Dual Maximum VAV Box Control Logic

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1 This article was published in ASHRAE Journal, December 212. Copyright 212 ASHRAE. Posted at This article may not be copied and/or distributed electronically or in paper form without permission of ASHRAE. For more information about ASHRAE Journal, visit Dual Maximum VAV Box Control Logic By Steven T. Taylor, P.E., Fellow ASHRAE; Jeff Stein, P.E., Member ASHRAE; Gwelen Paliaga, Member ASHRAE; and Hwakong Cheng, P.E., Member ASHRAE Most variable air volume (VAV) systems use the same VAV box control logic that was used when the system first became popular in the 197s. But with modern direct digital controls (DDC), zone control logic can be much more sophisticated and higher performing. This article discusses one such strategy dubbed dual maximum VAV box control logic and shows how this control logic improves energy efficiency and occupant comfort. Conventional VAV Box Logic VAV boxes with reheat coils were traditionally controlled using the control logic shown in Figure 1. The supply airflow setpoint is reset from the zone maximum airflow setpoint when the space is at full cooling proportionally down the zone minimum when no cooling is required. This minimum airflow rate is maintained as the space temperature falls through the deadband into heating mode. The hot water valve then opens (or electric heat stages or modulates up) to maintain the space at the heating setpoint until it is fully open (or all electric heat stages are on). The minimum airflow setpoint with this logic is determined by the following: 1. No less than the larger of: a. The zone minimum outdoor air rate, e.g., that required by code or Standard b. The amount of air required to heat the space at a supply air temperature appropriate for the application. For overhead supply and return systems, ASHRAE Standard (Section ) does not allow the supply air temperature to exceed 2 F (11.1 C) above the space temperature to avoid excessive stratification and short circuiting. To maintain a zone air distribution effectiveness of 1. for ceiling supply/return systems, ASHRAE Standard 62.1 requires that the supply air temperature be no more than 15 F (8.3 C) above space temperature. So for typical overhead systems, the supply air temperature is limited to 85 F to 9 F (29 C to 32 C), assuming a 7 F (21 C) space heating setpoint temperature. c. The lowest setpoint allowed by the VAV box controls. With pneumatic controls, the minimum setpoint About the Authors Steven T. Taylor, P.E., and Jeff Stein, P.E., are principals, and Gwelen Paliaga and Hwakong Cheng, P.E., are senior engineers at Taylor Engineering in Alameda, Calif. 16 ASHRAE Journal ashrae.org December 212

2 is quite high. But with modern direct digital controls, the minimum can be very low, as discussed in more detail later. 2. No more than 3% of the cooling maximum airflow setpoint per the prescriptive requirements of ASHRAE Standard (Section ) and California s Title Energy Standards 3 for pneumatic controls. This fairly low setpoint significantly limits the use of VAV reheat terminals in zones with high heating loads, such as exterior zones with large windows in cold climates. For these zones, other systems that do not rely on the cooling airflow for heating must be used, such as fan-powered VAV boxes, radiant heat, or convectors. This logic results in a large amount of reheat energy, shown in Figure 1 by the purple area. For a given supply air temperature, the lower the minimum airflow setpoint, the lower the reheat energy, which is why the energy standards referenced above limit this value. Dual Maximum VAV Box Logic The previous conventional logic still applies to zones that do not have DDC, but the 28 version of Title 24 and an addendum proposed to ASHRAE Standard require that VAV boxes with DDC provide what has been dubbed dual maximum logic. Here is the language from Title (the Standard 9.1 addendum has similar language): A. For each zone with direct digital controls (DDC): i. The volume of primary air that is reheated, re-cooled, or mixed air supply shall not exceed the larger of: a. Fifty percent of the peak primary airflow; or b. The design zone outdoor airflow rate per Section 12.1 [which is the minimum ventilation section]. ii. The volume of primary air in the deadband shall not exceed the larger of: a. Twenty percent of the peak primary airflow; or b. The design zone outdoor airflow rate per Section iii. The first stage of heating consists of modulating the zone supply air temperature setpoint up to a maximum setpoint no higher than 95 F (35 C) while the airflow is maintained at the dead band flow rate. iv. The second stage of heating consists of modulating the airflow rate from the dead band flow rate up to the heating maximum flow rate. The logic this language requires is shown in Figure 2. The term dual maximum logic comes from the fact that there are now two maximum airflow setpoints: one for heating in addition to the one for cooling. With dual maximum logic, the minimum airflow setpoint is determined as follows: 1. No less than the larger of: a. The zone minimum outdoor air rate. Meeting ventilation requirements with low minimums is discussed further below. b. The lowest setpoint allowed by the VAV box controls. With modern DDC, the controllable minimum for most DDC manufacturers is usually not an issue because it is normally below the ventilation requirement. Minimum Figure 1: Conventional VAV reheat control diagram. Maximum Supply Air Temperature Maximum Heating Minimum Reheat Valve Position Heating Loop Heating Loop Dead Band Supply Air Temperature (Requires Discharge Temperature Sensor) Dead Band Maximum Cooling Loop Maximum Cooling Cooling Loop Figure 2: Dual maximum VAV reheat control diagram. Minimum controllable setpoints are generally much lower than those published by VAV box manufacturers, which are based on very conservative assumptions regarding the capability of the digital controller that is seldom provided with the VAV box. For a detailed discussion of how to determine the lowest controllable setpoint for a VAV box and controller, see the Journal article Sizing VAV Boxes. 5 Table 1 shows typical VAV box performance with a velocity pressure probe with a 2.3 amplification factor and a digital controller capable of controlling airflow to.4 in. w.g. (1 Pa) velocity pressure reading. The minimum controllable setpoint is about 8% of the design maximum airflow rate if the box is selected at.5 in. w.g. (125 Pa) pressure drop with a two-row hot water coil, up to about 12% if the design airflow rate is at a value just above the maximum of the next smallest box size. A reasonable rule-of-thumb is to assume the minimum can be no lower than 1% of the cooling maximum. Another approach is for designers to ignore this constraint in their VAV box schedules and rely on the controls vendor to determine if any scheduled minimum setpoint is not possible and adjust the setpoint accordingly. December 212 ASHRAE Journal 17

3 2. No more than 2% of the cooling maximum airflow setpoint. The heating maximum airflow setpoint must be: 1. No less than the larger of: a. The minimum airflow rate determined above. b. The amount of air required to heat the space at a supply air temperature appropriate for the application or as limited by Title 24/9.1, Box Inlet Diameter Maximum cfm at.5 in. w.g. Pressure Drop Minimum cfm at.4 in. w.g. Sensor Reading Minimum Ratio at Highest Maximum (%) Minimum Ratio at Lowest Maximum (%) % % 13.6% 1 1, % 12.7% 12 1, % 11.8% 14 2, % 11.3% 16 2, % 1.9% Table 1: Typical VAV box and controller performance. Maximum Heating Minimum Maximum Heating Minimum Reheat Valve Position Heating Loop Reheat Valve Position Heating Loop Cooling Loop Figure 3: Control diagrams for example configurable controllers. typically 85 F to 95 F (29 C to 35 C) for overhead supply/return systems as discussed earlier. 2. No more than 5% of the cooling maximum airflow setpoint. This is a larger percentage than the 3% allowed using conventional logic which allows VAV reheat terminals to be used in more applications. This does result in more reheat at high heating loads, but the lower minimum airflow setpoint (a maximum of 2% vs. 3% with the conventional control logic) offsets this by reducing reheat at lower heating loads. Since most zones spend far more time at low heating loads than they do at high heating loads, the dual maximums approach will result in overall reheat savings. To optimize the design, engineers should resist the shortcut to simply specify 2% minimum and 5% heating maximum setpoints. These are the highest setpoints allowed by the energy standards and higher than necessary for most applications. Rather, designers should use these values as limits only as they are intended and determine the actual setpoints based on the ventilation requirement (minimum airflow setpoint) and heating requirements (heating maximum airflow setpoint). To implement the dual maximum logic shown in Figure 2 will require the following controls: A fully programmable DDC zone controller, or a configurable controller with the dual maximum logic already installed. Configurable controllers have fixed control logic; only setpoints can be adjusted. Dual maximum logic is not new (the authors have specified it since the late 199s) but it is not yet common enough that many configurable controllers include it. Rather, they use conventional logic or one of the two logic diagrams shown in Figure 3. Neither of these meets the requirements of the Title 24/9.1 language shown earlier based on simulations that showed they used more energy with a heating maximum setpoint of 5% than conventional logic did at a 3% minimum. Prior to specifying or installing a VAV controller, the designer should verify that the controller is capable of providing the control logic shown in Figure 2 (Page 17). A supply air temperature sensor. This is needed to ensure that the supply air temperature does not get too hot and result in stratification, as would likely occur when the zone is heating at the low minimum airflow rate. As shown in Figure 2, the zone heating control loop does not modulate the hot water valve directly. Instead, cascading logic is used: the loop determines the supply air temperature setpoint and then the hot water valve is modulated to maintain that setpoint. The supply air temperature sensor (typically costing less than $1) required for this logic can also be used for diagnosing failed hot water control valves and other system faults. For electric resistance heat, a modulating controller (e.g., silicon controlled rectifier) and an electronic airflow sensor capable of sensing very low airflow rates and limiting heater capacity according to the available airflow. Electric heat with step controls generally cannot meet the requirements of the Title 24/9.1 language shown above because of high minimum airflow rates required to close safety switches. Figure 4 shows trend data from a VAV zone programmed with dual maximum logic. As the zone temperature drops below the heating setpoint (third row), the heating PID loop out- 18 ASHRAE Journal ashrae.org December 212 Dead Band Dead Band Maximum Cooling Maximum Cooling Cooling Loop

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5 Figure 4: Trend data from a VAV zone programmed with dual maximum logic. cfm Temperature ( F) Percent Temperature ( F) 9 8 DAT DAT :3 8:45 9: 9:15 9:3 9:45 1: 1:15 1:3 1: Heating % Cooling % Reheat Valve 2 8:3 8:45 9: 9:15 9:3 9:45 1: 1:15 1:3 1: Cooling Heating Zone Temperature 8:3 8:45 9: 9:15 9:3 9:45 1: 1:15 1:3 1:45 2, Cold Duct cfm 1, 8:3 8:45 9: 9:15 9:3 9:45 1: 1:15 1:3 1:45 put (Htg%) starts to increase (second row). As it does, the supply air temperature setpoint is increased from 55 F (13 C) up to a maximum of 85 F (29 C) (first row). The hot water valve (second row) is modulated to maintain supply air temperature at setpoint (first row). The valve loop is not very stable because of the low airflow rate and the changing setpoint, but this is not perceptible to occupants in the space. (Note that conventional logic without the benefit of supply air temperature feedback would have opened the valve equal to the zone PID loop output value and heated the air well above the 85 F to 9 F (29 C to 32 C) needed to limit stratification.) At 5% heating PID loop output (Htg%), the primary airflow setpoint ( Cold Duct CFM ) starts to increase from the 4 cfm (189 L/s) minimum (2% in this case) up to the heating maximum of 1, cfm (472 L/s) (5%) before the zone temperature starts to rise above the heating setpoint and the loop output starts to fall. In addition to ensuring that the supply air temperature is never too warm, another advantage of controlling the hot water valve off of supply air temperature is that the hot water system is selfbalancing even during transients such as warm-up, presuming a two-way valve variable flow system is used in accordance with Section of Standard 9.1. The control valves will never supply more than the design hot water flow due to the supply air temperature feedback without the need for balancing valves or balancing labor. With standard valve control off of space temperature, all valves go full wide open during warm-up, possibly resulting in flow imbalance unless the system is manually balanced. Dual Maximum Performance: RP-1515 Figure 5 shows data collected for a large office complex in Sunnyvale, Calif., for ASHRAE Research Project where VAV boxes initially used conventional logic with 3% minimums for several months after which the logic was changed to dual maximum logic with minimums set to the California Title 24 minimum ventilation rate. Even in warm weather, most of the zones operate at very low loads and, therefore, low airflow rates. The fact that zones are at minimum airflow even in warm weather also explains a surprising result from RP-1515: comfort improved when dual maximum logic was installed, from 77% acceptance to 88% acceptance in one building. Surveys (Figure 6) show that the lower acceptance with conventional logic is primary due to cold complaints caused by the high minimums pushing the zone down to the heating setpoint (e.g., 7 F [21 C]) even in warm weather when occupants are likely wearing summer apparel. Per ASHRAE Standard 55, 7 occupants are likely to find temperatures below about 74 F (23 C) too cool when wearing lightweight summer clothing. Energy savings from dual maximum logic are from reduced fan power, a small savings in mechanical cooling Fraction of Time Dual Maximum Logic Conventional Logic Percent of Design Figure 5: RP-1515 data for dual maximum logic vs. conventional logic (warm season data). 2 ASHRAE Journal ashrae.org December 212

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7 Percent of Total Votes ( 3) Cold 4% 18% ( 2) Cool ( 1) Slightly Cool () Neutral Conventional Logic Dual Maximum Logic (1) Slightly Warm (2) Warm (3) Hot Power Consumption (kw) Dual Maximum Logic Conventional Logic Thermal Sensation Figure 6: RP-1515 comfort survey, How do you rate your thermal sensation right now? Outside Air Temperature ( F) energy, and reduced natural gas use due to reduced reheat energy. Figure 7 shows measured AC unit power (fan and DX cooling) for an RP-1515 building while Figure 8 shows weather normalized annual savings for five buildings developed using correlations from the RP-1515 measurements and typical annual weather. Potential Issues with Low Minimum Rates Concerns regarding the low minimum airflow rates inherent with dual maximum logic include: 1. Low Air Diffusion Performance Index (ADPI) and diffuser dumping causing draft complaints. Bauman 8 using an instrumented test chamber with segmented thermal mannequins found that acceptable ADPI could be maintained at 25% of design flow. RP-1515 used a laboratory mockup similar to Bauman s to test performance at low rates. Tests of perforated diffusers with blades in the neck show negligible difference in ADPI between 8% and 18% of design flow (all close or equal to 1% ADPI), more uniform temperature at lower flow, and lower air speeds in the occupied region at lower flow (Figure 9). RP-1515 field studies showed improved thermal comfort and no increase in draft complaints. These tests indicate that declining diffuser performance at low airflow rates is not an issue with respect to occupant comfort. Tests were not run below 18% of design flow due to test lab limitations but the researchers fully expect the conclusions also apply to flow rates as low as 1% of design flow. 2. Low ventilation effectiveness. Most studies have shown that supply of cold air from the ceiling results in fully mixed spaces (air change effectiveness ~ 1.) regardless of diffuser selection and airflow rates. Fisk 9 found that to be the case even at 25% flow in cooling mode and found that effectiveness actually increased as percentage airflow rate was Annual Fan & Cooling Electricity Use kwh/ft 2 Annual Heating Gas Use Therms/ft Figure 7: RP-1515 fan and cooling power vs. OA temperature. reduced. RP-1515 will measure air change effectiveness at low airflow fractions in laboratory mockup with a single occupant and heating load produced by a controllable cold wall. The tests are incomplete at this time but we do not expect low airflow rates to significantly impact zone ventilation effectiveness in heating mode. 3. Meeting ventilation requirements. Complying with ventilation code in California is easy; the code simply requires that overall building outdoor air rates be maintained and that each space be supplied with the minimum ventilation rate without regard to the fraction of outdoor air being supplied outdoor air and return air are equally acceptable. This contrasts with ASHRAE Standard 62.1 which requires a more sophisticated approach using the socalled multiple spaces equations that track dilution from multiple airflow paths for each space. Research projects such as RP have shown that the multiple spaces approach in Standard 62.1 is Electricity Savings (Fan & Cooling) Building 1 Building 2 Building 3 Building 4 Building 5 Gas Savings (Reheat) Building 1 Building 2 Building 3 Building 4 Building 5 Figure 8: RP-1515 annual energy savings. Conventional Logic Dual Maximum Logic 22 ASHRAE Journal ashrae.org December 212

8 valid. California s simplified approach probably works from an indoor air quality perspective because California rates are typically larger than Standard 62.1 rates, about 3% larger for typical offices, for example. Unfortunately, it is difficult to use the Standard 62.1 multiple spaces equations because there are myriad assumptions required and the results for both overall outdoor air rate and zone minimum airflow rates can vary significantly depending on these assumptions. Some designers are extremely conservative and assume, for instance, that all zones are at their design airflow except for the critical zone which is fully occupied yet somehow at its minimum airflow rate. A more realistic approach is to use simulations to more accurately track airflow rates delivered by the zone control logic relative to the requirements for ventilation. Figure 1 shows the results of a DOE-2.2 simulation of a typical office building in Oakland, Calif., served by a VAV air handler with an outdoor air economizer. The y-axis is the ratio of the air handler outdoor air rate to that required by Standard 62.1 (V ot ) using the multiple spaces equation. Figure 11 shows the same results without an air economizer. The building model was extremely detailed. For instance, each room was provided with multiple realistic occupancy and internal load schedules that varied from one zone to another, day to day, and week to week. Occupancy sensors were modeled so that lights were shut off in unoccupied spaces. Minimum airflow rates were modeled as the larger of the California code ventilation requirement for offices or the controllable VAV box minimums. Conference rooms were modeled with CO 2 sensors that would raise this minimum rate as required to provide 15 cfm/ person (7 L/s per person), the California code occupant component ventilation requirement for offices. (CO 2 demand controlled ventilation [DCV] is required for these densely occupied spaces by Title 24 and Standard 9.1.) The results of the hourly analysis were exported to a spreadsheet where every hour was tested for compliance with Standard 62.1 using the multiple spaces equation, including assuming a.8 zone air distribution effectiveness when the zones were in heating mode. The results indicate that when the AHU has an outdoor air economizer, outdoor air rates met Standard 62.1 requirements for every hour and the average annual outdoor air supplied was 364% of the Standard 62.1 rate. However, without an economizer there were 151 hours (4% of the total occupied hours) that did not comply with Standard 62.1 but the average annual outdoor air rate was 168% larger than Standard 62.1 rates. To provide full Standard 62.1 compliance without an economizer, the AHU outdoor air rate would have to increase by about 5% above Title 24 minimum rates in this example. The results suggest the following approach to ensuring ventilation rates are met: Minimum zone airflow rates should be no lower than the Title 24 building rate of.15 cfm/ft 2 (.76 L/s m 2 ). CO 2 DCV should be used on all densely occupied spaces to allow the occupant ventilation rate component to be dynamically determined. Outdoor air economizers should ensure Standard 62.1 compliance as well as improve energy performance and thus are strongly encouraged. Air Speed (fpm) % Flow 33% Flow 18% Flow Distance from Diffuser (ft) Figure 9: RP-1515 laboratory measured air speed at 42 in. (1 m) above floor. Note: Diffuser located in center of 2 ft (6 m) wide room. Ratio of AHU OA Rate to Std Required OA Rate (Std 62.1 V ot = 1.) Outside Air Temperature ( F) Figure 1: Simulated outdoor air rate with economizer vs. Standard 62.1 using multiple spaces equations. Ratio of AHU OA Rate to Std Required OA Rate (Std 62.1 V ot = 1.) Outside Air Temperature ( F) Figure 11: Simulated outdoor air rate without economizer vs. Standard 62.1 using multiple spaces equations. Another control option is to dynamically reset zone minimum airflow setpoints based on air-handling system outdoor air supply, e.g., reducing minimums when the AHU is supplying 1% outdoor air in economizer mode and increasing minimums as outdoor air percentage falls when not in econo- December 212 ASHRAE Journal 23

9 mizer mode. This will be the subject of a future article. Conclusions The benefits of dual maximum logic over conventional logic include: Lower fan energy, lower reheat energy, and lower cooling energy use; Improved thermal comfort by not pushing zone temperature to heating setpoints during the cooling season; Reduced stratification due to supply air temperature control, and Self-balancing of hot water systems with two-way valves. Disadvantages include: Added cost of a discharge temperature sensor (very minor and also useful for diagnostics); Added cost for more complex control programming (a one-time cost for the control vendor); and Difficulty in determining zone minimum airflow setpoints and air handler minimum outdoor airflow setpoints that meet Standard 62.1 (not an issue under Title 24 ventilation code). RP-1515 results show that low minimum airflow setpoints rates improve air distribution performance and occupant comfort, results supported by the authors more than 1 years of successful experience using dual maximum logic in California. To ensure Standard 62.1 compliance at low minimum rates, we recommend using system simulations (rather than conservative assumptions in spreadsheets) to determine zone minimum setpoints along with CO 2 DCV and outdoor air economizers. References 1. ANSI/ASHRAE Standard , Ventilation for Acceptable Indoor Air Quality. 2. ANSI/ASHRAE Standard , Energy Standard for Buildings Except Low-Rise Residential Buildings. 3. Building Energy Efficiency Standards for Residential and Nonresidential Buildings. 28. Title 24, California Code of Regulations, Part 6. California Energy Commission. 4. Standard , Addendum ck. 5. Taylor, S., J. Stein. 24. Sizing VAV boxes. ASHRAE Journal 46(3). 6. Center for the Built Environment, Taylor Engineering, Price Industries ASHRAE RP 1515, Thermal and air quality acceptability in buildings that reduce energy by reducing minimum airflow from overhead diffusers. In progress; report expected January ASHRAE Standard 55-21, Thermal Environmental Conditions for Human Occupancy. 8. Bauman, F., C. Huizenga, T. Xu, T. Akimoto Thermal Comfort With a Variable Air Volume (VAV) System. Center for Environmental Design Research, University of California, Berkeley. 9. Fisk, W.J., D. Faulkner, D. Sullivan, F.S. Bauman Air change effectiveness and pollutant removal efficiency during adverse conditions. Indoor Air 7: Yuill, D., G. Yuill, A. Coward. 27. ASHRAE RP 1276: A Study of Multiple Space Effects on Ventilation System Efficiency in Standard and Experimental Validation of the Multiple Spaces Equation.

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