far greater than for a recirculating system with SACO 2.) Individual CO 2 sensors for each space with minimum outdoor air intake controlled to satisfy the space with the greatest demand is possible in theory, but unattractive in practice. As most spaces could potentially have the greatest demand under some partial occupancy condition, the number of sensors required is generally large. As the number of sensors increases, so does the cost, the maintenance needs and the risk of system failure. This method depends on every sensor working properly. If any sensor reads significantly high, it will increase the outdoor air intake for the whole system, quite possibly to 100%. Limiting sensors to selected typical spaces, or disregarding a certain percentage of high readings, risks ignoring and underventilating the spaces that are most heavily occupied at a particular time. Multipoint space CO 2 sensing uses one sensor to sample many spaces plus outdoor air (just as one sensor can sample supply air and outdoor air). This would improve cost and reliability, but the number of spaces that can be effectively sensed with one sensor is limited by the sampling time needed to draw air from a space and get a stable reading. Average return air CO 2 sensing generally is not a satisfactory method of minimum outdoor air control because it does not ensure delivery of the required minimum outdoor air per person to a fully occupied space when the overall system is only partially occupied (see sidebar Why Average RA CO 2 Control Does Not Work ). It is similar to controlling the temperature of a whole building from a single thermostat. Setting the return air CO 2 control setpoint for the worst partial occupancy condition is no solution, as it would cause corresponding overventilation during full occupancy. 2 6 A S H R A E J o u r n a l a s h r a e. o r g O c t o b e r

40 Advantages of SACO 2 Control: 1. Excellent probability of good ventilation because: a. Accurate setup is easy at any time of year. b. SACO 2 corrects for damper linkage wear and avoids the potential problems of damper position setpoints being changed or lost during maintenance or repair. c. It corrects for short circuiting of relief air back into the outdoor air intake. d. It is quick and easy to check the outdoor air system operation and sensor accuracy. e. The CO 2 sensor readings can be monitored and charted to demonstrate performance. f. Outdoor air readings can trigger an alarm if they suggest that the sensor needs recalibration or is faulty. g. If the original total building design occupancy is exceeded, SACO 2 automatically adjusts the outdoor air intake. 2. Energy savings can be large if the average number of occupants is much less than the peak number or if the system replaces one that doesn t operate correctly. 3. Installation and maintenance costs are low, as a single sensor serves one or more entire systems. 4. SACO 2 is easily applied to new or existing buildings. Space Ventilation Calculations These calculations apply to any multiple space recirculating system regardless of the method of minimum outdoor air control. The example is based on the outdoor air per person rates of ASHRAE Standard 62 ( editions) rather than the two-part rates of Addendum 62n but refers to Addendum 62n where a concept is not dealt with in the base standard. For convenience, fl ow terms and equations have been calculated on a per person (R terms) or per unit area basis (V terms). In principle, one could start by choosing a design outdoor air fraction in the supply air (Z s ) and go straight to Step 4 to calculate the corresponding minimum space supply airflow rate for each space type. In practice, Steps 1 to 3 help optimize between outdoor air intake and minimum space fl ow needs: Step 1: Select the space type on which design will be based. Select the common space type with the least supply air per person. Other spaces will be supplied with minimum flow rates calculated in Step 4. In this case, the dominant occupancy type is offi ce space and the offi ces that receive the least supply air per person will be private interior offi ces with low thermal loads. As this is a VAV system, space supply will be least when the supply air is coldest. Step 2: Determine the minimum supply flow per person to the space. Assume for this example, a 150 ft 2 (14 m 2 ) offi ce with two people plus three lights emitting 210 W, negligible equipment load (e.g., laptop computers) and a supply air T of 15 F (8 C). The resulting space supply fl owrate R sz without reheat is: R sz = (2P x 250 Btu/h/P W x Btu/h/W)/ 15 F/1.08 Btu/h/ F/cfm/2P = 37.5 cfm/person. Step 3: Calculate the outdoor air fraction (Z ) required in the supply air. From Standard the ventilation target for occupants of offi ce spaces is R obz = 20 cfm/person. The room ventilation effectiveness E z = 1 (62n, table 6.2) and as E z = R obz /R oz where R oz is the required space supply rate to achieve the required ventilation rate in the breathing zone R obz : R oz = R obz /E z = 20 /1 = 20 cfm/person And the minimum required outdoor air fraction in the space supply air, Z z = R oz /R = 20 cfm/person/37.5 cfm/person = 0.53 (i.e., 53%). sz As the primary supply air is the sole source of ventilation to the space, the minimum required outdoor air fraction in the system supply air to satisfy this space (Z s ) is also Step 4: Establish minimum supply flow rates for each space type. Based on the above outdoor air fraction, calculate the corresponding minimum supply rate required to ventilate each space type when it is at peak design occupant density as shown in the following table. Space Type Min. OA Rate/Person cfm/person R obz Space Air Distribution Effectiveness E z OA Fraction In Supply Air Z s Min. Supply Air Rate/ Person cfm/p R sz = R obz / E z Z Design Max. Occ. Density ft 2 /P A z / P z Min. Supply Air Rate/ Unit Area cfm/ft 2 V sz = R sz /(A z / P z ) Office a Office (heating) b b 0.53 a d 75 d f Meeting Room c Reception c e g 0.5 Notes a. Interior private offi ce; the occupancy type upon which the minimum OA fraction in the supply air was based. b. Perimeter offi ce: When heating from the ceiling with air, supply air can bypass the occupants, particularly if VAV turndown results in hot supply air and low discharge velocity. This example uses E =0.8 (20% bypass), as per 62n. If supply air is too hot and the velocity is low, E z can be lower than 0.8. In this case however, stratifi caz tion usually causes thermal discomfort and correcting the comfort problem also improves E z. c. Used E = 1.0, even for perimeter meeting rooms, as the high internal heat gain minimizes heating needs. z d. 75 ft 2 /person (7 m 2 /person) for private offi ces because two people routinely meet in a 150 ft 2 (14 m 2 ) private offi ce. (62-89 s default density of 7 people/1,000 ft 2 [93 m 2 ] is insuffi cient for private offi ces in any building I have ever seen. 62n s default of 5 people /1,000 ft 2 (93 m 2 ) is even less realistic for private offi ces but is reasonable for the diversifi ed building density) e. 20 ft 2 /person (1.9 m 2 /person) (50 people/1,000 ft 2 [93 m 2 ]) for meeting rooms as per 62-89, 62n, and personal experience. f. The minimum ventilation supply rate for perimeter offi ces heating with air is rarely a practical concern when they are on the same system as interior offi ces. E may v be less than 1.0 but the minimum space supply rate to achieve comfort, and maintain VAV terminal control on turndown, is usually higher than is needed for adequate ventilation. g. Meeting rooms in offi ce buildings generally require reheat to maintain room temperature if they are supplied with adequate ventilation air from the primary system alone. Alternative solutions are discussed in the Sidebar Meeting Rooms. Table 2: Minimum supply flow rates for each space type. 2 7 A S H R A E J o u r n a l a s h r a e. o r g O c t o b e r

41 CO 2 Calculations Step 1: Calculate the rise in CO 2 concentration if all outdoor air was used. If the building only was supplied outdoor air at the rate it is used by the occupants, there would be no unused outdoor air remaining in the building return. The concentration rise under this steady state condition can be calculated using an equation similar to that in ASHRAE Standard , Appendix C. C ru C o = 1,000,000 N b M /R os = 1,000,000 x x 1.2 / ppm where C ru = CO 2 concentration in recirculated air if all outdoor air supplied to the building is used C o = CO 2 concentration outdoors 1,000,000 is conversion from a fraction to ppm N b = CO 2 generation rate per person at base metabolic rate = cfm/person ( L/s per person) M = Relative metabolic rate in met units from Standard , Appendix C, Figure C-2 = 1.2 R os = The rate at which occupants use outdoor air, i.e. the average design outdoor air ventilation rate for the area served by the system = 20 cfm/person (10 L/s per person). This was based on the worst occupant distribution (greatest fraction of occupants in spaces requiring the highest ventilation rates). Return air may be regarded as a mixture of outdoor air that has been used and outdoor air that was supplied in excess of requirements and remains unused. In this case, the return air may be regarded as a mix of used outdoor air at 550 ppm above outdoors and unused outdoor air at 0 ppm above outdoors. Step 2: Calculate the supply air CO 2 concentration rise. Supply air may be regarded as a mix of recirculated used outdoor air at 550 ppm above outdoors and unused outdoor air. This unused outdoor air is made up of recirculated unused outdoor air plus fi rst pass outdoor air directly from the outdoor air intake. If supply air is 100% outdoor air, its concentration is 0 ppm above outdoors, and from Step 1, a CO 2 concentration of 550 ppm above outdoors corresponds to 0% outdoor air (for this example). Hence, the rise in concentration corresponding to the desired outdoor air fraction in the supply air from the Space Ventilation Calculation sidebar (Z s = 0.53) is: C C o = Z s 0 + (1 Z s ) 550 = 0 + (1 0.53) 550 = 260 ppm C s Minimum OA Intake Calculation (Optional) Calculating the minimum outdoor air intake airflow is not essential to implement SACO 2, but is needed for equipment sizing and is useful for optimization as discussed below. Calculation Example For this example, the overall system supply flow rate (V s ) is assumed to be 1 cfm/ft 2 (5 L/s per m 2 ) and the average occupant density across the area served by the system is assumed to be a 200 ft 2 /P (19 m 2 /P ). As in Sidebar CO 2 Calculations, the average required ventilation rate at peak system conditions is assumed to be 20 cfm/person (10 L/s per person). The outdoor airflow used by all occupants of the system, V ou = 20 cfm/person/200 ft 2 /person = 0.1 cfm/ft 2 (0.004 L/s per m 2 ). The outdoor air used by all occupants expressed as a fraction of the system supply, s = V ou / V = 0.1/1.0 = 0.1. os As the outdoor air fraction in the supply air Z s = 0.53 from sidebar Space Ventilation Calculations, and of this, s = 0.1 is used, the outdoor air fraction remaining in the system return air, Z r = Z s s = = From this, Y,, the outdoor air intake flow expressed as a fraction of the supply airflow can be calculated with a simple mixing calculation: Y = (Z Z s Z r ) / (1 - Z r ) = ( ) / (1 0.43) = This means that 18% first pass outdoor air directly from the intake is suffi cient to provide a 53% outdoor air fraction in the supply air. The remaining 35% is outdoor air that has passed through the building without being used, and is then recirculated. Is the Answer Right? Let s check the result against the multiple spaces equation from ASHRAE Standard Expressing Equation 6-1 in the terms from this article and using the same data: Y = s /(1 + s Z s ) = 0.1/( ) = The result is identical, which is not surprising as the logic is that used to develop multiple spaces equations: 1,3,4,5 1) Calculate the outdoor air fraction needed in the supply air. 2) Calculate the outdoor air used in the whole area served by the system. 3) From 1 and 2, calculate the outdoor air fraction remaining in the return air. 4) From 1 and 3, calculate the outdoor air intake required. (Steps 3 and 4 are usually combined into one more complex equation and this makes the logic less apparent.) Why So Simple? Engineers who have applied Standard 62 s multiple spaces calculations know how much work it can be and may be puzzled that the outdoor air calculation for a large system can be this simple. The answer lies in the way the calculation process is organized. From reading Standard 62, the most obvious process is that outlined in 62n. Calculations are done for each space based upon predetermined d minimum space flow rates. System level calculations then are based on adding space values and applying population diversity. If the spacess change, the calculations are redone. While this method is valid, it is akin to weighing a ton of sand a grain at a time then reweighing it if the arrangement of grains changes! It is really only suitable for people who have a lot of free time and love doing calculations! The method used in this article calculates the outdoor air fraction in the supply air and the outdoor air intake needed to satisfy one space by looking only at that space and at the system as a whole. The distribu- tion of area, people and airflows in the other spaces is irrelevant for this purpose. This can be seen from the original derivations of the multiple space equations. 1,3,4,5 Minimum supply rates for other space types are back calculated based on the outdoor air fraction in the supply air being set to meet the needs of the base space. All of this is done on a per unit area basis, so that it can be done without knowing partitioning details and to establish a simple table showing the minimum supply rate per unit area needed to adequately ventilate each space type. The table provides simple space ventilation rules for both the original designers and those designing future building changes. Optimization Minimum space supply airflows and minimum outdoor air intake flow are interdependent. If the outdoor air intake flow seems excessive, it can be lowered by recalculating based on higher minimum space supply flow rates, or alternatively the higher space flow rates can be directly calculated by rearranging the formulae. Similarly, if the minimum space supply flow rates seem too high, they can be lowered and the effect on outdoor air intake checked. O c t o b e r A S H R A E J o u r n a l 2 8

42 Limitations of Supply Air CO 2 Control: 1. For single space systems, SACO 2 works, but space CO 2 sensing is more effective. 2. SACO 2 may not be affordable for very small systems, unless they share a common CO 2 sensor. 3. As with any minimum outdoor air system, SACO 2 should be periodically checked and recalibrated. (Having a CO 2 sensor makes this need more obvious, but makes the task easier.) 4. If no space requires the design minimum fraction of outdoor air in the supply air, SACO 2 only responds to the overall reduction in occupancy (usually not an issue). 5. After any unoccupied period, SACO 2 initially will not bring in enough outdoor air to flush any contaminants that have built up during the unoccupied period. A simple flushing cycle should be implemented to provide good air quality following unoccupied periods. 6. SACO 2 cannot be applied if minimum outdoor air needs are based totally on makeup air requirements or an outdoor air per unit area based rate. 7. If minimum outdoor air needs are based on a combination of airflow per person and makeup air requirements and/or an area-based rate, SACO 2 cannot be used alone but can be part of a more complex control strategy. ASHRAE Standard 62 Addendum n 62n changes the ventilation rate calculation section of Standard One signifi cant change is the shift from outdoor air per person ventilation rates, to additive rates made up of a smaller rate per person, plus a rate per unit area. Another is the reduction in net ventilation for most fully occupied spaces. The reduction in ventilation rate varies depending upon occupant density and space type. Fully occupied seating areas of auditoriums drop from 15 cfm/person (7 L/s per person) to 5.4 cfm/person (2.5 L/s per person). A 150 ft 2 (14 m 2 ) private offi ce with two occupants drops from 20 cfm/person (10 L/s per person) to 9.5 cfm/person (4.5 L/s per person) though the net reduction in a multiple space system with diverse occupancy is slightly less. Some other space types have much smaller changes. Health and productivity are not discussed in the addendum or its foreword, but there is broad acceptance that worse health, absenteeism and task performance are associated with ventilation rates below 20 cfm per person (10 L/s per person). 6 In view of this, why the reduced rates? Part of the answer is that to the extent that the 62n rates and additive structure were based on scientifi c data, this data was more about perceived odor rather than health. A second factor is a change in objectives. Unlike previous versions of , the rates in 62n were chosen with the goal of setting an absolute minimum standard rather than a guideline to what is desirable. One consideration was to change from a ventilation rate at which 80% of people would fi nd odor acceptable when they fi rst arrived, to a much lower ventilation rate at which 80% of people would fi nd odor acceptable after they had adapted to it. While 62n generally reduces ventilation rates under normal operating conditions, it has the reverse effect when the building is almost empty. At night, a system serving 100,000 ft 2 (9290 m 2 ) of offi ce space would have to operate and provide more than 0.06 cfm/ft 2 100,000 ft 2 = 6,000 cfm (2831 L/s) of outdoor air to meet the needs of one or two people who are working late. If a roving security offi cer is entitled to the same ventilation quality, then this would be required all night. While some odor studies are supporting this, the cost and energy consumption seem excessive. My experience suggests that the benefi t is limited and that few owners will run the ventilation systems in nearly empty buildings regardless of what the standard says. Standard 62 s rates always have relied heavily on judgment and a compromise between conflicting views as to whether they should rise or fall. Through the years, the rates have varied widely, and they probably will change again in response to different weighting of health or other issues and more scientifi c information. Most countries base their ventilation standards on rates per person, and the extent to which 62n s rates will be adopted into code in North America remains to be seen. Even in jurisdictions that adopt 62n s minimum rates, it seems desirable to operate with a higher rate at peak occupancy, except possibly in extreme weather conditions. 62n and Demand Control Ventilation With 62n s two part rates, when the number of occupants in a space changes, so does the net rate per person and so does the rise in CO 2 concentration corresponding to the 62n requirement. Similarly, when the number of occupants served by a recirculating system varies, so does the rise in supply air CO 2 concentration corresponding to the outdoor air fraction in the supply air that is required to meet 62n. SACO 2 can be adapted to handle 62n by adding a sampling point in the return air. The control system calculates the system occupancy based on the rise in CO 2 concentration between the supply air and the return, then calculates the rise in supply air CO 2 concentration that corresponds to the 62n requirement. When the occupant density is relatively high, this method of control is satisfactory. However, as the building empties, the ventilation rate per person required by 62n becomes very high and the CO 2 concentration differences become too low for practical measurement. When this point is reached, the minimum intake flow must be maintained at approximately this level. Sometimes exhaust makeup will provide enough outdoor air to meet this requirement. When it does not, relying on a minimum position signal to the outdoor air damper is one possibility. A more accurate option is to use a separate fi xed position minimum outdoor air damper as a measuring orifi ce and control the suction pressure across it by modulating the recirculation damper. An alternative approach is to sense the rise in CO 2 concentration between outdoors and the return air and measure the minimum outdoor air intake flow. Based on these two values, the control system can calculate the outdoor air intake airflow required to maintain the necessary outdoor air fraction in the supply air and adjust the outdoor air intake accordingly. This method can theoretically track the 62n requirement right down to zero, but it costs more and requires additional maintenance. As minimum outdoor air intake requirements under 62n are less dependent on the number of occupants, why would one wish to use CO 2 -based demand ventilation control? Possible reasons include: Local code authorities may not have adopted the 62n rates; A desire to improve air quality when the building is fully occupied and save energy when it is not; Concern about the performance of other methods of controlling minimum outdoor air; and A desire to have a measured record of ventilation delivery. 2 9 A S H R A E J o u r n a l a s h r a e. o r g O c t o b e r

43 Conclusions SACO 2 offers significant benefits for systems that recirculate from multiple spaces. These include energy savings, simple maintenance, better assurance of adequate ventilation and the ability to measure and record performance. Benefits are largest where indoor air quality is of concern, there are large variations in the number of occupants, the climate is extreme, or energy rates are high. It is not suited to single space systems, is not needed where makeup needs exceed ventilation needs, and requires modification if used with two-part ventilation rates (Addendum n). Acknowledgments The author would like to acknowledge the support of Kevin Jeffries and Jack Meredith of the British Columbia Buildings Corporation in developing SACO 2. References 1. ANSI/ASHRAE Standard , Ventilation for Acceptable Indoor Air Quality. 2. Addendum 62n, ANSI/ASHRAE Standard , Ventilation for Acceptable Indoor Air Quality. 3. Kowalczewski, J Quality of air in air conditioning. AI- RAH (2). 4. Warden, D Outdoor air, calculation and delivery. ASHRAE Journal 37(6). 5. Ke, Y-P., S.A. Mumma A generalized multiple spaces equation to accommodate any mix of close-off and fan powered VAV boxes ASHRAE Transactions 102(1). 6. Seppanen, O.A., W.J. Fisk, M.J. Mendell Association of ventilation rates and CO 2 -concentrations with health and other responses in commercial and institutional buildings Indoor Air 9: Krarti, M., et al. 1999, Techniques for measuring and controlling outside air intake rates in variable volume systems. JCEM TR/99/3 (ASHRAE Research Paper). 8. Emmerich S.J., A.K. Persily State-of-the-art review of co2 demand controlled ventilation technology and application. NIST, NISTIR Persily, A.K., et al Simulations of indoor air quality and ventilation impacts of demand controlled ventilation in commercial and institutional buildings. NIST, NISTIR Relief Air Outdoor Air C o ppm Supply Air (C o + 360) ppm max. for 15 cfm/person in a fully occupied classroom. (C o + 590) ppm, when partially occupied and under RA CO 2 control. Return Air CO 2 Sensor Setpoint C o ppm Classroom 1 Fully Occupied School 33 cfm SA/Person 15 cfm OA/Person (C o + 670) ppm Classroom 2 Fully Occupied School 33 cfm SA/Person 16 cfm OA/Person (C o + 670) ppm Classroom 3 Fully Occupied School 33 cfm SA/Person 15 cfm OA/Person (C o + 670) ppm Classroom 4 Fully Occupied School 33 cfm SA/Person 15 cfm OA/Person (C o + 670) ppm Partially Occupied School 33 cfm SA/Person 11 cfm OA/Person (C o + 900) ppm Partially Occupied School No Occupants in Classroom 2 (C o + 590) ppm Partially Occupied School No Occupants in Classroom 3 (C o + 590) ppm Partially Occupied School No Occupants in Classroom 4 (C o + 590) ppm Figure 3: RA CO 2 control with a simple classroom system. CO 2 concentrations were calculated based on a met rate of 1.2. Peak occupancy ventilation rates and CO 2 concentrations are shown in grey. Partial occupancy ventilation rates and CO concentrations are shown in black with problem conditions in red. 2 Why Average RA CO 2 Control Does Not Work This example illustrates RA CO 2 control with a simple classroom system. The outdoor air intake is controlled from a CO 2 sensor installed in the return air. The return air CO 2 setpoint is 670 ppm above outdoor CO 2 concentration, as this is the concentration rise when all classrooms are occupied and receiving the intended minimum outdoor air ventilation rate. During partial overall occupancy, any fully occupied classroom is underventilated. Occupancy patterns like this are common. In a school, it may occur when some classes are at a special event, when some classrooms in a wing are unused due to reduced enrollment or when the facility is used for night school. In an offi ce building it can occur at night, on weekends or when a building is partially tenanted. CO 2 concentrations were calculated using the method illustrated in Sidebar 2, based on a met rate of 1.2. Peak occupancy ventilation rates and CO 2 concentrations are shown in grey. Partial occupancy ventilation rates and CO 2 concentrations are shown in black with problem conditions in red. During partial occupancy, supply air bypassing through empty rooms reduces the rise in CO 2 concentration between the supply air and the average return air from 310 ppm to approximately 80 ppm. As the return air CO 2 is controlled at 670 ppm above outdoors and this is 80 ppm above the supply air, the supply air CO 2 concentration is = 590 ppm above outdoors. This is 230 ppm above the maximum CO 2 concentration acceptable in the supply air to a fully occupied classroom. The rise in CO 2 concentration between the supply air and the fully occupied classroom is unchanged, so the CO 2 concentration in this classroom also is 230 ppm high. ( = 900 ppm above outdoors). The result is that the fully occupied classroom has an outdoor air ventilation rate that is only 11 cfm/person (5 L/s per person) rather than the 15 cfm/person (7 L/s per person) that was intended. 3 0 A S H R A E J o u r n a l a s h r a e. o r g O c t o b e r

44 The following article was published in ASHRAE Journal, October Copyright 2004 American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. It is presented for educational purposes only. This article may not be copied and/or distributed electronically or in paper form without permission of ASHRAE. Addendum 62n Single-Zone & Dedicated-OA Systems By Dennis Stanke, Member ASHRAE ANSI/ASHRAE Standard 62, Ventilation for Acceptable Indoor jair Quality, 1 has been modified by Addendum 62n 2, whose ventilation requirements alter the very heart of Standard 62. Addendum 62n contains a long-awaited update to the minimum prescribed ventilation rates last updated in ASHRAE Standard and it incorporates ventilation airflow additivity for dilution of both people-source and building-source contaminants. The updated version of the once-familiar table of prescribed breathing-zone ventilation rates now contains both per-person and per-unit-area values for each occupancy category. Addendum 62n updates the calculation procedure for zone ventilation airflow, incorporating an adjustment for air distribution effectiveness. It also updates the calculation procedure for system intake airflow for different ventilation systems and clarifies the prescribed approach for multiple-zone system design required for years but also widely misunderstood and largely ignored by designers. (Incorrect intake calculations often result in multiple-zone recirculating systems that * Although Addendum 62n shows ventilation rates in both IP and SI units, this paper uses IP, except in selected specific calculations. This is because 62n uses rational conversions, not mathematical. provide too little ventilation especially for some fully occupied VAV zones.) Finally, the addendum specifically identifies some operational control options that can reduce (or increase) intake airflow to match ventilation capacity with a changing ventilation load, saving preconditioning energy while maintaining the required ventilation to the occupants. What Are the New Ventilation Rates? Table 1 shows minimum breathing-zone ventilation rates* for several important occupancy categories, comparing the prescribed rates from Addendum 62n, Table 6.1, with the previous rates from Standard The minimum cfm/person rate (R p ) dropped for many categories (except for some retail categories, where it increased from the previously prescribed zero per-person rate) because the ventilation basis for people-source contaminants changed from satisfying unadapted visitors to satisfying adapted occupants. However, the addition of a minimum cfm/ft 2 rate About the Author Dennis Stanke is a staff applications engineer with Trane, La Crosse, Wis. He is vice chair of SSPC A S H R A E J o u r n a l a s h r a e. o r g O c t o b e r

45 (R a ) for each category to dilute building-source contaminants moderates those drops in the effective per-person breathingzone rates (i.e., the sum of the people- and area-related rates, divided by zone population). But be careful of simple, general comparisons; new default occupant densities and new options for population averaging can result in significant changes to traditional ventilation rates. Underlying the prescribed rates, which only apply to nosmoking areas, is the premise that all other general requirements in the standard are met (i.e., that the drain pan drains, humidity is limited, filters are used, and so on). Of course, compliance has always meant meeting general requirements, but Standard and its addenda extend and clarify these requirements, moving many design decisions from good practice to mandatory. As in the past, Table 6.1 includes default occupant densities for each category, but as before, these default values should only be used if reasonable for the application or if the actual design density is unknown. What Is the New Procedure? Addendum 62n prescribes a step-by-step procedure for ventilation system design to help designers consistently find the minimum ventilation airflow required in the breathing zone, in the ventilation zone, and at the outdoor air intake. Zone Ventilation Calculations Addendum 62n includes the following three-step procedure, which results in correct application of the prescribed rates to each ventilation zone. 1. Referring to Table 1 (an excerpt from 62n, Table 6.1), look up both the peoplesource ventilation rate (R p in cfm/person) and the building-source ventilation rate (R a in cfm/ft 2 ). Establish the zone floor area and zone population. The latter is the largest (or average) number of occupants expected to occupy the zone during normal use. Using these values, solve Equation 6-1 (V bz = R p P z + R a A z ) to find the required outdoor airflow for the breathing zone. 2. Referring to Table 2 (an excerpt from 62n, Table 6.2) look up the default value for zone air-distribution effectiveness (E z ), which is based on selection and placement of supply diffusers and return grilles. This effectiveness value, which is similar to air-change effectiveness (described in ASHRAE Standard 129) for wellmixed spaces, allows the designer to account for any ventilation air (delivered by the diffusers) that bypasses the breathing zone. 3. Solve Equation 6-2 (V oz = V bz / E z ) to find the outdoor airflow required in the air supplied to each ventilation zone. This volume of outdoor air must be supplied for ventilation, regardless of heating or cooling airflow requirements. Let s use numbers to demonstrate how these three steps determine the needed outdoor airflow for a north-facing office area with overhead supply diffusers and return grilles. Referring to Table 1, the office space occupancy category requires a people outdoor air rate of R p = 5 cfm/person** and an area outdoor air rate of R a = 0.06 cfm/ft 2. If the office comprises 1,500 ft 2 of open-plan space and is occupied by 10 people (150 ft 2 /person) solving Equation 6-1 (V bz = = = 140), tells us that at least 140 cfm (67 L/s) of outdoor air must be delivered to the breathing zone. The default zone air-distribution effectiveness for this configuration is 1.0 when cooling and 0.8 when heating. Now, we can solve Equation 6-2 and learn that our example office space requires at least (V oz = V bz / E z = 140/1.0 = 140 cfm) 14 cfm/person of outdoor airflow when cooling and at least (V oz = 140/0.8 = 175 cfm) 17.5 cfm/person when heating. Since 1989, Standard 62 has required 20 cfm/person for office-space breathing zones, so this particular office space needs less outdoor air per person to comply with 62n. However, lower occupant density might require more airflow per person, and higher occupant density (common in private offices) would require much less airflow per person. If our example was an executive office, designed for five people (300 ft 2 /person), it would require 115 cfm or 23 cfm/person (10.9 L/s per person), but if it was an technical professional office, designed for 20 people (75 ft 2 /person), it would require 190 cfm or only 9.5 cfm/person (4.6 L/s per person). System Ventilation Calculations After determining outdoor airflow (V oz ) for each ventilation zone, outdoor air intake flow (V ot ) is calculated for the ventilation system as a whole. The procedure for finding the required outdoor air intake flow varies with the configuration of the ** cfm = L/s; ft = m 2 O c t o b e r A S H R A E J o u r n a l 1 3

46 Occupancy Category People Outdoor Air Rate , Table 2 62n, Table 6.1 Per Person Rate Area Outdoor Air Rate People Outdoor Air Rate Area Outdoor Air Rate Default Occupant Density n Rp Ra cfm/person cfm/ft 2 cfm/person cfm/ft 2 #/1000 ft 2 cfm/person cfm/person Office Space Conference/Meeting Art Classroom Classroom (Ages 5 8) Classroom (Ages 9+) Lecture Classroom Multiuse Assembly Retail Sales Table 1: Comparison of breathing zone ventilation rates for several occupancy categories. ventilation system. Addendum 62n defines three configurations: single-zone, 100% (or dedicated ) outdoor air, and multiplezone recirculating systems. Single-Zone Systems. In a single-zone ventilation system, one air handler supplies one ventilation zone with a mixture of outdoor air and recirculated return air. Single-zone rooftop units, packaged terminal air conditioners, classroom unit ventilators, and so on, are single-zone systems. For these systems, Addendum 62n defines required outdoor air intake flow as equal to the required zone outdoor airflow according to Equation 6-3 (V ot = V oz ). Apparently for simplicity, this equation does not account for supply-duct leakage, which would increase V ot, or recirculation of outdoor air that bypasses the breathing zone, which would reduce V ot. Dedicated Outdoor Air Systems. In a dedicated outdoor air system (DOAS), called a 100%-outdoor air system in Addendum 62n, one air handler serves the ventilation requirements of one or more ventilation zones, delivering the appropriate minimum outdoor airflow without recirculated return air to each zone. Terminal units (e.g., fan-coil units, water-source heatpumps, or even chilled ceiling panels) handle the thermal loads Air Distribution Configuration Ceiling Supply of Cool Air Ceiling Supply of Warm Air and Floor Return Ceiling Supply of Warm Air At Least 8 C (15 F) Above Space Temperature and Ceiling Return Ceiling Supply of Warm Air Less Than 8 C (15 F) Above Space Temperature and Ceiling Return Provided That the 0.8 m/s (150 fpm) Supply Air Jet Reaches to Within 1.4 m (4.5 ft) of Floor Level. within each zone. (Some designers call this type of ventilation system a hybrid system, because it comprises a dedicated unit for ventilation air-handling and separate terminal units to handle thermal loads in occupied zones.) Like single-zone systems, all outdoor air entering the central air handler (assuming negligible duct leakage) reaches the ventilation zone diffusers. The required intake airflow is defined as the sum of the zone outdoor airflow values, according to Equation 6-4 (V ot = V oz ). Again for simplicity, the equation does not account for supply-duct leakage, which E z Table 2: Zone air distribution effectiveness (E z ) for several configurations. Not all configurations from Table 6.2 are listed here Making Sense of Additivity ASHRAE Standard 62 specifies minimum ventilation rates that are intended to result in indoor air that s free of harmful concentrations of known contaminants and that satisfies the senses of at least 80% of the occupants. (Though the rates prescribed in the standard seem to be based primarily on dilution of odors and irritants, they are presumed to be sufficient to adequately dilute potentially harmful contaminants as well.) Occupant satisfaction relates to the perceived intensity of odors and/or irritants from various indoor contaminant sources. These contaminants originate both from occupants (and their activities) and from the building (and its furnishings). While the relationships are complex, most experts agree that adding the outdoor airflow needed to dilute one odor or irritant to that needed to dilute another generally is the best simple model for dilution of odor and irritation effects. Accounting for the additive effect of contaminant sources really isn t new. Since 1989, the standard did so behind the scenes: Dilution rates for building-related contaminants were added to the per-person dilution rate for each occupancy category. For example, the standard previously required 20 cfm per person for offices: 15 cfm to dilute people-related odors and an additional 5 cfm to dilute building-related odors. Using Equation 6-1 of Addendum 62n, engineers now can independently account for people-related and building-related contaminants using two ventilation rate requirements: one rate per occupant (cfm/person) and the other per unit of occupiable floor area (cfm/ft²). To determine the required ventilation, simply multiply the per-person rate by the number of people in the space and the per-unit rate by the floor area; then, add the resulting airflow values together. 1 4 A S H R A E J o u r n a l a s h r a e. o r g O c t o b e r

47 Cooling Heating Ventilation Zone People Outdoor Air Rate Zone Population Area Outdoor Air Rate Zone Floor Area Zone Ventilation Efficiency Zone Outdoor Airflow Zone Ventilation Efficiency Zone Outdoor Airflow Rp Pz Ra Az Ez Voz Ez Voz cfm/person cfm/ft 2 ft 2 cfm cfm South Offices West Offices North Offices East Offices Interior Offices ,000 2,000 2,000 2,000 20, , ,125 North Conference Room * , South Conference Room ** , Total Zone-Level Outdoor Airflow V oz = 3,150 V oz = 3,830 Single-Zone Systems: Total Intake Air V oz = 3, %-Outdoor-Air System V ot = 3,830 * Average population (72% of 20-person peak population) ** Average population (77% of 30-person peak population) Table 3: A ventilation design example with calculations for a small office building. would increase V ot. And, even though Equation 6-4 increases V ot to account for breathing-zone bypass (when E z < 1.0), non-recirculating systems cannot reuse this bypassed air or any unused outdoor air from partially occupied zones; as a result, in some cases, dedicated outdoor air systems may actually require more intake airflow than multiple-zone systems. Multiple-Zone Recirculating Systems. In multiple-zone recirculating systems, such as constant-volume reheat systems and nearly all varieties of VAV systems, one air handler supplies a mixture of outdoor air and recirculated return air to two or more ventilation zones. The required outdoor air intake flow can only be determined by properly accounting for system ventilation efficiency. The reason for this is because the intake airflow must be sufficient to ventilate the critical zone, which is the zone that requires the highest fraction of outdoor air in its primary airstream. Since a multiple-zone system delivers the same primary air mixture to each ventilation zone, proper minimum ventilation in the critical zone overventilates all other zones. As a result, some unused outdoor air recirculates (reducing required V ot ) while some leaves the building via the relief, exhaust, and exfiltration air streams (increasing required V ot ). Addendum 62n recognizes this behavior and accounts for it by incorporating system ventilation efficiency in the ventilation calculations. System ventilation efficiency (E v ) can be determined using the default maximum values found in Addendum 62n, Table 6.3. These default values are based on the critical-zone ventilation fraction, found using Equation 6-5 (Z p = V oz /V pz ). Alternatively, system ventilation efficiency can be determined using the more accurate calculation procedure found in normative Appendix G of Addendum 62n. In either case, having found E v, outdoor airintake flow (V ot ) for the multiple-zone system must be determined using Equations 6-6, 6-7 and 6-8. Equation 6-6 (V ou = D R p P z )+ R a A z )) establishes the required uncorrected outdoor air-intake flow while allowing the designer to account for system population diversity determined with Equation 6-7 (D = P s / P z ), provided breathing-zone outdoor airflow is based on peak (not average) zone population. Equation 6-8 (V ot = V ou /E v ) finds the minimum outdoor air-intake flow by dividing the uncorrected outdoor air intake flow by the system ventilation efficiency. Ventilation Design Examples To demonstrate the use of the design procedure prescribed in Addendum 62n, this article considers the design of two ventilation systemsa single-zone system and a 100%-outdoor air system. Each system is applied in two different building types: an office and a school. (Future articles will discuss the design procedure for multiple-zone ventilation systems.) An Office Building Let s review the design of a ventilation system for a small office building that contains the ventilation zones described in Table 3. We began by looking up the outdoor air rates (R p and R a ) in Table 6.1 and then used Equation 6-1 to find the breathing-zone outdoor airflow (V bz ) for each zone. In our example, we used the peak population as the expected occupancy for each office area. For the conference rooms, however, we elected to use the optional population-averaging approach described in Section We used Table 6.1 default occupant densities for some zones and significantly higher densities (which may be more typical in individual offices) for others. According to a recent industry report, 3 average occupant density for some office workers exceeds 15 people per 1,000 ft 2 significantly higher than the default density of five people per 1,000 ft 2 in Table 6.1. Next, we established the zone air-distribution effectiveness (E z ) O c t o b e r A S H R A E J o u r n a l 1 5

48 for each zone based on the supply-air distribution configuration and the values shown in Table 6.2 (Table 2). Although the use of these values appears to be mandatory, it s reasonable to consider them as defaults, which are to be used whenever actual values are unknown. Higher-than-default values decrease required intake airflow and should be used cautiously (when justified by measurement or experience), since underventilation could result. Lower-thandefault values, which should be used if more-than-typical bypass is expected, increase required intake airflow and may increase first cost and operating cost. For instance, if the ceiling-mounted supply diffusers and return grilles are tightly spaced, it may be reasonable to assume a lower E z value during cooling, as we did for the north offices and the conference rooms in our example. In our example office, the zone air-distribution effectiveness for heating is less than for coolingi.e., E z = 0.80 (or less) vs That s because our design uses ceiling supply and return, and delivers supply air that s warmer than 90 F (32 C) during heating operation. The lower E z value for heating accounts for the tendency of warm supply air to float above the cooler, denser breathing-zone air. We then used Equation 6-2 to find each zone s minimum outdoor airflow (V oz in Table 3) during cooling and heating. At this point, we knew the zone-level outdoor airflow requirements, but how much outdoor airflow must enter the building at the intake(s)? The answer varied with the type of ventilation system selected. Single-Zone System Design. Initially, we assumed that each ventilation zone was served by a single-zone, constant-volume rooftop unit, making it necessary to use Equation 6-3 to find the outdoor air intake flow (V V ot ) required at each rooftop unit. Because we also assumed that all intake air reaches the supply diffusers (no significant duct leaks), the minimum intake airflow for each rooftop unit equals the minimum zone outdoor airflow (V V ot = V ) for the oz ventilation zone it serves. As shown in Table 3, the highest intake Averaging Zone Population for Ventilation Design In earlier versions of the standard, only intermittent occupancy zones (at peak population for three hours or less) could be designed for ventilation at the average population (but not less than one-half of peak population). Now, any zone may be designed d for average population. According to Addendum 62n, Section 6.2.5, the system must be designed to deliver the required outdoor airflow to each occupied breathing zone, but the design may be based on averages, in some cases. For instance, if occupancy (or supply airflow or intake airflow) varies, ventilation system design may be based on average population over a specific time period rather than on peak population. The averaging time T for a given zone is determined according to Equation 6-9, using zone volume and the breathing-zone outdoor airflow that would be needed at peak population. This averaging concept replaces the traditional populationaveraging approach for intermittent occupancy. Why averaging time T? The indoor concentration of any contaminant can be modeled using a first order ordinary differential equation. As most engineers vaguely remember, the solution to such an equation has the form ( ). Space time-constant is space volume divided by the outdoor airflow rate ( = v/ v V o ). In response to a step-change in contaminant sources, the concentration in the space rises exponentially, reaching 95% of its steady state value after three time constants. So, Equation 6-9 finds a three-time-constant averaging period, giving the zone a reasonable chance to respond to various short-term conditions in the space. Averaging time may be applied to make design adjustments when changing conditions in the zone can be predicted. For instance, if fluctuations in zone population can be predicted, design breathing zone outdoor airflow may be calculated based on the highest average population over any T hour period. Table 4 shows estimated population profile, averaging time and average population for several of the example zones used in this article. A cautionary note: overly aggressive zone-population averaging applied with the new, lower prescribed breathing-zone rates can sometimes lead to very low intake rates and inadequate ventilation. Ventilation Zone Example Offi ce North Conference Room South Conference Room Art Classroom Multiuse Assembly Area Zone Floor Area A z ft 2 1,500 2,000 3,000 2,000 3,000 Zone Pop. P z People Breathing- Zone Outdoor Airflow V bz cfm Ceiling Height ft Averaging Time T hr 5.5* Percentage of Peak Population by Time Interval Operating Time Blocks * For example, T = 3 v/ v V bz = /140 = 321 minutes or about 5.5 hours. ** For example, by inspection, highest average population as a percentage of peak = ( )/5.5 = 91%, so P z -avg = 91 10/ 100 = 9.1 people. Table 4: Estimated population profile, averaging time, and average population for several of the example zones used Avg. 91** Avg. Zone Pop. People A S H R A E J o u r n a l a s h r a e. o r g O c t o b e r

49 airflow needed for each zone occurs during heating operation. The resulting total outdoor airflow for the system is 3,830 cfm. Heating operation requires more intake airflow than cooling because we assumed that the supply-air temperature during heating is quite high, and that the discharge velocity of the diffusers is quite low. If each rooftop unit has only one minimum setting for the outdoor air damper, we would set it for the higher minimum airflow and size the heating and cooling coils for the corresponding percentage of outdoor air. If the rooftop can accommodate two minimum settings, however, we could reduce intake airflow to the lower value during cooling operation, thereby reducing the percentage of outdoor air and the required cooling coil capacity. An even better approach might be to simply lower the supplyair heating temperature to 90 F (32 C) or less while increasing the diffuser discharge velocity. According to Table 6.2, this would raise the maximum zone air-distribution effectiveness to 1.0 and reduce the heating intake airflow for each zone to match the cooling intake airflow. The benefits include less intake airflow year-round, only one minimum outdoor air-damper setting, and increased comfort since reduced discharge temperature and longer throws result in less vertical temperature stratification. Dedicated Outdoor Air-System Design. To determine the effect of a different ventilation system on outdoor air intake flow, we replaced the single-zone systems for our office with a dedicated (100%) outdoor air system. In this case, we assumed each ventilation zone is served by a single-zone, constant-volume heat pump (with no outdoor air intake). We also assumed all outdoor air is delivered directly to the ceiling mounted heat pumps from a central, constant-volume, dedicated outdoor air unit. This dedicated unit preconditions the outdoor air and delivers it through a ventilation duct system to each heat pump. Each heat pump, in turn, delivers both outdoor air and locally recirculated air from the plenum to the ventilation zone it serves. For this system configuration, we found outdoor air intake flow (V ot ) at the dedicated outdoor air unit using Equation 6-4. Assuming negligible duct leakage, all intake air reaches all supply diffusers, so minimum intake airflow simply equals the sum of the minimum zone-outdoor airflow values. In this example, we assumed that the preconditioned outdoor air mixes with local return air, and that the mixture is then cooled or heated by the heat pump before it enters the ventilation zone. In this configuration, the highest of the heating or cooling V oz value for each zone must be used to find V ot. With the values for zone air-distribution effectiveness shown in Table 3, and assuming that all zones are in heating, as might be the case in cold weather for the first few hours of operation, we found that the dedicated outdoor air unit must handle 3,830 cfm of outdoor air. In a system with several ventilation zones, it might seem reasonable that accounting for system population diversity could lower the intake requirement. This is not the case, however. Fluctuations in zone population can be incorporated (via Equations and Variables from Addendum 62n [6-1] V bz = R p P z + R a A z [6-2] V oz = V bz /E z [6-3] V ot = V oz single-zone systems [6-4] V ot = V oz 100% outdoor-air systems [6-5] Z p = V oz /V pz [6-6] V ou = D allzones R p P z + allzones R a A z = D allzones V bzp + allzones V bza [6-7] D = P s / P allzones z [6-8] V ot = V ou /E v multiple-zone recirculating systems [6-9a] T = 3v/V bz IP version [6-9b] T = 50v/V bz SI version where A z is zone floor area, the net occupiable floor area of the zone, ft² (m²) D is occupant diversity, the ratio of system population to the sum of zone populations E v is ventilation efficiency of the system E z is air-distribution effectiveness within the zone P s is system population, the maximum simultaneous number of occupants in the area served by the ventilation system P z is zone population, the largest expected number of people to occupy the ventilation zone during typical usage (See caveats in Addendum 62n Section ) R a is area outdoor air rate, the required airflow per unit area of the ventilation zone determined from Addendum 62n Table 6.1, cfm/ft² (L/s m²) R p is people outdoor air rate, the required airflow per person determined from Addendum 62n Table 6.1, in cfm/person (L/s person) T is averaging time period, minutes v is ventilation-zone volume, ft³ (m³) V bz is breathing-zone outdoor airflow, the outdoor airflow required in the breathing zone of the occupiable space(s) of the ventilation zone, cfm (L/s) V ot is outdoor air intake flow, adjusted for occupant diversity and corrected for ventilation efficiency, cfm (L/s) V ou is the uncorrected outdoor air intake flow, cfm (L/s) V oz is zone outdoor airflow, the outdoor airflow that must be provided to the zone by the supply-air-distribution system at design conditions, cfm (L/s) V pz is zone primary airflow, the primary airflow that the air handler delivers to the ventilation zone; includes both outdoor air and recirculated return air Z p is zone primary outdoor air fraction, the fraction of outdoor air in the primary airflow delivered to the ventilation zone for VAV systems, Z p for design purposes is based on the minimum expected primary airflow, V pzm. O c t o b e r A S H R A E J o u r n a l 1 7

50 Ventilation Zone People Outdoor Air Rate Zone Population Area Outdoor Air Rate Zone Floor Area Cooling Zone Ventilation Efficiency Zone Outdoor Airflow Heating Zone Ventilation Efficiency Zone Outdoor Airflow South Classrooms (Age 9+) West Classrooms (Age 9+) North Lecture Classrooms East Lecture Classrooms Interior Offices North Art Classroom South Multiuse Assembly Total Zone-Level Outdoor Airflow Single-Zone Systems: Total Intake Air 100% Outdoor Air System R p cfm/person P z * 276** R a cfm/ft * Average population (81% of 40-person peak population) ** Average population (92% of 300-person peak population) Table 5: Ventilation calculations for example school building served by single-zone system and 100% outdoor air system. A z ft 2 4,000 4,000 4,000 4,000 1,000 2,000 3,000 E z V oz = V ot = V oz cfm 1,880 1,880 2,190 2, ,250 11,200 11,200 E z V oz = V oz = V oz cfm 2,350 2,350 2,190 2, ,250 12,300 12,300 averaging) into zone ventilation calculations, but population diversity for the entire system cannot. In other words, even though the system as a whole never contains the sum-of-zones design population, each zone must be ventilated as though it is occupied at design (peak or average) population. A dedicated outdoor air system does not receive credit for recirculated outdoor air from over-ventilated zones, and typically does not modulate ventilation airflow to the zones served. When determining intake airflow for the dedicated outdoor air unit, each zone must be considered to be at design occupancy. A School Building Next, let s review two ventilation system designs for a different type of building: a school that contains the ventilation zones described in Table 5. As before, we followed the three-step procedure in 62n to find the zone-level ventilation requirements. To do so, we first used Equation 6-1 and Table 6.1 to find V bz for each zone. For zone population, we used the expected peak occupancies for the classrooms and office areas, but we used average population for the art classroom and multiuse assembly area. Second, we established the zone air-distribution effectiveness for each zone configuration, based on the values in Table 6.2 (see Table 2). We assumed that supply air during heating is less than 90 F (32 C), but we also assumed that the classroom and office diffusers do not provide sufficient velocity to deliver 150 fpm (0.75 m/s) air at the 4.5-ft (1.4 m) level. Therefore, we used E z = 1.0 for the lecture rooms and assembly area, but we used an E z value of 0.8 for the classrooms and offices during heating operation. Third, we used Equation 6-2 to find the minimum outdoor airflow needed for each zone in both cooling and heating modes. Because the minimum required outdoor air intake flow depends on the type of ventilation system, we again looked at two system types. Zone Air-Distribution Effectiveness All supply air leaving the ventilation zone diffusers may not actually arrive in the breathing zone. During heating operation, for instance, some portion of warm air from an overhead diffuser may simply float on cooler air in the space and never drop into the breathing zone (between 3 and 72 in. from the floor). Or, during cooling operation, some portion of cool air from an overhead diffuser may short circuit to a nearby (poorly placed) return grille, never reaching the breathing zone. The fraction of the air supplied to a space that actually reaches the breathing zone can be characterized by zone air-distribution effectiveness (E z ). It is similar to air-change effectiveness (ACE) for well-mixed spaces, which can be determined in the lab using the methods in ANSI/ASHRAE Standard 129, Measuring Air Change Effectiveness. For laminar flow spaces, like some displacement and underfloor systems, ACE depends on ceiling height, so it is not necessarily a good indicator of the fraction of supply air reaching the occupants. (Both ACE and E z values may be high with a nine-foot [2.75 m] ceiling, but with a 30-foot [9 m] ceiling, ACE would be higher than E z, not equal to it; using an inappropriately high value [that is, E z = ACE ], would lead to underventilation when using Equation 6-2 to find zone outdoor airflow.) Note that air diffusion performance index (ADPI), defined in the ASHRAE Handbook and ANSI/ASHRAE Standard 113, Method of Testing for Room Air Diffusion, is not directly related to air-distribution effectiveness; a poorly performing diffuser that dumps cold air onto an occupant s head has a low ADPI but a high E z value. 1 8 A S H R A E J o u r n a l a s h r a e. o r g O c t o b e r

51 Single-Zone System Design. First, we assumed that each ventilation zone is served by a single-zone, constant-volume unit ventilator or rooftop unit. Equation 6-3 (V ot = V oz ) is used to find the outdoor air intake flow needed at each unit; it simply equals the minimum zone outdoor airflow. As shown in Table 5, some zones need more intake airflow in heating mode than in cooling because of lower zone air-distribution effectiveness. The total outdoor airflow for the system in this case is 12,300 Advertisement formerly in this space. cfm. As mentioned earlier, lowering supply air temperature when heating, and/or selecting better diffusers can reduce the required intake airflow (and unit capacity) as well as increase comfort by reducing vertical temperature stratification. Dedicated Outdoor Air-System Design. Next, we assumed that each ventilation zone is served by a single-zone, constantvolume blower-coil unit (or a recirculating air handler) without an outdoor air intake. Further, we assumed that all outdoor air enters the system through a central, constant-volume, dedicated outdoor air unit. The entering outdoor air is preconditioned, so it is dry and cool (or neutral during heating operation), and is ducted to and discharged directly into each zone through ventilation diffusers. Given the system configuration, we used Equation 6-4 (V ot = V oz ) to find the minimum required outdoor air intake flow at the dedicated outdoor air unit; it s simply the sum of the minimum zone-outdoor airflow values For this school example, preconditioned outdoor air enters each zone directly, and it is never hot. We ignored the somewhat higher heating values and instead used the cooling V oz values, which means that the dedicated outdoor air unit must handle 11,200 cfm of outdoor air, a little less total intake airflow than we needed using single-zone recirculating systems. As in the previous example, we cannot account for system population diversity in this constant-volume system, even though peak population does not occur in all zones simultaneously. What About Part-Load Operation? In the preceding discussion, we found the minimum design outdoor air intake flow for two different ventilation systems in an example office and an example school. Operationally, however, ventilation load varies with changes in supply-air temperature and population. Addendum 62n explicitly permits dynamic reset of intake airflow so ventilation capacity can be matched to ventilation load. This can save outdoor airconditioning energy during periods of low ventilation load and avoid underventilation problems during periods of high ventilation load. Because demand-controlled ventilation is a broad topic, it will be addressed separately in a future article. Suffice to say, single-zone systems may be cycled on and O c t o b e r A S H R A E J o u r n a l 1 9

52 off, or the intake damper minimum setting may be modulated in response to population changes. But, in dedicated outdoor air systems, the 100% outdoor air unit must deliver design ventilation air to all zones whenever any zone is occupied, unless individual zones include modulating dampers to control outdoor airflow based on zone ventilation demand. Note that the ASHRAE energy standard 4 requires energy recovery on 100% outdoor air units delivering 5,000 cfm (2360 L/s) or more; a dedicated outdoor air system with energy recovery may use less energy than multiple single-zone systems without energy recovery. Anything Else? For zones with strong contaminant sources, Addendum 62n prescribes minimum exhaust airflow to remove contaminants from the building. Minimum exhaust rates for nineteen zones are prescribed in Table 6.4 (not included here). This may be regarded as a significant change from Standard , which specified exhaust rates for only a few zones and did not clarify whether it was just minimum exhaust flow, or both minimum exhaust and outdoor air supply flow (e.g., public restrooms). While most zones require either one or the other, some zones such as art classrooms require both minimum supply and exhaust airflow. Makeup air from outdoors (to replace exhausted air) may be supplied to a zone by any combination of first-pass outdoor air or outdoor air transferred from other zones. Summary The Addendum 62n heart transplant certainly changes the ventilation standard. There are no guarantees, but these expectations seem reasonable: that the generally lower breathing-zone rates will reduce some designer complaints about overventilation, that people- plus building-ventilation additivity will reduce effective per-person ventilation in high-density spaces but increase it in low-density spaces, and that the prescribed calculation procedures will reduce underventilation in some systems while increasing calculation consistency among designers. By changing the prescribed rates and at the same time encouraging better design procedures, Addendum 62n may lower ventilation system costs for some systems and improve indoor air quality in others. Time will tell. References 1. ANSI/ASHRAE Standard , Ventilation for Acceptable Indoor Air Quality. 2. ANSI/SHRAE Addendum n to ANSI/ASHRAE Standard International Facility Management Association Project Management Benchmarks Report. In IFMA Research Shows Office Space Shrinking. Cited Aug. 20, Available from: org/about/prdetail.cfm?id=331&actionbig=3&actionlil= ANSI/ASHRAE Standard , Energy Standard for Buildings Except Low-Rise Residential Buildings. Advertisement formerly in this space.

53 O 2 The following article was published in ASHRAE Journal, December Copyright 2008 American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. It is presented for educational purposes only. This article may not be copied and/or distributed electronically or in paper form without permission of ASHRAE. CO 2 Criteria for Outdoor Air Monitoring Selecting By Thomas M. Lawrence, Ph.D., P.E., Member ASHRAE provided. The California Title Building Energy Efficiency Standards The operation of buildings is as, or is more important than, the original specifies a maximum CO 2 concentration of 600 ppm above ambient as a one size design in maintaining a healthy indoor environment and energy efficiency. fits all criteria for spaces using demand ventilation controls. However, this only A number of certification programs and standards specify the use of ventilation applies to California because it has its own ventilation standard, which is not system performance monitoring. Monitoring is to be done either based on CO 2 based on Standard Decisions must be made regarding concentrations in the occupied space or actual measurement of outdoor airflow, the balance between having adequate outdoor air ventilation to the occupied depending on the space design occupancy and ventilation type (mechanical or space while still considering the energy natural). For example, outdoor air delivery monitoring is a credit option for new requirements needed to condition that CO ventilation air. Section in ANSI/ construction, existing buildings and the core and shell LEED programs from ASHRAE/IESNA Standard , Energy Standard for Buildings Except the U.S. Green Building Council. Low-Rise Residential Buildings, describes the allowances for dynamic reset The current standards or program tion for Acceptable Indoor Air Quality, of outdoor air ventilation for densely descriptions do not provide detailed guidance for determining what concentration concentrations are. LEED-NC, version published many articles during the past but does not list what the actual 2 occupied spaces. ASHRAE Journal has of CO 2 should be considered the maximum concentration. The LEED-EB (for (EQ Credit 1) for monitoring of 2 control (demand-controlled ventilation 2.2, (new construction) offers a credit decade that deal with outdoor airflow existing buildings) credit mentions CO 2 concentrations in mechanically ventilated concentrations 15% above that expected densely occupied ( 25 people per 1,000 to occur with a corresponding minimum ft² [100 m 2 About the Author ]) spaces, and in all naturally ventilated spaces. Again, no spe- service associate with the Faculty of Engineering at Thomas M. Lawrence, Ph.D., P.E., is a public outdoor airflow rate required by ANSI/ ASHRAE Standard , Ventilacific maximum CO 2 concentrations are the University of Georgia in Athens, Ga. 18 ASHRAE Journal ashrae.org December 2008

54 or DCV) using CO 2. These include discussions on demandcontrol ventilation design guidelines, 1 applying supply air sampler to determine CO 2 levels. While locating the CO 2 sensor zones at the required locations and uses a central monitoring CO 2 control, 2 retrofitting CO 2 control in existing buildings, 3 directly in the zone usually works sufficiently for the purposes the potential sustainability benefits for CO 2 DCV, 4 and the risk of CO 2 monitoring, using a centralized sampling system may versus rewards involved. 5 Various options beyond occupancy sensing using 2 are possible as well for outdoor airflow rate control and dynamic reset. 6 The existence of differing criteria, or the lack of guidance on the exact criteria to use for determining the upper limit of CO 2 CO be more accurate, especially if the monitoring is based on comparing 2 concentrations, such as breathing zone levels compared to the outdoor air. Another consideration is the 2 sensor accuracy level and if it is sufficient for use in zone monitoring. Typical cut sheet in a monitoring program or for a DCV system can lead accuracies range between 50 and 100 ppm at 2,000 ppm. At the to confusion. It is generally thought of as the responsibility of lower range of 50 ppm, consider the concentration error for two the design engineer to define the upper allowed limit. However, sensors (zone and outside air) in a space designed for 15 cfm this may not always be done. In an existing building, the building operator may be forced to make a decision on what is the between zone and outside air). This net 100 ppm potential error (7 L/s) per person (with a corresponding 700 ppm differential highest acceptable concentration of 2. is roughly equivalent to an error in outdoor airflow monitoring This article is intended to provide overall guidance and direction on what should be considered a too high concentration of ppm accuracy were used in the same situation, the equivalent to of ±2 cfm (1 L/s) of outdoor air per person. If sensors with 75 CO 2 in a given space undergoing a 2 -based outdoor air monitoring program. The results also can be used to help determine Advances in 2 sensor technology over the past decade now error in outdoor airflow monitoring would be ±3 cfm (1.5 L/s). the upper control limit for a system using DCV. The guidance is allow for less frequent recalibration, e.g., once every five years. built upon the fundamental concepts described in key ASHRAE Some devices also are advertised to conduct self-calibration, documents, such as Standard 62.1, the corresponding User s but this technique is still relatively new. Manual and the ASHRAE Handbooks, with added engineering Questions can be legitimately raised as to whether a space judgment to focus on the key aspects needed for a 2 -based can ever be assumed to be fully mixed, and at steady-state with outdoor air monitoring program. respect to 2 concentration. Whether or not the space can 2 be considered fully mixed is related to the zone air-distribution Indoor Air Quality Monitoring Options effectiveness. Values for this can be found in Standard The two most common options used for monitoring indoor 2007, Table 6-2. Under constant occupancy conditions, the time air quality of a space are the direct measurement of the total outdoor airflow and the monitoring of CO 2 concentrations. The direct measurement of outdoor airflow may appear to be conceptually straightforward, but its application in practice can be technically difficult with issues such as low intake airflow velocities, flow through exhaust dampers, or measuring low differentials in velocity pressures. 7,8 CO 2 monitoring is commonly used as an indirect indication of indoor air quality problems due to insufficient outdoor air ventilation rate, but it certainly is not without the potential for error and misinterpretation. One reason is that it neglects pollutants not associated with occupants, such as off-gassing from building materials and furnishings, making it only an indirect method of determining adequate ventilation. needed for a building or room to reach equilibrium is a function of the air change rate (assuming the occupancy, ventilation rate and outdoor concentration remain steady as well). A time period with stable occupancy and ventilation rate equal to three times the room time constant is required for room CO 2 concentrations to reach 95% of their steady-state value. 9 In a room with a low air change rate, it may take up to 12 hours of constant conditions for true equilibrium conditions to be reached. 10 Persily 11 specified the following criterion for a building to be considered in equilibrium: Where The conversion constant ( ) is based on SI units with: C = change in indoor-outdoor concentration in one hour, Key Issues With CO 2 Monitoring, Equipment and Systems When developing a CO 2 -based outdoor air monitoring mg/m³ h program, a number of issues need to be considered. These G = CO 2 generation rate in the zone, L/s include how and where to measure the CO 2 concentrations; V = Volume of the zone, L what type of sensing system to use; the required accuracy level of the sensor(s); calibration requirements for whatever devices For example, consider a typical K 12 school classroom with are chosen; and whether to base the evaluation on absolute assumed default occupancy of 35 persons per 1,000 ft² (100 m²) concentrations of CO 2 or on the differential with respect to and a total CO 2 generation rate of 0.18 L/s (0.38 cfm) based on the ambient air. The type of ventilation system being used can the data shown in Figure C-2 of Standard Assuming influence these choices. a room ceiling height of 9 ft (3 m), the resulting typical room A CO 2 sensor need not be located within the zone itself; rather volume is 30,000 ft³ ( L). Using this equation, the a system could be chosen that draws samples of air from the room could be considered in equilibrium if the change in CO 2 December 2008 ASHRAE Journal 19 (1)

55 Determining the Ambient CO 2 Concentration The general worldwide average CO 2 concentration is around 380 ppm, 16 which varies by season and location, but the trend is upward. The concentration is lower over marine surface sites and higher over land, especially in towns and cities. Therefore, a simple assumption for CO 2 would be 400 ppm, which is conservative (higher than actual) in most building locations. Alternatively, the actual ambient concentration surrounding the building could be measured. This is particularly recommended for sites located in larger urban settings or near major roadways, where local CO 2 concentrations can easily be several hundred ppm higher than 400 ppm. If two sensors are used to determine the CO 2 concentration difference, sensor accuracy becomes important, especially when the indoor concentration approaches the outdoor (or supply) concentration at low occupancy levels. concentration was less than 108 ppm by volume (194 mg/m³ for CO 2 ) over a one-hour period. In practice, most spaces do not remain steady in occupancy and airflow for long enough time periods to reach full steadystate equilibrium of CO 2 concentration. However, for simple and indirect checking of outdoor airflow per person, it is not necessary for a space to be in complete equilibrium for the values measured to give at least a good indication of the current apparent rate of outdoor airflow with respect to occupancy. Related to that, Lawrence and Braun showed that a steady-state analysis was sufficient for evaluating a building for potential CO 2 -based DCV retrofits. 12 When considering decisions that are broader in nature, such as is this concentration of CO 2 too high, which implies that the rate of outdoor airflow reaching the breathing zone is too low for the occupancy, then simplifying assumptions can be made. We must keep in mind that the actual purpose of CO 2 monitoring is to provide an indication of the level of bioeffluents in the space, which is what we are trying to control using ventilation. Taylor 13 indicated, in the sidebar titled Assumptions of Steady-State Conditions, that the generation of bioeffluents should closely mimic the rate of generation of CO 2 since both are generated at a rate related to the number of people in the space and their activity level. Determining Expected CO 2 Concentrations Standard specifies minimum outdoor air ventilation rates in the breathing zone based on the zone occupancy category. Table 6-1 in this standard provides the required values per person and per floor area for the different occupancy types, and also includes a default value for the net combined ventilation rate per person using default occupancy values. Expressed as the flow rate per person, the default combined values can range widely, from 5 cfm (2.7 L/s) per person to 25 or more cfm (12.4 L/s). Thus, the normal expected CO 2 concentration at the design outdoor airflow rate and occupancy could vary widely as well, even at steady-state conditions, and a simple single CO 2 concentration above ambient conditions would not apply universally to all occupancy space categories. The ASHRAE Journal article by Taylor 13 and the 62.1 User s Manual, Appendix A 14 provide a good technical review of the calculations involved with determining required outdoor ventilation airflow and the resulting steady-state CO 2 concentrations for single-zone systems. While those articles provide a good derivation of the equations involved, the following is a brief summary of how steady-state room CO 2 concentration can be calculated from an overall mass balance of CO 2 into and out from the space. The rate of CO 2 generation and removal can be expressed on a per person basis. The rate of CO 2 generation is based on the activity level (m, or metabolic rate). 13,15 The rate of CO 2 removal is a function of the outdoor air ventilation rate being supplied to the breathing zone. Standard specifies that the outdoor ventilation air required is computed based on a component for the occupants, as well as for the building area, or: Where R p P z R a A z = ventilation rate per person = number of people in the zone = ventilation rate per building unit area = building area The combined outdoor ventilation air rate per person to the breathing zone can be found by dividing Equation 4 by the number of occupants. The amount of ventilation air to the breathing zone (V bz ) is determined based on the outdoor air rate supplied to the zone (V oz ) and the zone air distribution effectiveness (E z ), as shown by Equation 5. Using the CO 2 generation rate per person and the combined outdoor ventilation air rate per person, the steady-state room CO 2 concentration (in ppm) that will occur in the breathing zone at design occupancy and assuming the minimum outdoor (2) (3) (4) (5) 20 ASHRAE Journal ashrae.org December 2008

56 Table 1: Computed and Recommended CO 2 Concentrations for Outdoor Airflow Monitoring or DCV Upper Control Limit Default Values Combined Outdoor Air Rate Per Person Assumed Activity Level CO 2 Generation Actual Steady-State Concentration Monitoring Program Concentration (Alarm Level) DCV Upper Control Limit Concentration (Caution Level) LEED-EB IEQ Credit 1 Concentration Occupancy Category cfm L/s (met) * (cfm per person) (ppm) (ppm) (ppm), (ppm), Correctional Facilities Cell ,072 1, ,233 Dayroom ,840 2,000 1,656 2,116 Guard Stations ,333 1,500 1,200 1,533 Booking/Waiting ,333 1,500 1,200 1,533 Educational Facilities Day Care (Through Age 4) ,141 1,300 1,027 1,312 Day Care Sickroom Classrooms (Age 5 8) , ,104 Classrooms (Age 9+) ,046 1, ,203 Lecture Classroom ,450 1,600 1,305 1,668 Lecture Hall (Fixed Seats) ,450 1,600 1,305 1,668 Art Classroom , ,070 Science Laboratories , ,142 University/College Lab , ,142 Wood/Metal Shop ,284 1,400 1,156 1,477 Computer Lab ,072 1, ,233 Media Center ,072 1, ,233 Music/Theater/Dance ,800 1,900 1,620 2,070 Multiuse Assembly ,975 2,100 1,778 2,271 Food and Beverage Service Restaurant Dining Rooms ,576 1,700 1,418 1,812 Cafeteria/Fast-Food Dining ,707 1,900 1,536 1,963 Bars, Cocktail Lounges ,707 1,900 1,536 1,963 General Break Rooms ,408 1,600 1,267 1,619 Coffee Stations ,316 1,500 1,185 1,514 Conference/Meeting ,800 1,900 1,620 2, ASHRAE Journal ashrae.org December 2008

57 Table 1: Computed and Recommended CO 2 Concentrations for Outdoor Airflow Monitoring or DCV Upper Control Limit (Cont.) Default Values Combined Outdoor Air Rate Per Person Assumed Activity Level CO 2 Generation Actual Steady-State Concentration Monitoring Program Concentration (Alarm Level) DCV Upper Control Limit Concentration (Caution Level) LEED-EB IEQ Credit 1 Concentration Occupancy Category cfm L/s (met) * (cfm per person) (ppm) (ppm) (ppm), (ppm), Hotels, Motels, Resorts, Dormitories Bedroom/Living Area ,011 1, ,163 Barracks Sleeping Areas ,240 1,400 1,116 1,426 Laundry Rooms, Central ,388 1,500 1,249 1,596 Laundry Within Dwelling ,092 1, ,256 Lobbies/Prefunction ,660 1,800 1,494 1,909 Multipurpose Assembly ,500 2,600 2,250 2,875 Office Buildings Office Space , ,142 Reception Areas ,840 2,000 1,656 2,116 Telephone/Data Entry ,080 2,200 1,872 2,392 Main Entry/Lobbies ,545 1,700 1,391 1,777 Miscellaneous Spaces Bank Vaults/Safe Deposit , ,028 Computer (Not Printing) , Pharmacy (Preparation Area) , ,048 Photo Studios ,092 1, ,256 Transportation Waiting ,450 1,600 1,305 1,668 Public Assembly Spaces Auditorium Seating Area ,080 2,200 1,872 2,392 Place of Religious Worship ,080 2,200 1,872 2,392 Courtrooms ,080 2,200 1,872 2,392 Legislative Chambers ,080 2,200 1,872 2,392 Libraries , ,028 Lobbies ,920 3,100 2,628 3,358 Museums (Children s) ,545 1,500 1,391 1,777 Museum/Galleries ,800 1,700 1,620 2,070 December 2008 ASHRAE Journal 23

58 Table 1: Computed and Recommended CO 2 Concentrations for Outdoor Airflow Monitoring or DCV Upper Control Limit (Cont.) Default Values Combined Outdoor Air Rate Per Person Assumed Activity Level CO 2 Generation Actual Steady-State Concentration Monitoring Program Concentration (Alarm Level) DCV Upper Control Limit Concentration (Caution Level) LEED-EB IEQ Credit 1 Concentration Occupancy Category cfm L/s (met) * (cfm per person) (ppm) (ppm) (ppm), (ppm), Retail Sales (Except Below) ,188 1,300 1,069 1,366 Mall Common Areas ,800 1,900 1,620 2,070 Barbershop ,408 1,600 1,267 1,619 Beauty and Nail Salons Pet Shops (Animal Areas) Supermarket ,240 1,400 1,116 1,426 Coin-operated Laundries ,469 1,600 1,322 1,689 Sports and Entertainment Spectator Areas ,975 2,100 1,778 2,271 Disco/Dance Floors ,600 1,700 1,440 1,840 Health Clubs/Aerobics Room ,927 2,100 1,735 2,216 Health Clubs/Weight Room ,369 1,500 1,232 1,575 Bowling Alley (Seating) ,369 1,500 1,232 1,575 Gambling Casinos ,520 1,700 1,368 1,748 Game Arcades , ,142 Stages, Studios ,545 1,700 1,391 1,777 * From User s Manual or taken directly or estimated from 2005 ASHRAE HandbookFundamentals, Chapter 8, Table 4. Assumes 400 ppm ambient. Steady-state concentration 10%. Steady-state concentration +15%. ventilation air is provided can be determined by Equation 6. (6) and an activity level (m) of 1.5, the corresponding room CO 2 concentration would be: (7) This is the equivalent to the Equation 13 in the article by Taylor. 13 As an example, consider a typical museum gallery. Table 6-1 of Standard defines the default occupancy of 40 people per 1,000 ft² (100 m²), and a default combined outdoor air rate of 9 cfm/person (4.6 L/s per person) delivered to the breathing zone. Assuming an ambient CO 2 concentration of 400 ppm (see the sidebar What is the Ambient CO 2 Concentration? ), a zone air-distribution effectiveness of 1.0 CO 2 Concentration Difference Criteria for OA Monitoring Starting with the expected steady-state concentration of CO 2 for any given space, we can develop values that would be useful in a CO 2 monitoring program. It may be desirable for a building or system operator to have two sets of CO 2 concentrations. One value would be to use as a caution point, warning that the concentrations are approaching the generally expected maxi- 24 ASHRAE Journal ashrae.org December 2008

59 mum. A second, and higher, concentration would be the level used as an alarm point signifying a situation where something was wrong. The alarm point would be expected to be used as an upper limit as part of a CO 2 monitoring program, while the caution point might be used as an upper setpoint limit used in a CO 2 -based DCV system. Overall, four sets of values are derived and presented here, with each being useful for different situations. 1. Actual expected maximum steady-state CO 2 concentration. This is the value computed based on the combined outdoor air ventilation rate per person (as defined in Table 6-1 of Standard ) and the expected metabolic rate for the occupants in that space, as defined in either Appendix A of the 62.1 User s Manual or estimated using Table 4 in Chapter 8 of the 2005 ASHRAE Handbook Fundamentals as a guide. 2. Upper limit for a CO 2 monitoring program (alarm point). This value is the critical action level for CO 2 concentration in the space that would trigger action or at least be recorded as an incident in the monitoring program. The values proposed here are based on the expected maximum steady-state CO 2 concentrations (Situation 1) rounded upward to allow for sensor error and to give concentrations in hundreds of ppm for simplicity. 3. Upper limit for CO 2 -based DCV (caution point). This is the upper limit used in the control system algorithm that adjusts outdoor ventilation airflow based on measured CO 2 concentrations. The values proposed here are based on the expected maximum steady-state CO 2 concentrations (Situation 1) minus 10% to allow for control system response times, sensor inaccuracy, etc. Finding the right minimum CO 2 concentration to begin damper position movement can be tricky, but the damper is opened gradually as CO 2 values rise toward the maximum concentration setpoint, at which the outdoor air damper would be at the maximum open position. 4. Upper limit for a building undergoing LEED-EB monitoring. This is the maximum value allowable in a space if a building were trying to obtain the LEED-EB program IEQ Credit 1. The values proposed here are based on the expected maximum steady-state CO 2 concentrations (Situation 1) +15%, as specified in the LEED-EB credit description. (The discussion below compares the results of the LEED-EB values and the alarm point concentrations.) These four different sets of values are summarized in Table 1, and are based on the assumption that the ambient concentration is 400 ppm, there is a zone air-distribution effectiveness of 1.0, and the zone is at design occupancy and is provided with the minimum outdoor air ventilation rate. The table is organized using the various occupancy categories listed in Table 6-1 of Standard Discussion The three columns on the right side of Table 1 summarize the key recommendations and purpose of this article. The Researching Multizones This article focuses on the development for single-zone systems. The situation is more complex when considering multizone systems, as each zone contributes to the overall CO 2 balance and concentrations in the supply and return airstreams. For example, one method for calculation of the zone ventilation effectiveness for multizone systems is given in Appendix A of Standard The equations needed for the concentration equations can become complex since the concentration in one zone depends on what is going on in every other zone. A research project is being initiated by ASHRAE Technical Committee 4.3, which should help shed some light in this area. A monitoring program for a multizone system still would be checking for the amount of outdoor air entering the breathing zone for any individual space. The rough mass balance of the amount of outdoor air entering the breathing zone carrying away CO 2 generated in the space still exists. The recommended concentrations in Table 1 might be the basis for application to multizone systems, but this analysis and discussion is reserved for other articles and is awaiting the results from ASHRAE research programs. primary purpose is to present proposed room CO 2 concentrations for use in a CO 2 monitoring program. These values are presented as an aid to a building operator, manager or systems designer responsible for developing a CO 2 monitoring program, or, alternatively, in determining control limits for a CO 2 -based DCV system. The concentration criteria allow for a potential sensor error in the range of 50 ppm to 75 ppm. When conducting CO 2 monitoring for the space, if the room concentration exceeds the amount recommended in Table 1 for that particular occupancy type, then this would indicate a potential problem with the system that must be investigated. It is interesting to compare the recommend CO 2 monitoring program concentration limits (the action level) presented to the concentrations that are a set 15% above the steady-state values (the LEED-EB criteria for IEQ Credit 1). These concentrations are the same, with differences due to rounding to the nearest hundreds of ppm concentration for the recommended CO 2 monitoring program action levels. The results presented here also can be applied to CO 2 -based DCV systems. The upper limit of the control algorithm is recommended to be the steady-state room concentrations minus 10%, and these values are given in the second from right column in Table 1. The DCV control should have as an alarm point the concentrations listed as the CO 2 monitoring program action levels. Potential Responses if Recommended Criteria Are Exceeded The actual response taken if the CO 2 concentration difference is exceed in a space as part of an outdoor air monitoring program can vary. The criteria given in Table 1 would be used 26 ASHRAE Journal ashrae.org December 2008

60 as the maximum allowable values above which a signal to the system operator or other response would be initiated. The actual response taken if the CO 2 criteria are exceeded is up to the individual building owner or operator. A local alarm or building automation system signal, or both, could be sent, or only a record could be made that this event occurred for future diagnostic purposes. Summary The use of CO 2 monitoring as part of an overall outdoor air monitoring program is becoming more common in the operation of a modern building. To date, minimal practical guidance has been given as to what CO 2 concentrations would be considered the upper limit. This article presents a set of recommended maximum CO 2 concentration differences as a function of the design combined minimum outdoor airflow and the anticipated metabolic rate for the zone occupants that can be used as upper acceptable limits for an outdoor air monitoring fault detection program. Advertisement formerly in this space. References 1. Schell, M. and D. Int-Hout Demand control ventilation using CO 2. ASHRAE Journal 43(2): Warden, D Supply air CO 2 control. ASHRAE Journal 46(10): Schell, M. and D. Smith Assessing CO 2 control in retrofits. ASHRAE Journal 44(11): Lawrence, T Demand-controlled ventilation and sustainability. ASHRAE Journal 46(12): Dougan, D. and L. Damiano CO 2 based demand control ventilation: do risks outweigh potential benefits? ASHRAE Journal 46(10): Stanke, D Standard system operation: dynamic reset options. ASHRAE Journal 48(12): Fisk, W.J., D. Faulkner, and D.P. Sullivan Measuring OA intake rates. ASHRAE Journal 48(8): Kettler, J.P Controlling minimum ventilation volume in VAV systems. ASHRAE Journal 40(5): Seppänen, O.A., W.J. Fisk, and M.J. Mendell Associations of ventilation rates and CO 2 concentrations with health and other responses in commercial and institutional buildings. Indoor Air 9: Nabinger, S.J., A.K. Persily, and W.S. Dols A study of ventilation and carbon dioxide in an office building. ASHRAE Transactions 100(2): Persily, A.K Evaluating building IAQ and ventilation with indoor carbon dioxide. ASHRAE Transactions 103(2): Lawrence, T.M. and J.E. Braun Modeling of CO 2 concentrations in small commercial buildings. Buildings and Environment 41(2): Taylor, S CO 2 -based DCV using ASHRAE Journal 48(5): ASHRAE User s Manual ASHRAE HandbookFundamentals, Chapter 8, Thermal Comfort. 16. Solomon, S., et al Climate Change 2007, The Physical Science Basis: Contribution of Working Group I to the Fourth Assessment Report of the Intergovernmental Panel on Climate Change. Cambridge, U.K.: Cambridge University Press. December 2008 ASHRAE Journal 27

61 Application Issues Copyright 2004, American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. This posting is by permission of ASHRAE IAQ Applications. This article may not be copied nor distributed in either paper or digital form without ASHRAE s permission. Contact ASHRAE at Transient Occupancy Ventilation By Monitoring CO 2 Standard 62 has produced an industry devoted to utilizing ventilation reset, thus reducing the energy used to condition OA, based generally upon holding the space with the highest CO 2 concentration at or below about 1,000 ppm. That may change with Addendum n. By Stanley A. Mumma, Ph.D., P.E., Fellow ASHRAE ANSI/ASHRAE Standard , Ventilation for Acceptable Indoor Air Quality, has produced an industry devoted to utilizing ventilation reset, thus reducing the energy used to condition OA, based generally upon holding the space with the highest CO 2 concentration at or below about 1,000 ppm. For example, a space designed for 40 people where each person is to receive 20 cfm (9.4 L/s) has a required ventilation rate of 800 cfm (378 L/s). If the occupancy drops to 20 people, dynamic reset would reduce the ventilation rate to 400 cfm (189 L/s), reducing the quantity of OA to be conditioned and energy use by 50%. Standard states Indoor air quality shall be considered acceptable if the required rates of acceptable outdoor air in Table 2 are provided for the occupied space. Table 2 lists the required ventilation rates in cfm (L/s) per person or cfm per ft 2 (L/s per m 2 ) for a variety of spaces. Where the ventilation rates are in cfm per person, the contamination produced is presumed to be proportional to the number of persons in the space. ASHRAE is on the verge of replacing Table 2 and the Ventilation Rate Procedure of Standard with Addendum 62n, which has been approved by ASHRAE and awaiting ANSI approval. It recognizes that indoor air pollutants are generated by both building occupants and their activities as well as by the contents of a building. Consequently, the ventilation requirement in the breathing zone is the summation of an occupancy component and an area component. Mostly, the design occupancy total ventilation requirements will not be greatly changed by Addendum 62n. High occupancy spaces, such as conference rooms, are an exception. The ventilation rate for such spaces generally will be less with Addendum 62n. Dynamic Reset of Outdoor Air Intake Section of 62n permits dynamic reset of the design outdoor air intake flow during several conditions. One condition is an estimate of occupancy or ventilation rate per person using occupancy sensors such as those based on indoor CO 2 concentrations. A literal reading would require the space CO 2 concentration to go down as the occupancy decreases. Consider a classroom for 25 students (ages 9 or older) that has a floor area of 1,050 ft 2 (98 m 2 ). Table 6.1 of Addendum 62n calls for 10 cfm (4.7 L/s) per person and 0.12 cfm/ft 2 (0.6 L/s per m 2 ), or a 250 cfm (118 L/s) person component and a 125 cfm (59 L/s) floor component for a 375 cfm (177 L/s) total ventilation rate (same as the old standard where the ventilation rate per person was 15 cfm [7 L/s]). If the OA CO 2 concentration was approximately 300 ppm, the resulting room concentration would be approximately 1,000 ppm. However, if the student population drops to 12, Addendum 62n would call for only a 120 cfm (57 L/s) person component and the same 125 cfm (59 L/s) floor component for a 245 cfm (116 L/s) ventilation rate (if the ventilation rate were divided by the occupancy the apparent per person ventilation rate has increased to 20.4 cfm [9.6 L/s] because of the constant floor component). The resulting room CO 2 concentration, assuming the OA concentration remained the same, would be approximately 800 ppm. An alternative reading of the condition for dynamic ventilation rate reset using indoor CO 2 concentrations, would be to hold the space CO 2 concentration less than 700 ppm above the outdoor air concentration (or approximately 1,000 ppm as in the past) and use the floor component only as a minimum OA requirement. This alternative reading is inspired by Standard , Section 6.1, Page 7, Comfort (odor) criteria with respect to human bioeffluents are likely to be satisfied if the ventilation results in indoor CO 2 concentrations less than 700 ppm above the outdoor air concentration. Controls to strictly maintain the minimum floor component may be tricky. If this alternative reading is what the drafters of Addendum 62n had in mind, it should have been explicitly stated. So read, it follows that, in practice, VAV system box minimums will be set at the floor component flow, or in the previous example, greater than 125 cfm (59 L/s). I believe this alternate reading of Addendum 62n is flawed since it essentially ignores a portion of the contribution from the floor component when the occupancy drops below design. IAQ Applications/Winter

62 Application Issues Figure 1: Logical blocks used to estimate occupancy. For example, eight people in the space designed for 25 would result in a space CO 2 concentration of 1,000 ppm with just the floor component flow of 125 cfm (59 L/s). Such operation would be 80 cfm (38 L/s) deficient and fail to provide the needed OA to adequately dilute both the occupant- and nonoccupant-related source of contaminants. CO 2 Testing Dynamic Occupancy Given this background, an experiment to test the ability of space CO 2 concentrations to accurately predict real-time occupancy was undertaken. If the real-time occupancy can be determined, the space CO 2 concentration setpoint can be computed dynamically, and conventional PID control loops can be used to modulate the OA flow to meet the new Addendum 62n requirements. The site of the experimental work is a 3,200 ft 2 (297 m 2 ) classroom/studio facility served by a dedicated outdoor air system. The design occupancy is 45 students and their studio critics. Addendum 62n, Table 6.1, calls for the ventilation rate to be 450 cfm (212 L/s) for the person component and 385 cfm (182 L/s) for the floor component for a total 835 cfm (394 L/s) of OA. The supply and relief air CO 2 concentrations and the supply airflow rate are monitored in real time. The commercial instrumentation used in the experiment is at or near the top of the line for quality and accuracy. The signals are connected to the system s DDC hardware, where the software computes the dynamic occupancy. Computing Occupancy From Measured Data Both steady state and transient equations can be used to solve for occupancy. 1 The transient equation in difference form is: ( V ( N N ) τ + SA ( N C )) ( 1,000,000) Pep = - 1 i G where Pep = number of occupants V = the space air volume, ft 3 N = the space CO 2 concentration at the present time step, ppm N 1 = the space CO 2 concentration one time step back, ppm τ = the time step, min. SA = the supply airflow rate, scfm C i = the CO 2 concentration in the supply air, ppm G = the CO 2 generation rate per person, scfm The transient equation is easily converted to a steady state V N N - 1, equal to zero. equation by setting the term, ( ) τ The graphic program blocks used to make this calculation are illustrated in Figure 1. The first block at the left is used to store the N 1 value as well as obtain the current value of N at time intervals τ. The second block solves the Pep equation dynamically. The third block limits the rate of change of Pep to 1 person/15 seconds to dampen transients caused by fluctuating CO 2 values and supply air fluctuations. Figure 2A illustrates CO 2 fluctuations that are 10 to 15 ppm. Figure 2B shows supply air fluctuations that are generally less than 25 cfm (11.8 L/s). The fourth and fifth blocks limit the occupancy numbers between 0 and 50. Finally, the sixth block limits the display of occupants to integer values. For the test case, the active space volume (room volume less the estimated volume of furniture, occupants, and other objects in the space) is 32,800 ft 3 (929 m 3 ). The assumed CO 2 generation rate per person, G, is 0.01 scfm (0.005 L/s). The time step, τ, varied considerably from 0.05 minutes to 1 minute in the experiments with little impact. To appreciate how well the third block dampens fluctuations, consider two cases. Case 1: When the τ was 0.5 minutes, the N and N 1 values were 644 and ppm respectively. The C i was 386 ppm and the SA flow rate about 850 scfm (401 L/s). Under these transient conditions, the computed occupancy was If the transient term of the equation were set to 0, the steady state occupancy would be 850 ( ) ( ,000,000) = people. At that moment, the displayed number of people was 21. Case 2: Occurring several minutes later, N and N 1 had values of 633 and ppm, respectively. The C i was ppm and the SA flow rate about 850 scfm (401 L/s). Under these transient conditions, the computed occupancy was If the transient term of the equation was set to 0, the steady state occupancy would be 850 ( ) ( ,000,000) = people. At that moment, the displayed number of people was 22. The presence of the transient term causes some oscillation in the estimated occupancy, but is able to respond correctly and rapidly to abrupt changes in occupancy. However, when the occupancy is steady or changing slowly, the transient equation oscillates by only one or two people. Implicit in the equation is the assumption that the space is held at a positive pressure, so that air with a CO 2 concentration higher or lower than the space air is not drawn inward, which would alter the concentration balances. When the space is held at a positive pressure, all of the room air leaves with an assumed uniform CO 2 concentration equal to that measured in the return airstream. In the case of the test facility, the use of high induction supply air diffusers produces well-mixed air. 22 Comments/Letters: iaq@ashrae.org IAQ Applications/Winter 2004

63 Application Issues SCFM People PPM CO 2 (EA) People FM 3 (SA) CO 2 (SA) 18:00 0:00 6:00 18:00 0:00 6:00 18:00 0:00 6:00 Figure 2: CO 2 data and estimated occupancy. Verification that the space pressures were maintained positive during testing is given in Figure 3. The upper curve is a plot of the space pressure, ranging from about in. w.g. (1.25 Pa) to 0.02 in. w.g. (5 Pa). The lower curve shows the supply and return airflow rates of approximately 850 and 760 scfm (401 and 359 L/s), respectively. CO 2 -Based Estimates, Observed Occupancy Walk-through occupancy counts have been made during several months and compared with the computed occupancy. The actual occupancy count agrees with estimated occupancy within two people. It also gives accurate counts when there is a rapid change in occupancy. Dynamically Resetting the Ventilation Rate By measuring the supply air CO 2 concentration and the supply airflow rate, real-time occupancy estimates are made. The estimated occupancy and the floor area can be used along with Table 6.1 of Addendum 62n to compute the space outdoor airflow at the breathing zone. The actual space OA airflow rate may be more or less than the summation using Table 6.1 due to the zone air distribution effectiveness factor (E Z ) presented in Table 6.2 of Addendum 62n. For cooling via ceiling diffusers, E Z is 1. By knowing the space outdoor airflow rate requirements, the space CO 2 setpoint easily can be computed using the following steps: 1. Compute the floor component ventilation air; 2. Compute the occupant component ventilation air; 3. Sum the components; 4. Divide the sum by the number of occupants to get an equivalent cfm/occupant, V o ; and 5. Compute the room CO 2 setpoint with the equation: room setpoint (ppm) = 10500/V o + C OA C OA is the OA CO 2 concentration, assumed to be 300 ppm in the example. Using the example classroom: The floor component is 125 cfm (59 L/s) A B SCFM in. w.g FM 5 (BA) FM 4 (RA) FM 3 (SA) 18:00 0:00 6:00 18:00 0:00 6:00 Figure 3: Measure of the space pressure. Room Setpoint Space Estimated Assuming Concentration Occupancy 300 ppm OA Above OA Level 25 Design Occupancy 1,000 ppm 700 ppm ppm 640 ppm ppm 570 ppm ppm 470 ppm ppm 300 ppm Table 1: Space CO 2 setpoints for estimated occupancies. The occupant component (10 people) is 100 cfm (47 L/s) The sum is 225 cfm (106 L/s) V o = 225/10 = 22.5 equivalent cfm (11 L/s) per person Room setpoint (ppm) =770 Table 1 gives room CO 2 setpoints for various other occupancy estimates. Note the setpoint is not constant, but decreases from 1,000 ppm for full occupancy to 600 ppm when only five people are present. The changing setpoint is brought about by the constant ventilation requirement of the floor component. Conclusion Dynamic OA intake reset with Addendum 62n should be accurately achieved where the design team chooses to measure occupancy based upon supply and return air CO 2 concentrations. The actual algorithms and stability issues associated with this approach obviously have not been addressed in this column. By choosing to use the occupancy estimates and a strict reading of Addendum 62n, contaminants from both the occupant- and non-occupant related sources can be accommodated. Dynamic reset is a valuable tool for VAV systems when it comes to saving heating and cooling energy. VAV systems fan energy reduction with Dynamic OA reset does not occur. Reference 1. Ke, Y., S. Mumma Using carbon dioxide measurements to determine occupancy for ventilation controls. ASHRAE Transactions 103(2): S.A. Mumma, Ph.D., P.E., is a professor of architectural engineering at Penn State University, University Park, Pa. IAQ Applications/Winter

64 Application Issues Copyright 2002, American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. This posting is by permission of ASHRAE IAQ Applications, Summer This article may not be copied nor distributed in either paper or digital form without ASHRAE s permission. Contact ASHRAE at Using Dedicated Outdoor Air Systems Demand Controlled Ventilation By Stanley A. Mumma, Ph.D., P.E. Fellow ASHRAE Carbon dioxide-based demand controlled ventilation (DCV) for VAV systems is intended to resolve the traditional conflict between operating cost and verifiably maintaining ventilation for acceptable indoor air quality as required in ANSI/ASHRAE Standard , Ventilation for Acceptable Indoor Air Quality. If properly applied, DCV can be expected to: Reduce energy operating costs during off-design occupancy, compared to all-air VAV systems operating with a fixed design minimum OA. (Savings are achieved by reducing overventilation of partially occupied zones.) Accommodate infiltration, exfiltration, local exhaust, and interzonal transfer. Maintain the desired ventilation Mumma rate per person compared to most allair VAV approaches designed to accommodate variable occupancy. It has been shown 1 that, for multiple space systems, a single CO 2 sensor in the common return causes critical spaces to be underventilated, in some cases by up to 90% (i.e., only 2 cfm [0.9 L/s] per person instead of the required 20 cfm [9.4 L/s] per person). Consequently, sensors must be located in enough zones to detect the critical spaces. Zones with short circuit paths between the supply and return air can create CO 2 measurement problems, which can lead to a failure to meet Standard 62. With properly operating DCV systems, the required minimum quantity of OA decreases roughly in proportion to the total building occupancy reduction. With a uniform percent occupancy reduction in every zone, the percent OA reduction is nearly the same as the total building occupancy reduction. And with non-uniform percent occupancy reduction in the building, the percent OA reduction is less than the total building occupancy reduction. In any event, the literature 2 warns that when DCV is used, Design Load, Tons DOAS 10,000 scfm ventilation rates should not be reduced below 20% to 50% of design. Maintaining this minimum rate supplies sufficient ventilation air to dilute building contaminant sources, even at low occupancy levels. Dedicated OA Systems The dedicated outdoor air systems (DOAS) discussed here supply the design minimum outdoor air directly into every zone of the building while working with a parallel comfort conditioning system. 3 The OA generally is preconditioned 4 to some extent, depending on: the parallel system used to meet zone loads; how the OA is introduced into the zone; and 67 VAV 12,000 scfm Figure 1: OA load on the chiller, Atlanta. VAV 17,000 scfm IAQ Applications/Summer

65 Application Issues 60, % OCC $4, % OCC 50,000 75% OCC 50% OCC 25% OCC 50,400 $4,000 $3,500 75% OCC 50% OCC 25% OCC $4,030 Ton-Hours/Year 40,000 30,000 20,000 35,000 26,625 17,750 37,800 25,200 Annual Operating Cost, $ $3,000 $2,500 $2,000 $1,500 $2,850 $2,138 $1,425 $3,023 $2,015 10, ,900 4,425 2,950 1,475 DOAS 10,000 scfm VAV 12,000 scfm VAV 17,000 scfm Figure 2: Annual ton-hours on the chiller for Atlanta with DCV. 8,875 12,600 $1,000 $500 $0 $470 $353 $235 $118 DOAS 10,000 scfm VAV 12,000 scfm VAV 17,000 scfm Figure 3: Operating costs to condition OA, based on 0.8 kw/ton and $0.10/kWh. $713 $1,008 the extent of energy recovery used. For buildings with design DOAS flow rates more than 7,000 cfm (3300 L/s), ANSI/ASHRAE/IESNA Standard , Energy Efficient Design of New Buildings Except Low-Rise Residential Buildings, requires total energy recovery. For much of the United States, the design wet-bulb temperatures are sufficiently high that Standard requires an enthalpy wheel. An enthalpy wheel, with an effectiveness of 80%, is capable of preconditioning the OA to within 20% of the zone(s) return air conditions (dry bulb and humidity ratio). This yields a large reduction in the outdoor air latent and sensible cooling, or heating and humidification loads. The result of total energy recovery is that the OA cooling and dehumidification loads are reduced by 80%. Heating of OA is nearly eliminated as is the need for winter humidification. Energy Consumption Atlanta weather data is presented for evaluating DCV s impact on annual energy consumption. The example assumes that the uncorrected OA at design occupancy is 10,000 scfm (4700 L/s) and the space conditions are maintained at 75 F (24 C) dry-bulb temperature and 50% RH. Typical Meteorological Year (TMY) data were used, assuming that the building would be occupied 6 days per week, from 7 a.m. to 7 p.m. When the multiple spaces equation of Standard 62 is applied, the uncorrected OA increases by 20% to 70%. Because of this, a fully occupied building served by an all-air VAV system would need to supply the corrected OA flow rate of between 12,000 to 17,000 scfm ( L/s), while a DOAS would only need to supply the uncorrected 10,000 scfm (4700 L/s)of OA. The design OA load on the chiller for the DOAS and for a VAV system with the lower and upper OA flow rates is presented in Figure 1. Because Standard requires the DOAS to use total energy recovery and needs to condition only the uncorrected OA flow, the peak OA load on the chiller is only 11 tons (39 kw). Since the DCV s intent is to reduce energy consumption, it is the author s experience that it is not used with total energy recovery equipment. Therefore, the design impacts of the VAV system s corrected OA flow rates of 12, Comments/Letters: iaq@ashrae.org IAQ Applications/Summer 2002