• Keep outside air damper closed during warm-up and cool-down periods

The fan energy savings increase significantly as the fan speed limit decreases. Figure 26.5 shows the theoretical fan power savings. When the fan speed limit is 50% of design fan speed, the potential fan energy savings are 75% of the fan energy even if the fan runs twice as long. The theoretical model did not consider the variable speed drive loss. The actual energy savings will normally be somewhat lower than the model projected value.

Note that if the outside air damper cannot be closed tightly, extra thermal energy may be required to cool or warm up outside air that leaks through the damper. This factor should be considered when this measure is used.

26.4.2 Operational Efficiency Measures for AHU Systems

Air handler systems normally condition and distribute air inside buildings. A typical AHU system consists of some combination of heating and cooling coils, supply and return air fans, filters, humidifiers, dampers, ductwork, terminal boxes, and associated safety and control devices, and may include an economizer. As the building load changes, AHUs change one or more of the following parameters to maintain building com fort: outside air intake, total airflow, static pressure, and supply air temperature and humidity. Both operating schedules and initial system set up, such as total airflow and outside airflow, significantly impact building energy consumption and comfort. The following ten major CC measures should be used to optimize AHU operation and control schedules:

Adjust total airflow for constant air volume systems

Set minimum outside air intake correctly Improve static pressure set-point and schedule Optimize supply air temperatures Improve economizer operation and control Improve coupled control AHU operation Valve off hot air flow for dual duct AHUs during summer

Install VFD on constant air volume systems Install airflow control for VAV systems Improve terminal box operation Adjust Total Air Flow and Fan Head for Constant Air Volume Systems

Air flow rates are significantly higher than required in most buildings primarily due to system over-sizing. In some large systems, an oversized fan causes over-pressurization in terminal boxes. This excessive pressurization is the primary cause of room noise. The excessive airflow often causes excessive fan energy consumption, excessive heating and cooling energy consumption, humidity control problems, and excessive noise in terminal boxes18. Set Minimum Outside Air Intake Correctly Outside air intake rates are often significantly higher than design values in existing buildings due to lack of accurate measurements, incorrect design calculations and balancing, and operation and maintenance problems. Excessive outside air intake is caused by the mixed air chamber pressure being lower than the design value, by significant outside air leakage through the maximum outside air damper on systems with an economizer, by the minimum outside air intake being set to use minimum total airflow for a VAV system, or by lower than expected/designed occupancy. Improve Static Pressure Setpoint and Schedule The supply air static pressure is often used to control fan speed and ensure adequate airflow to each zone. If the static pressure setpoint is lower than required, some zones may experience comfort problems due to lack of airflow. If the static pressure setpoint is too high, fan power will be excessive. In most existing terminal boxes, proportional controllers are used to maintain the airflow setpoint. When the static pressure is too high, the actual airflow is higher than its setpoint. The additional airflow depends on the setting of the control band. Field measurements19 have found that the excessive airflow can be as high as 20%. Excessive airflow can also occur when terminal box controllers are malfunctioning. For pressure dependent terminal boxes, high static pressure causes significant excessive airflow. Consequently, high static pressure often causes unnecessary heating and cooling energy consumption. A higher than necessary static pressure setpoint is also the primary reason for noise problems in buildings. Optimize Supply Air Temperatures

Supply air temperatures (cooling coil discharge air temperature for single duct systems; cold deck and hot deck temperatures for dual duct systems) are the most important operation and control parameters for AHUs. If the cold air supply temperature is too low, the AHU may remove excessive moisture during the summer using mechanical cooling. The terminal boxes must then warm the over-cooled air before sending it to each individual diffuser for a single duct AHU. More hot air is required in dual duct air handlers. The lower air temperature consumes more thermal energy in either system. If the cold air supply temperature is too high, the building may lose comfort control. The fan must supply more air to the building during the cooling season, so fan power will be higher than necessary. The goal of optimal supply air temperature schedules is to minimize combined fan power and thermal energy consumption or cost. Although developing optimal reset schedules requires a comprehensive engineering analysis, improved (near optimal) schedules can be developed based on several simple rules. Guidelines for developing improved supply air temperature reset schedules are available for four major types of AHU systems20. Improve Economizer Operation and Control

An economizer is designed to eliminate mechanical cooling when the outside air temperature is lower than the supply air temperature setpoint and decrease mechanical cooling when the outside air temperature is between the cold deck temperature and a high temperature limit, which is typically less than 70°F. Economizer control is often implemented so it controls mixed air temperature at the cold deck temperature or simply 55°F. This control algorithm is far from optimum. It may, in fact, actually increase the building energy consumption. The economizer operation can be improved using the following steps:

1. Integrate economizer control with optimal cold deck temperature reset. It is tempting to ignore cold deck reset when the economizer is operating, since the cooling is free. However, cold deck reset normally saves significant heating.

2. For a draw-through AHU, set the mixed air temperature 1°F lower than the cold deck temperature setpoint. For a blow-through unit, set the mixed air temperature at least 2°F lower than the supply air temperature setpoint. This will eliminate chilled water valve hunting and unnecessary mechanical cooling.

3. For a dual duct AHU, the economizer should be disabled if the hot air flow is higher than the cold air flow since the heating energy penalty is then typically higher than cooling energy savings.

4. Set the economizer operating range as wide as possible. For dry climates, the economizer should be activated when the outside air temperature is between 30°F and 75°F, between 30°F and 65°F for normal climates, and between 30°F and 60°F for humid climates. When proper return and outside air mixing can be achieved, the economizer can be activated even when the outside air temperature is below 30°F.

5. Measure the true mixed air temperature. Most mixing chambers do not achieve complete mixing of the return air and outside air before reaching the cooling coil. It is particularly important that mixed air temperature be measured accurately when an economizer is being used. An averaging temperature sensor should be used for the mixed air temperature measurement. Improve Coupled Control AHU Operation

Coupled control is often used in single-zone single-duct, constant volume systems. Conceptually, this system provides cooling or heating as needed to maintain the setpoint temperature in the zone and uses simultaneous heating and cooling only when the humidistat indicates that additional cooling (followed by reheat) is needed to provide humidity control. However, the humidistat is often disabled for a number of reasons. To control room relative humidity level, the control signals or spring ranges are overlapped. Simultaneous heating and cooling often occurs almost continuously. Valve Off Hot Air Flow for Dual Duct AHUs During Summer

During the summer, most commercial buildings do not need heating. Theoretically, hot air should be zero for dual duct VAV systems. However, hot air leakage through terminal boxes is often significant due to excessive static pressure on the hot air damper. For constant air volume systems, hot air flow is often up to 30% of the total airflow. During summer months, hot air temperatures as high as 140°F have been observed due to hot water leakage through valves. The excessively high hot air temperature often causes hot complaints in some locations. Eliminating this hot air flow can improve building thermal comfort, reduce fan power, cooling consumption, and heating consumption. Install VFD on Constant Air Volume Systems The building heating load and cooling load varies significantly with both weather and internal occupancy conditions. In constant air volume systems, a significant amount of energy is consumed unnecessarily due to humidity control requirements. Most of this energy waste can be avoided by simply installing a VFD on the fan without a major retrofit effort. Guidelines for VFD installation are available for dual duct, multi-zone, and single duct systems21. Airflow Control for VAV Systems

Airflow control of VAV systems has been an important design and research subject since the VAV system was introduced. An airflow control method should: (1) ensure sufficient airflow to each space or zone; (2) control outside air intake properly; and (3) maintain a positive building pressure. These goals can be achieved using the variable speed drive volume tracking (VS-DVT) method22,23. Improve Terminal Box Operation

The terminal box is the end device of the AHU system. It directly controls room temperature and airflow. Improving the set up and operation are critical for room comfort and energy efficiency. The following measures are suggested:

• Set minimum air damper position properly for pressure dependent terminal boxes.

• Use VAV control algorithm for constant air volume terminal boxes.

• Integrate lighting and terminal box control.

• Integrate airflow and temperature reset.

• Improve terminal box control performance.

in a Previously Retrofit Building24 Case Study Building Description

The building studied in this paper is the 295,000 gross ft2 (226,000 ft2 net) Zachry Engineering Center (ZEC), located on the Texas A&M University campus (30°N, 96°W) where the average January temperature is 50°F and average July temperature is 84°F, and pictured in Figure 26.6. The building has four-floors plus an unconditioned basement parking level. It was constructed in the early 1970s and is a heavy structure with 6 in. concrete floors and insulated exterior walls made of precast concrete and porcelain-plated steel panels. About 12% of the exterior wall area is covered with single-pane, bronze-tinted glazing. The windows are recessed approximately 24 in. from the exterior walls, which provides some shading. Approximately 2835 ft2 of northeast-facing clerestory windows admit daylight into the core of the building.

The ZEC includes offices, classrooms, laboratories and computer rooms and is open 24 hours per day, 365 days per year with heaviest occupancy during normal working hours between 8 a.m. and 6 p.m. on weekdays. Occupancy, electrical consumption and chilled water consumption show marked weekday/weekend differences with peak weekend electrical consumption less than 10% above the nightly minimum; weekday holiday occupancy is similar to weekend usage with intermediate usage on weekdays between semesters when class

Figure 26.6 The Zachry Engineering Center on the Texas A&M campus25.

rooms are not in use, but laboratories and offices are occupied.

HVAC Systems. Twelve identical dual-duct systems with 40 hp fans rated at 35,000 cfm and eight smaller air handlers (3 hp average) supply air to the zones in the building. Supply and return air ducts are located around the perimeter of the building. These were operated with a constant outdoor air intake at a nominal value of 10% of design flow. The large dual-duct constant air volume (DDCAV) systems were converted to dual-duct variable-air volume (DDVAV) systems accompanied by connection to the campus energy management and control system (EMCS) in 1991. This retrofit successfully reduced fan power consumption by 44%, cooling consumption by 23%, and heating consumption by 84%.

Monitoring of energy use. In the engineering center about 50 channels of hourly data have been recorded and collected each week since May 1989. The sensors are scanned every 4 seconds and the values are integrated to give hourly totals or averages as appropriate. The important channels for savings measurement are those for air handler electricity consumption and whole-building heating and cooling energy use. Air handler electricity consumption is measured at the building's motor control center (MCC) and represents all of the air-handling units and most of the heating, ventilating, and air-conditioning (HVAC)-related pumps in the building. Cooling and heating energy use are determined by a Btu meter which integrates the monitored fluid flow rate and temperature difference across the supply and return lines of the chilled- and hot-water supply to the building. The majority of the 50 channels of monitored information come from one air handler that was highly instrumented.

4.3.2 Continuous Commissioning of the Zachry Engineering Center

The Continuous Commissioning process was applied to the Zachry Engineering Center in 1996-97. In this case, the initial survey and specification of monitoring portions of the CC process were not performed since the university president had decided to implement CC on campus based on its success in numerous other locations rather than on the results of individual building surveys. Metering had been installed much earlier in the Zachry Engineering Center as part of the retrofit process, so performance baselines were already available.

Conduct system measurements and develop proposed CC measures

The facility survey found that the building control system set-up was far from optimum and found numerous other problems in the building as well. The basic

Figure 26.7 ZEC daily motor control center (MCC) electric consumption in 1990 before the retrofit, in 1994 after the retrofit, and in 1997 after CC26.

control strategies found in the building are summarized in Table 26.3 in the column labeled "Pre-CC Control Practice." The ranges shown for constant parameters reflect different constant values for different individual air handlers.

The control practices shown in the table are all widely used, but none are optimal for this building. The campus controls engineer worked closely with the CC engineers during the survey. The items shown in Table 26.3 could all be determined by examination of the control system in the building, but the facility survey also examines a great deal of the equipment throughout the building and found numerous cases of valves that let too much hot or cold water flow, control settings that caused continuous motion and unnecessary wear on valves, air ducts that had blown off of the terminal boxes, kinks in air ducts that led to rooms that could not be properly heated or cooled, etc. Following the survey, the building performance was analyzed and CC measures including optimum control schedules were developed for the building in cooperation with the campus controls engineer.

Implement CC measures

Following the survey, the building performance was analyzed and optimum control schedules were developed for the building in cooperation with the campus control engineer. The air handlers, pumps and terminal boxes had major control parameters changed to values shown in the "Post-CC" column of Table 26.3. Most of the control parameters were optimized to vary as a function of outside air temperature, Toa, as indicated.

In addition to optimizing the control settings for the heating and cooling systems in the building, numerous problems specific to individual rooms, ducts, or terminal boxes were diagnosed and resolved. These included items like damper motors that were disconnected, bent air ducts that could not supply enough air to properly control room temperature, leaking air dampers, dampers that indicated open when only partly open, etc.

Problems of this sort often had led to occupant complaints that were partially resolved without fixing the real problem. For example, if a duct was constricted so inadequate flow reached a room, the pressure in the air handler might be increased to get additional flow into the room. "Fixes" like this typically improve room comfort, but sometimes lead to additional heating and cooling consumption in every other room on the same air handler.

Document energy savings

Implementation of these measures resulted in significant additional savings beyond the original savings from the VAV retrofit and controls upgrade as shown in Figures 26.7, 26.8 and 26.9. Figure 26.7 shows the motor control center power consumption as a func

Table 26.3. Major control settings in the Zachry Engineering Center before and after implementation of CC27.


Pre-CC Control Practice

Post-CC Control Practice

Pressure in air ducts

Constant at 2.5 - 3.5 in H2O

1- 2 inH2O as Toa increases

Cold air temperature

Constant at 50°F - 55°F

60°F - 55°F as Toa increases

Hot air temperature

Constant at 110°F - 120°F

90°F - 70°F as Toa increases

Air flow to rooms

Variable - but inefficient

Optimized min/max flow and damper operation

Heating pump control

Operated continuously

On when Toa>55°F

Cooling pump control

Variable speed with shut-off

Pressure depends on flow

tion of ambient temperature for 1990, 1994 and 1997. It is evident that the minimum fan power has been cut in half and there has been some reduction even at summer design conditions. Figure 26.9 shows the hot water consumption for 1990, 1994 and 1997, again as a function of daily average temperature. The retrofit reduced the annual hot water (HW) consumption for heating to only 16% of the baseline, so there is little room for further reduction. However, it can be seen that the CC measures further reduced HW consumption, particularly at low temperatures. The largest savings from the CC measures are seen in the chilled water consumption as shown in Figure 26.8. The largest fractional savings occur at low ambient temperatures, but the largest absolute savings occur at the highest ambient temperatures.

The annualized consumption values for the baseline, post-retrofit and post-CC conditions are shown in Table 26.4. The MCC consumption for 1997 was 1,209,918 kWh, 74% of the 1994 consumption and only 41% of the 1990 consumption. On an annual basis, the post-CC HW consumption normalized to 1994 weather was 1940 MMBtu, a reduction to only 10% of baseline consumption and a reduction of 34% from the 1994 consumption. The CC measures reduced the post-CC chilled water (CHW) consumption to 17,400 MMBtu, a reduction of 17,800 MMBtu which is noticeably larger than the 13,900 MMBtu savings produced by the retrofit. The CHW savings accounted for the largest portion of the CC savings in this cooling-dominated climate.

Generalized application of case study and conclusions

The major energy savings from the CC activities in the case study building resulted from five items.

1. Optimization of duct static pressures at lower levels. This reduces fan power and also reduces damper leakage that increases both heating and cooling consumption.

2. Optimization of cold air temperatures. Most buildings in our experience use a constant set-point for

Table 26.4. Consumption at the Zachry Engineering Center before and after retrofit and after implementation of CC measures28.

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