Energy Efficient BuildingsVariable-Air-Volume and Variable-Flow-Pumping Systems

IntroductionIn the past, most heating and air conditioning systems in multi-zone buildings were designed to deliver a constant volume of low-temperature supply air to each zone. The heating and cooling requirements of the individual zones were met by blending cool and warm supply air or by reheating the cool supply air. These systems are called constant-air-volume, CAV, systems. Today, most HVAC systems are designed to vary the quantity of air to each zone. These systems are called variable-air-volume, VAV, systems. VAV systems reduce the quantities of heating, cooling and fan energy required to condition a building.

Similarly, in the past, most pumping systems in multi-zone buildings were designed to deliver a constant volume of water to the end-uses independent of the actual building load. Today, variable speed drives coupled to intelligent control vary the flow of water me match building loads, and hence reduce pumping energy use.

This chapter shows how both constant-flow and variable-flow systems work and how to calculate energy savings from constant to variable flow conversions.

Single-Duct CAV and VAV SystemsA single-duct commercial heating and air conditioning system is shown below. For simplicity, the figure shows only two zones even though large commercial buildings have many zones. Despite this simplification, building energy use can often be accurately modeled using simple two zone models with an interior and exterior zone.

Single-duct system with two zones.

In single-duct constant-air-volume system, SD-CAV, the flow of cool air to the zone remains constant, but heat is added in a reheat box at each zone to meet the cooling/heating load in the zone.

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In a single-duct variable-air-volume system, SD-VAV, zone temperature is maintained by varying the flow of cool air down to the minimum air flow required for ventilation purposes by a VAV box. For energy efficiency, ASHRAE recommends minimum flows of 30% of peak flow or less. If the zone requires heating after the flow of cool air is at the minimum, then heat is added in the VAV/reheat box. VAV boxes for an internal zone that never requires heating and an external zone that sometime requires heating are shown in the pictures below.

VAV box without reheat VAV box with hot-water reheat

The volume flow rate of air from the cooling coil, Vcc, and quantity of heat added in the reheat box, Qh, for as functions of zone temperature are shown below.

SD-CAV SD-VAV

SD-CAV Heating and Cooling Energy UseIn a SD-CAV system, the flow of supply air through the cooling coil, Vsa, is constant, and the enthalpy, hsa, and temperature, Tsa, of supply air leaving the cooling coil are known. Thus, the energy extracted by the cooling coil, Qcc, can be calculated from an energy balance on the cooling coil, based on the enthalpy, h, and density, p, of the mixed air, ma, and supply air, sa and water, w.

Qcc = Vsa psa (hma – hsa) - Vsa psa (wma – wsa) hw

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The heating energy added by the heating coil in reheat box 1, Qhc1, can be calculated from an energy balance on the zone and reheat box, based on product of the air density and specific heat pcpa.

Qhc1 = Vsa1 pcpa (Tra – Tsa) – Qsen1

The total building heating energy requirement is the sum of the heating energy required by each reheat box. SD-VAV Heating and Cooling Energy UseIn a SD-VAV system, the flow of supply air through the cooling coil, Vsa, varies and is the sum of the air flow required by each zone. The procedure is:

Vsa1 = Qsen1 / (pcpa (Tra – Tsa))If Vsa1 < Vsa1,min then ‘heating required

Vsa1 = Vsa1,minQhc1 = Vsa1 pcpa (Tra – Tsa) – Qsen1

Else ‘no heating requiredVsa1 = Qsen1 / (pcpa (Tra – Tsa))Qhc1 = 0

End If

The total building heating energy requirement is the sum of the heating energy required by each VAV/reheat box. The total air flow over the cooling coil, Vsa, is the sum of the air flows through each zone. And, as before, the energy extracted by the cooling coil, Qcc, can be calculated from an energy balance on the cooling coil.

Qcc = Vsa psa (hma – hsa) - Vsa psa (wma – wsa) hw

Dual-Duct CAV and VAV SystemsA dual-duct commercial heating and air conditioning system is shown below. For simplicity, the figure shows only two zones even though large commercial buildings have many zones. Despite this simplification, building energy use can often be accurately modeled using simple two zone models with an interior and exterior zone.

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Dual-duct system with two zones.

In a dual-duct constant-air-volume system, DD-CAV, warm and cool air streams are mixed in a mixing box at each zone to meet the cooling/heating load in the zone. Note that the total air flow to the zone remains constant. In a dual-duct variable-air-volume system, DD-VAV, zone temperature is maintained by varying the amount of cold or warm air introduced into the zone. When the zone calls for heating, all cooling air is shut off. When the zone calls for cooling, all heating air is shut off. The warm and cold air streams are mixed only during very low-load conditions, to maintain a minimum air flow into the zone. The result is that the total flow of air into the zone varies, rather than remaining constant as in CAV systems. The volume flow rate of air from the cooling coil, Vcc, and heating coil, Vhc, as functions of zone temperature are shown below.

DD-CAV DD-VAV

DD-CAV Heating and Cooling Energy UseIn a DD-CAV system, the flow of supply air through the supply air fan, Vsa, is constant; however, the flow rate through the cooling and heating coils, Vcsa and Vhsa, varies. The fraction of supply

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air to each zone from the cooling coil, fcc, can be determined from an energy balance on each zone and mixing box.

fcc1 = [Qsen1 / (Vsa1 pcpa) + (Thsa- Tra)] / [Thsa – Tcsa]

The flow of air through the cooling and heating coils for each zone, Vcsa and Vhsa, can then be calculated.

Vcsa1 = fcc1 x VsaVhsa1 = (1-fcc1) x Vsa

The total air flows through the cooling and heating coils, Vcsa and Vhsa, are the sum of the flows from each zone. The cooling and heating coil energy use, Qcc and Qhc, can then be calculated from energy balances on each coil.

Qcc = Vcsa psa (hma – hsa) - Vsa psa (wma – wsa) hwQhc = Vhsa pcpa (Tsa – Tma)

DD-VAV Heating and Cooling Energy UseIn a DD-VAV system, the flow of supply air through the cooling coil, Vsa, varies and is the sum of the air flow required by each zone. The procedure to calculate savings begins by calculating the cooling and heating air flows through each zone from an energy balance on the zone and mixing box.

If Qsen1 > 0 then ‘coolingVhsa1 = 0Vcsa1 = [-Qsen + Vha1 pcp Tra – Vhsa1 pcpa Thsa] / [pcpa Tcsa] If Vcsa1 < Vsa,min1 then ‘use method for DD-CAV, but with Vsa1 = Vsa,min1

fcc1 = [Qsen1 / (Vsa,min1 pcpa) + (Thsa- Tra)] / [Thsa – Tcsa]Vcsa1 = fcc1 x Vsa,min1Vhsa1 = (1-fcc1) x Vsa,min1

End ifElse ‘heating

Vcsa1 = 0Vhsa1 = [-Qsen + Vsa1 pcp Tra – Vcsa1 pcpa Tsa] / [pcpa Thsa]If Vhsa1 < Vsa,min1 then ‘use method for DD-CAV, but with Vsa1 = Vsa,min1

fcc1 = [Qsen1 / (Vsa,min1 pcpa) + (Thsa- Tra)] / [Thsa – Tcsa]Vcsa1 = fcc1 x Vsa,min1Vhsa1 = (1-fcc1) x Vsa,min1

End ifEnd if

The total air flows through the cooling and heating coils, Vcsa and Vhsa, are the sum of the flows from each zone. The cooling and heating coil energy use, Qcc and Qhc, can then be calculated from energy balances on each coil.

Qcc = Vcsa psa (hma – hsa) - Vsa psa (wma – wsa) hwQhc = Vhsa pcpa (Tsa – Tma

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Fan and Pump CurvesThe volume flow rate generated by a fan depends on the total system pressure drop. Fans can generate high volume flow rates at low system pressure drops or low volume flow rates at high system pressure drops. A fan curve shows the relationship between total pressure drop and volume flow rate for a specific fan.

Modified graph from ASHRAE Handbook: HVAC Systems and Equipment 2008

It is common for fan manufacturers to publish fan performance data in terms of “static pressure” versus flow. The “static pressure’ in performance data is actually the difference between the static pressure at the fan outlet and the total pressure at the fan inlet.

Pstatic,performance data = Pstatic,2 – (Pstatic,1 + Pvelocity,1)

All fan performance calculations should be performed using total pressure. For example, the methods to calculate pressure loss through ducts and fittings, calculate total pressure loss. In addition, the power requirement of a fan is a function of total pressure loss, not static pressure loss. Thus, for fan calculations, it is important to add the velocity pressure of the air leaving the fan outlet to the static pressure reported in fan performance data to determine the relationship between total pressure and flow for the fan. In some cases manufactures list outlet velocity for the fan. In other cases, outlet velocity can be calculated from outlet dimensions of the fan and airflow. The total pressure of the fan is then:

Ptotal = Pstatic,performance data + Pvelocity,2 = (Pstatic,2 + Pvelocity,2 )– (Pstatic,1 + Pvelocity,1)

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In U.S. dimensional units, the relation is:

Ptotal (in-H20) = Pstatic,performance data (in-H20) + (V2 / 4,005)2 where V2 is in ft/min

System CurveThe total pressure rise that the fan must produce to move air is determined by the duct system. This total pressure of the duct system is the sum of the pressure due to inlet and outlet conditions and the pressure loss due to friction. In a duct system, pressure loss due to friction increases with increasing fluid flow; thus, system curves have positive slopes on pump performance charts. The operating point of a fan is determined by the intersection of the fan and system curves.

The equations for pressure loss from friction through ducts and through fittings are:

Pp = (f L fluid V2) / (2 D)

Pf = kf fluid V2 / 2

These equations clearly show that for a given duct system, the pressure drop is proportional to the square of the velocity, and hence the square of the volume flow rate.

Pfriction = C1 V2 = C2 V2

This quadratic relationship can be plotted on the fan curve to show the “system curve”. The operating point of the fan will be at the intersection of the fan and system curves, as shown below.

Modified graph from ASHRAE Handbook: Systems 2008

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Plotting System CurvesA system curve for a duct system with negligible inlet/outlet pressure differences is a parabola of the form hheadloss = C2 V2. The curve passes through the origin because the inlet/outlet pressure difference, sometimes called the static head, is zero. The coefficient C2 can be determined if the operating point is known by substituting the known pressure drop and flow rate into the equation and solving for C2. The fluid work required to push the fluid through the duct is the product of the volume flow rate and system pressure drop and is represented graphically by the area under the rectangle defined by the operating point.

Controlling Flow Using Throttling/Outlet Dampers: Controlling flow by closing a flow-control valve/damper downstream of the pump/fan increases pressure drop and causes the operating point to move up and left on the pump/fan curve.

This results in relatively small energy savings, since

Wf2 = V2 P2 where V2 < V1 but P2 > P1

Thus, throttling is an energy inefficient method of flow control.

Controlling Flow by Reducing Fan SpeedThe most energy efficient method of varying flow is by controlling pump/fan speed. If the required flow varies over time, speed control is best facilitated by an electronic variable speed drive (VSD) which can continuously and smoothly adjust pump/fan speed as needed. One time reductions in flow are more cost effectively accommodated by replacing the pump/fan pulley with a larger diameter pulley to slow pump/fan rotation or by reducing pump impellor diameter. The following example illustrates energy savings from reducing flow by slowing fan speed.

The figure below shows fan performance at various speeds and a system curve. At full speed with no throttling, the fan operates at B. At B, the pressure is 6.25 in-H20, the volume flow rate is 7,600 cfm and the required power to the fan would reduce to about 12.8 hp. If the flow is throttled, the fan would operate at point A. The volume flow rate at A would be about 6,700

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cfm and the system pressure drop would be about 7.5 in-H20. From the chart, the required power to the fan would be about 13 hp.

To realize significant energy savings at low flow, it is necessary to slow the fan speed. To determine the fan speed required to deliver the initial air flow of 6,700 cfm, it is necessary to develop a system curve for the new duct system. Pressure drop always varies with the square of flow rate. Thus, the equation for a system curve can be written as:

h = C V2

The coefficient, C, for the new system curve can be found by substituting the values of pressure drop and volume flow rate for point B.

C = h / V2 = 6.25 / 76002 = 1.082 x 10-7

Thus the pressure drop through the new duct system at 6,700 cfm would be about:

h = C V2 = 1.082 x 10-7 (6700)2 = 4.86 in-H20

According to the fan curves, the fan would deliver 6,700 cfm at 4.86 in-H20 if the fan speed were slowed to about 2,500 rpm at point C. At this operating point, the fan would require about 9 hp of power. Thus, the savings from reducing flow by slowing the fan would be about:

13 hp – 9 hp = 4 hp

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Controlling Flow by Reducing Pump SpeedThe most energy efficient method of varying flow is by controlling pump/fan speed. If the required flow varies over time, speed control is best facilitated by an electronic variable speed drive (VSD) which can continuously and smoothly adjust pump/fan speed as needed. One time reductions in flow are more cost effectively accommodated by replacing the pump/fan pulley with a larger diameter pulley to slow pump/fan rotation or by reducing pump impellor diameter. The following example illustrates energy savings from reducing flow by slowing pump/fan speed.

The figure below shows pump performance at various speeds and a system curve.

Assume the original operating point, A, is 1,200 gpm at 55 ft-H20. According to the chart, the required power to the pump at this operating point is about 23 hp. Alternately, pump power could be calculated as:

WA = 1,200 gpm x 55 ft-H20 / (3,960 gpm-ft-H20/hp x 0.74) = 22.6 hp

If the flow were reduced to 900 gpm with a flow control valve, the operating point would move along the pump curve to 900 gpm at 62 ft-H20. According to the chart, the required power to the pump at this operating point is about 20 hp. Alternately, pump power could be calculated as:

W = 900 gpm x 62 ft-H20 / (3,960 gpm-ft-H20/hp x 0.70) = 20.1 hp

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Thus, the pump power savings from reducing flow from 1,200 gpm to 900 gpm with a flow-control valve would be about:

22.6 hp – 20.1 hp = 2.5 hp

Alternately, the flow could be reduced from 1,200 gpm to 900 gpm by slowing the pump speed with a VSD. Reducing the pump speed from 1,200 rpm at point A to 900 rpm at point B would reduce the volume flow rate from 1,200 gpm to 900 gpm. The reduced volume flow rate would also generate less friction, and the system pressure drop would be reduced from 55 ft-H20 to 30 ft-H20. The power required to pump a fluid is the product of the volume flow rate and pressure drop; hence, the areas enclosed by the rectangles defined by each operating point represent the fluid power requirements, WA and WB, at the different flow rates.

WA = 1,200 gpm x 55 ft-H20 / 3,960 gpm-ft-H20/hp = 16.7 hpWB = 900 gpm x 30 ft-H20 / 3,960 gpm-ft-H20/hp = 6.8 hp

The power, WB, required by the pump at point B can be read from the chart to be about 10 hp. Alternately, the power could be calculated as:

WB = 900 gpm x 30 ft-H20 / (3,960 gpm-ft-H20/hp x 0.67) = 10.1 hp

Pump power savings would be the difference between PA and PB.

Savings = PA – PB = 22.6 hp – 10.1 hp = 12.5 hp

Alternately, power savings from reducing the volume flow rate can be estimated from the pump affinity laws. Theoretically, pump work varies with the cube of volume flow rate. Use of the cubic relationship would predict:

PB = PA (VB/VA)3 = 22.6 hp x (900 gpm / 1200 gpm) 3 = 9.5 hp

The 9.5 hp predicted by the pump-affinity law is less than the 10.1 hp predicted by the pump curve. This example demonstrates how use of the cubic relationship typically exaggerates savings. In practice, the efficiencies of the VSD, pump and motor typically decline as flow rate decreases, resulting in slightly less savings than would be predicted using this ‘cubic’ relationship. Thus, we estimate that pump/fan work varies with the 2.5 power of flow rather than the cube of flow. Using this relationship, if we measured PA to be 22.6 hp at 1,200 gpm, we would estimate PB for 900 gpm to be about:

PB = PA (VB/VA)2.5 = 22.6 hp x (900 gpm / 1200 gpm) 2.5 = 11.0 hp

Thus, savings would be about:

Savings = PA – PB = 22.6 hp – 11.0 hp = 11.6 hp

This slightly lower estimate of savings incorporates the reduction in motor efficiency, and power loss by the VSD.

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Fan/Pump Affinity LawsThe fundamental fluid mechanic relationships developed thus far can be modified to generate other useful relations between fan parameters. These relationships are known as fan affinity laws. The two most important relationships are derived below.

As shown in the section of system curves, friction head loss is proportional to the square of the volume flow rate.

Pfriction = C1 V2 = C2 V2

By substitution, fluid work is proportional to the cube of volume flow rate

Wf = V Pfriction = V C2 V2 = C2 V3

Since Wf / V3 is constant, it follows that:

(Wf / V3)1 = C = (Wf / V3)2

Wf2 = Wf1 (V2 / V1)3

This relation shows that a small reduction in the volume flow rate results in a large reduction in the fluid work. For example, reducing the volume flow rate by one half reduces fluid work by 88%!

Wf2 = Wf1 (1/2)3 == Wf1 (1/8)(Wf1 – Wf2) / Wf1 = [Wf1 - Wf1 (1 /8)] / Wf1 = 1 – (1/8) = 88%

Another useful relation can be derived from the relationship between volume flow rate V and the rotational speed of the pump fan. In centrifugal pumps and fans, the volume flow rate is proportional to the rotational speed of the pump fan.

V = C RPM

Since V/RPM is constant, it follows that:

(V / RPM)1 = C = (V / RPM)2

V2 = V1 (RPM2 / RPM1)

Thus, volume flow rate varies in proportion to pump/fan speed.

VAV ControlA VFD installed on the AHU fan motor typically uses a pressure sensor signal to control the speed of the fan. The principle types of VFD control are:

Fan-Outlet Control: In fan-outlet control, the pressure sensor is located at the outlet of the fan and sends a signal to the VFD to maintain a duct static pressure set-point at design conditions.

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Supply-Duct Control: In supply-duct control, the pressure sensor is typically downstream in the duct system. The pressure sensor sends a signal to the VFD to maintain a set static pressure at this location to meet the design pressure requirement of the terminals downstream of the sensor.Critical-Zone Control: In critical-zone control, the sensor can be located anywhere in the supply air duct (but typically at the fan outlet); the duct static pressure set-point is dynamically reset to meet the flow requirement of the most critical zone at any moment.

The location of the pressure sensors is depicted in the simple duct system shown below with three zones (1, 2 and 3) served by VAV boxes.

In fan-outlet control, the pressure sensor controlling the VFD is located in position A. In supply-duct control, the pressure sensor controlling the VFD is located at B or C. Location B represents the common rule of thumb of 2/3 of the distance down the duct. Location C represents the location immediately upstream of the most distant zone.

Fan and system curves for different types of control are plotted in Figure 1. When the fan is running at full speed, a design system flow rate of Q0 can be met at a system pressure of P0. Fluid power is the area of the rectangle defined by the pressure drop and flow rate. To visualize fan power requirements as a function of control types and flow rate, consider a change in system flow rate from Q0 to Q’. In an AHU without a VFD the closing of dampers at VAV terminals shifts the system curve, which raises the pressure from P0 to P1 on the fan curve and reduces system flow from Q0 to Q’. The fan power requirement is reduced, but not proportionally with the reduction in flow rate since pressure drop increases.

In fan-outlet control with a VFD, the duct static pressure is maintained at P0. Fan power requirement is reduced proportionally with the reduction in flow rate because the VFD slows the fan speed.

Savings can be increased by controlling VFD fan speed to maintain static pressure near the most distant VAV box(es). This is called supply-duct control. However, supply-duct control risks starving VAV boxes in parallel branches whose pressure requirement exceeds the requirement of the branch where the pressure sensor is located. To mitigate this risk, the pressure sensor is often placed upstream of parallel branches. A common rule of thumb is to place the sensor 2/3 of the way down the longest duct; however, this location negates energy savings from zones downstream of the sensor and results in an artificially high static pressure.

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Critical zone control eliminates this risk by identifying the zone with the greatest duct static pressure requirement at its VAV terminal. In critical-zone control, the duct static pressure set-point is constantly adjusted to meet the flow requirement of the most critical zone; therefore, no minimum pressure needs to be maintained and the system curve can approach zero system pressure when system flow rate approaches zero. Thus, critical-zone control eliminates the risk of starving VAV boxes, and results in greater fan power reduction than supply duct control when the pressure sensor is placed upstream of parallel zones.

In summary, fan-outlet control allows the fan to slow down to a new fan curve while maintaining the set-point pressure of P0. Supply-duct control allows the system pressure to be reduced to P2. Critical-zone control allows the system pressure to be greatly reduced to P3. Since the system flow rate is determined by the terminal unit controllers through the actions of terminal dampers and the AHU system efficiency is not directly controllable, minimizing the duct static pressure set-point is the primary strategy to minimize AHU power. Among the three common control methods, critical-zone control results in the lowest system pressure at any flow requirement and thus offers the greatest AHU power savings.

Figure 1. System and fan curves for AHUs without VFD, with fan-outlet control, with supply-duct control and with critical-zone control illustrate the total system power at reduced system flow.

The implementation of critical zone reset control depends on the particular constraints of a system. Critical-zone control methods depend on collecting information from the VAV terminal units and using the information in an algorithm to dynamically reset the static pressure set-point. Zheng and Liu (2005) proposed resetting the chiller pump VFD set-point pressure based on the pressure drop across the identifiable most resistant loop (MRC). In most commercial buildings, however, the most critical zone changes constantly depending on thermal loads in each zone and thus must be algorithmically identified in the control strategy. Hartman (1993), in

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his Terminal Regulated Air Volume (TRAV) algorithm, described a method to vary fan speed to meet flow requirement in each zone. Warren and Norford (1993) improved upon the TRAV algorithm by resetting the static pressure set-point to minimize the number of starved VAV terminal units. Moult (1999) proposed an algorithm to modulate fan speed in order to maintain the damper position of the most open VAV terminal in a specified range.

Ma et al. (2014) proposed a new algorithm that reduced to modulate fan speed. The proposed strategy uses the number of open terminal dampers as a control input, allowing larger adjustments to the static pressure set-point and maintaining the functionality of the controls even when a terminal damper malfunctions. The control strategy is computationally simple and utilizes data readily available on most current and many older BAS.

The control loop sequence is illustrated in Figure 2. The algorithm executes the following steps in a loop:

1. Poll through all terminal unit controllers and determine the number of terminals with damper position greater than Pos_open.

2. If the number of open terminals is greater than N_max and the duct static pressure is less than P_max, increase the duct static pressure set-point by delP.

3. Else if the number of open terminals is less than N_min and the duct static pressure is greater than P_min, decrease the duct static pressure set-point by delP.

4. Else, maintain current duct static pressure ste-point.5. Delay by an amount of time defined in t_delay.

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Figure 2. Algorithm to execute the proposed control strategy

Fan Energy Use and Control Efficiency in VAV SystemsThe power required to move air through a duct system is:

Power = P x V /

Where, ΔP is the pressure rise across the fan, V is the system flow rate, and η is the total efficiency of the fan, drive, motor and VFD. As shown in the preceding sections, in VAV systems, each VAV box modulates air flow to the minimum needed by the zone it serves. Thus, the total system flow rate V is sum of the required flow rates to each VAV box. Because VAV systems require substantially lower volume air flow rates than CAV systems, fan energy savings can be significant.

However, to minimize power, a control system must also minimize the pressure rise across the fan P to the lowest achievable to meet the required flow rate V. The most energy efficient method of reducing P is by slowing fan speed with a variable frequency drive (VFD). In an ideal system, the pressure will be decreased along the design system curve as the flow rate decreases. When this occurs, the fluid work supplied by a fan or pump varies with the cube of volume flow rate. In this case, the fluid work supplied by the fan, Wf2, at reduced volume flow rate Vsa2 is the product of the original fluid work supplied by the fan, Wf1, at the original volume flow rate, Vsa1, and the cube of the ratio of the reduced and original flow rates:

Wf2 = Wf1 (Vsa2 / Vsa1)3

Unfortunately, this ideal condition is difficult to achieve in real VAV systems in which a single fan serves multiple zones. Depending on the type of control and location of required flow rates to each zone, the actual fluid work produced by the fan is often much greater than this ideal minimum. The ratio of actual to minimum fan power due to the type of control and location of required flow rates is called the VAV control efficiency.

Control Efficiency and Air Flow in Serial Duct SystemsTo understand VAV control efficiency, consider the simple duct system shown below with three zones (1, 2 and 3) served by VAV boxes.

In fan-outlet control, the pressure sensor controlling the VFD is located in position A. In supply-duct control, the pressure sensor controlling the VFD is located at B or C. Location B represents

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the common rule of thumb of 2/3 of the distance down the duct. Location C represents the location immediately upstream of the most distant zone. The table below shows control efficiencies, e_cntrl, for pressure sensor locations A, B and C in a single duct system. Control efficiencies are calculated for the design flow rate of 7,500 cfm and for different combinations of VAV box flow rates that result in a total flow of 5,000 cfm. The table shows that, regardless of the flow scenario, control efficiency increases as the sensor is located farther downstream. The table also shows that even when the pressure sensor is located at the most distant position C, control efficiency is less than 100% whenever the flow through the last zone (3) is less than the design (maximum) flow rate.

Design Case

V2 & V3 Reduced

V1 & V3 Reduced

V1 & V2 Reduced

Balanced Reduction

V1 (cfm) 2,500 2,500 1,250 1,250 1,667V2 (cfm) 2,500 1,250 2,500 1,250 1,667V3 (cfm) 2,500 1,250 1,250 2,500 1,667

Pressure Setpoint at Control Location A

4.00 4.00 4.00 4.00 4.00

e,cntrl = dh,act / dh,max 1.00 0.64 0.69 0.82 0.67Pressure Setpoint at Control Location B

3.00 3.00 3.00 3.00 3.00

e,cntrl = dh,act / dh,max 1.00 0.72 0.78 0.92 0.75Pressure Setpoint at Control Location C

2.00 2.00 2.00 2.00 2.00

e,cntrl = dh,act / dh,max 1.00 0.82 0.85 1.00 0.83

In critical-zone control, the sensor can be located anywhere in the supply air duct, but is typically located at the fan outlet. The duct static pressure set-point is dynamically reset to meet the flow requirement of the most critical zone at any moment. Thus, by design, critical-zone control achieves 100% control efficiency and is the most efficient type of VAV control.

Control Efficiency and Air Flow in Parallel Duct SystemsThe advantages of critical-zone control become even greater in more complicated duct systems with parallel flow. For example, consider the duct system shown below with two parallel ducts and six zones.

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If the pressure sensor were located at C, then it is possible to starve zones 4, 5 and 6 by not providing sufficient pressure at the fan for these zones. Starving can occur when zones 1, 2 and 3 require minimum air flows and zones 4, 5 and 6 require high air flows; the sensor at C can’t detect the high flow, and hence, pressure requirement in the other duct. Unfortunately, this type of unbalanced flow between parallel supply ducts is common; for example, it can be caused by the sun moving from one exposure of the building to the other exposure.

The table below shows control efficiencies and required flow rates and actual flow rates for different flow scenarios in the parallel duct system. In this simulation, the total air flow requirement is reduced from the design condition of 15,000 cfm to 12,500 cfm. As before, the table demonstrates that control efficiency increases as the control location moves downstream. But it also shows that downstream control locations increase the probability of starving zones in the parallel ducts. Flow rates to starved zones are highlighted and their corresponding control efficiencies labelled as not applicable (N.A.). Reducing the probability of starved zones is one reason why VAV control designers often move the pressure sensor upstream of clusters of parallel ducts, albeit at the expense of reduced control efficiency. The 100% control efficiency and minimal probability of starving zones make critical-zone control highly attractive.

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AB

AD

Design

Case

V2 & V3

Reduced

V1 & V3

Reduced

V1 & V2

Reduced

V1 & V4

Reduced

V3 & V6

Reduced

Balanced Reduction in Zone 1,2,3

Balanced Reduction in Zone 4,5,6

V1 (cfm) 2,500 2,500 1,250 1,250 1,250 2,500 1,667 2,500V2 (cfm) 2,500 1,250 2,500 1,250 2,500 2,500 1,667 2,500V3 (cfm) 2,500 1,250 1,250 2,500 2,500 1,250 1,667 2,500V4 (cfm) 2,500 2,500 2,500 2,500 1,250 2,500 2,500 1,667V5 (cfm) 2,500 2,500 2,500 2,500 2,500 2,500 2,500 1,667V6 (cfm) 2,500 2,500 2,500 2,500 2,500 1,250 2,500 1,667

Pressure Setpoint at Control Location A

9.25 9.25 9.25 9.25 9.25 9.25 9.25 9.25

e,cntrl = dh,act / dh,max 1.00 0.77 0.77 0.77 0.68 0.60 0.77 0.77V4_actual (cfm) 2,500 2,500 2,500 2,500 1,250 2,500 2,500 1,667V5_actual (cfm) 2,500 2,500 2,500 2,500 2,500 2,500 2,500 1,667V6_actual (cfm) 2,500 2,500 2,500 2,500 2,500 1,250 2,500 1,667Pressure Setpoint at Control Location B

3.00 3.00 3.00 3.00 3.00 3.00 3.00 3.00

e,cntrl = dh,act / dh,max 1.00 N.A. N.A. N.A. 1.00 0.89 N.A. 1.00V4_actual (cfm) 2,500 2,500 2,500 2,500 1,250 2,500 2,500 1,667V5_actual (cfm) 2,500 2,500 2,500 2,445 2,500 2,500 2,357 1,667V6_actual (cfm) 2,500 1,449 1,449 1,543 2,500 1,250 1,687 1,667Pressure Setpoint at Control Location C

2.00 2.00 2.00 2.00 2.00 2.00 2.00 2.00

e,cntrl = dh,act / dh,max 1.00 N.A. N.A. N.A. 1.00 0.95 N.A. 1.00V4_actual (cfm) 2,500 2,500 2,500 2,500 1,250 2,500 2,500 1,667V5_actual (cfm) 2,500 2,250 2,440 2,412 2,500 2,500 2,020 1,667V6_actual (cfm) 2,500 1,069 1,076 1,133 2,500 1,250 1,665 1,667

CAV to VAV: Energy Saving Recommendation ExampleThe methods for calculating savings shown above can be easily incorporated into computer program, such as ESim, which calculates energy use on an hourly data and sums the results for an entire year. The case study example shown here demonstrates the use of ESim to estimate savings from switching from a CAV to VAV system.

RecommendationWe recommend that retrofitting the air distribution system from a DD-CAV to a DD-VAV system. This would entail replacing the reheat boxes at each zone with VAV boxes with damper controls and adding VSDs to each supply air fan.

Estimated SavingsBuilding HVAC energy use is affected by outside air temperature, outside air humidity, solar radiation, heat transfer through the building envelope, internal electricity use, heat gain from building occupants, the efficiency of chillers and boilers, and the type and control of the air distribution system. Building energy simulation software attempts to model these complex interactions and estimate building energy use.

We modeled energy use in the administration and engineering buildings using the ESim building energy simulation software (Kissock, 1997). ESim uses a description of the building and its

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energy using equipment with hourly meteorological data from TMY2 data files (NREL, 1990) to predict hour-by-hour building energy use. Figure 3.2 below shows simulated and actual electricity use for the engineering and administration buildings. Figure 3.3 shows simulated and actual steam use for the engineering and administration buildings. The close match between simulated and measured energy consumption indicates that the model relatively accurately captures the major energy interactions within the buildings.

Figure 3.2 Simulated and actual electricity use for the engineering and administration buildings with CAV air distribution systems.

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Figure 3.3 Simulated and actual steam use for the engineering and administration buildings with CAV air distribution systems.

Next, we simulated building energy use with a VAV system instead of a CAV air distribution system. The predicted energy consumption with each system, and the savings from switching to a VAV system, are shown in Table 3.1. The large reduction in steam energy use suggests that in most zones internal electricity use and heat from the building occupants is sufficient to heat the zone most of the time. The relatively small reduction in electrical energy use suggests that the current practices of slowing down the fans and employing an economizer cycle successfully reduce fan energy use and mixing of hot and cold air streams; thus, electricity savings are relatively small.

Table 3.1 Simulated energy use and savings from CAV and VAV systems.Electricity (kWh/yr)

Steam(1000 lb/yr)

DD-CAV 9,297,133 11,277DD-VAV 8,899,816 2,255Savings 397,317 9,022Percent Savings 4.3% 80%

To estimate cost savings, we assume that steam costs $6 per thousand pounds and use the average cost of electricity of 4.8 cents per kWh. Based these simulation results, estimated savings from converting to a DD-VAV systems would be about:

9,661,149 kWh/yr x 4.3% = 415,429 kWh/yr415,429 kWh/yr x $0.048 /kWh = $19,941 /yr

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12,547 ThousLb/yr x 80% = 10,038 ThousLb/yr10,038 ThousLb/yr x $6 /ThousLb = $60,266 /yr

$19,941 /yr + $60,266 /yr = $80,167 /yr

Estimated Implementation CostManagement and maintenance personnel roughly estimate that converting the CAV zone mixing boxes to VAV zone mixing boxes would cost about $500 per mixing box. Roughly assuming that there are 100 mixing boxes in the building, the conversion cost would be about:

100 mixing boxes x $500 /mixing box = $50,000

VSD suppliers estimate that installed cost of a VSD for a 100-hp supply fan is about $10,000, and the installed cost of a VSD for a ~50-hp supply fan is about $6,000. Thus, the cost of VSDs for the supply fans would be about:

(6 x $10,000) + (3 x $6,000) = $78,000

Depending on the design, it may be possible to use the current return-air fans without VSDs. Based on these estimates, the total implementation cost would be about:

$50,000 + $78,000 = $128,000

Estimated Simple Payback$128,000 / $80,167 /yr x 12 mo/yr = 19 months

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CAV to VAV: Measured Savings Example 1The graphs below show electricity use, chilled water energy use and hot water energy use before and after a CAV to VAV retrofit of the Texas A&M University Zachry Engineering Center.

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CAV to VAV: Measured Savings Example 2The graphs below show electricity use, chilled water energy use and hot water energy use before and after a CAV to VAV retrofit of a building at the University of Texas.

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VSD Pumping: Commercial Building ExampleA typical constant-flow commercial-building chilled water system is shown below. This system provides constant flows through the condenser and evaporator of each chiller whenever a chiller is operational. The primary chilled water pumps are typically much smaller than the secondary pumps, since a primary pumps only have to move water through the chiller evaporators and not through the entire building. The secondary chilled water pumps are much larger than the primary pumps and provide a constant flow of chilled water to the AHUs. Each AHU varies the quantity of chilled water through the coil and bypasses unneeded chilled water.

Close-ups of a typical piping configuration at the air handler cooling coils in a constant-flow chilled-water supply system are shown below. The three-way valves direct chilled water either through the cooling coil or around the cooling coil via the bypass loop. The flow of chilled water through the cooling coils is varied to maintain the temperature of the air leaving the cooling coils at a constant temperature. In a VSD retrofit, the bypass valves would be closed, and a differential-pressure sensor would be installed between the supply and return headers at the air handler located farthest from the pump. In some cases, it may be necessary to replace the three-way valves with two-way valves if the three-way valves were not designed to handle larger pressure drops in a VSD situation.

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Piping configuration at air handling units.

The greatest pump energy savings come from changing the secondary chilled-water loop from constant to variable flow. This is done by:

Removing or blocking the bypass piping on each AHU Replacing 3-way valves with 2-way valves on each AHU Adding VFDs to the secondary chilled water pumps Controlling the VFDs based on the differential pressure between the supply and return

headers

A typical variable flow secondary chilled water loop system is shown below.

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Modern chillers are designed to accommodate variable flow through the evaporators and condensers. This enables full variable flow chilled water plants. A variable-flow chilled water plant with a flow control and bypass valve to guarantee minimum flow to the chillers is shown below.

The savings from constant to variable flow are represented in the Figure below.

In practice, actual savings depend on DP sensor location

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Bypass Valve

Pump

VSD

Process Lines

Chiller

ΔP

ΔP

ΔP

ΔP

Savings also depend on the quantity of bypass flow.

Bypass Valve

Pump

VSD

Process Lines

Chiller

ΔP

Savings also depend on the dP setpoint. Setting it too high, reduces savings.

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Bypass Valve

Pump

VSD

Process Lines

Chiller

ΔP

Case Study

ReferencesASHRAE Handbook: Fundamentals, 1977 and 1985, ASHRAE.

Engineering Design Reference Manual, 1990, United McGill.

Kreider and Rabl, 1994, Heating and Cooling of Buildings, McGraw-Hill Inc.

Larson, E.D. and Nilsson, L.J., 1991, “Electricity Use and Efficiency in Pumping and Air Handling Systems, ASHRAE Transactions, pgs. 363-377.

McQuiston and Parker, 1994, Heating Ventilating and Air Conditioning, John Wiley and Sons, Inc.

McQuiston, F., Parker, J. and Spitler, J., 2000, Heating, Ventilating and Air Conditioning: Analysis and Design, , John Wiley and Sons, Inc.

Mott, 2000, Applied Fluid Mechanics, Prentice Hall, Inc.

Nadel, S., Shepard, M., Greenberg, S., Katz, G., and Almeida, A., 1991, “Energy Efficient Motor Systems”, American Counsel for an Energy Efficient Economy, Washington D.C.

Tutterow, Energy Efficient Fan Systems, Industrial Energy Technology Conference, Houston, TX.

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