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Modeling the performance of ECM and SCR parallel fan-powered terminal units in single-duct VAV systems.


Figure 1 shows a typical single-duct variable air volume (SDVAV) system that uses fan-powered terminal units (FPTUs) to control the flow of conditioned air supplied by the primary fan. The FPTU has a small fan that is mounted either in series or in parallel with the primary system fan. Figure 2 shows a schematic of a typical parallel FPTU. The unit is called a parallel FPTU because the air flow path of the terminal fan is parallel to the air flow path of the primary fan.


In Figure 1 the terminal unit fan induces air from the return air plenum into the mixing chamber while primary air flows into the mixing chamber. The mixed air stream leaves the mixing chamber and is then supplied to the zone.

First generation FPTUs had constant speed fans where the fan motor speed was regulated with a silicon controlled rectifier (SCR) controller. Once the fan speed was set during the installation and commissioning process, it was operated as a constant speed on or off device. Manufacturers have begun to make available electronically commutated motors (ECM) in FPTUs which allow the speed of the fan to vary as the load in the zone changes. ECM motors also use about one-third of the power of a similar sized SCR controlled motor but they cost more than conventional single speed motors (Edmondson et. al. 2011). The parallel unit fan only operates at low primary air flow conditions when heat is added to maintain space temperature set points. Air properties for the system were calculated as described in the next section.


The simulation model included a number of equations that were used to calculate air properties, determine airflow, leakage, and power, and accounted for the difference in operation when in either cooling or heating mode. The equations and the logic used in the simulation model are described below.

Calculating Changes in the Properties of Air

Air property calculations were simplified by assuming constant properties which, for the temperature ranges of interest to this project, resulted in less than a 0.2% error in the calculations (Cengal et. al. 2005). Operating conditions for supply and return air were assumed to be 55[degrees]F (12.8[degrees]C), 95% RH and 78[degrees]F (25.5[degrees]C), 50% RH respectively. The density of air at 55[degrees]F (12.8[degrees]C), 95% RH was assumed to be 0.076 lb/ft3 (1.22 kg/ m3) and was used to derive Equations 1 and 2. Equation 1 was used throughout the simulations to model sensible heating and cooling of air as it passed through the cooling coil or when heat was added at the terminal as a result of the operation of the terminal unit fan motor or when heat was added to maintain space conditions.


Q = [C.sub.p s] Q ([increment of T]) (1)

Where Q was the heat energy transferred to or from the air, [] was the coefficient that accounted for the specific heat, density, and time, and Q was the volumetric flow rate (ft3/min (l/s)). Equation 2 was used throughout the simulations to model the energy change resulting from a change in the humidity ratio (such as latent heat removal at the cooling coil.

Q = [] Q ([[omega].sub.2] - [[omega].sub.1]) (2)

Where Q was the heat energy transferred to or from the air, [] was the coefficient that accounted for the latent heat, density, and time, and Q was the volumetric flow rate.

The model described in this paper was based on the verified calculation procedure that was implemented as described in Bryant et al. 2009. The program was developed to facilitate the incorporation of the model into existing commercially available programs.

The parallel FPTU has two basic operating modes - fanon and fan-off. Although the ECM control units can be used in a variable speed mode as an adjustable reheat mixing box, only the constant speed "on" or "off" fan operation was considered in this model which means that when the fan was on, it operated at a constant speed.

The minimum primary air flow rate, [Q.sub.p,min], for this project was 20% of design flow. When the fan was operational, the induced air flow rate was 50% of the design flow. The 20% minimum primary plus the 50% return air from the fan totaled 70% of the design flow rate when the fan was operating.

The primary air temperature [T.sub.p], and primary air relative humidity R[H.sub.p], were set at 55[degrees] F (12.8[degrees] C) and 95% respectively. The setpoint temperature of the space [T.sub.sp], was 78[degrees]F (25.5[degrees]C) for both heating and cooling loads. The primary air leakage rate [L.sub.p], was input during the simulation and was set at 5% (0.05), 10% (0.10) or 20% (0.20).

Terminal Unit Leakage

First documented by Furr, et al. (2008) and later confirmed by Cramlet (2008) and Edmondson et al. (2011), parallel terminal units leak during both fan-off and fan-on operation. Leakage is the air that passes through the primary air valve into the mixing chamber that does not pass through the supply port into the conditioned space. The source of this leakage includes seams, penetrations for electrical power into the terminal unit and the back draft damper. The effect on the air distribution system is to increase the amount of air that must be delivered by the primary fan.

The primary leakage rate, [L.sub.p], was input by the user as a decimal percent of the total primary flow rate as shown by Equation 3 and allowed the user to investigate the sensitivity of the system operation to leakage.

[L.sub.p] = [Q.sub.L]/[Q.sub.s]+[Q.sub.L] = [Q.sub.L]/[Q.sub.P] (3)

Calculations for the Parallel FPTU

When the Fan Is Off


Figure 3 shows a flow chart of the algorithm used to model operation of the parallel variable-air-volume (VAV) terminal unit and the zone supplied by the terminal unit. The first step in the analysis was to determine whether or not the zone sensible space load was a cooling load. If the load was a cooling load, then the terminal unit fan was off and the supply air properties were the same as the primary air properties.

The supply air flow rate, [Q.sub.s], required to meet the cooling load was calculated using Equation 4. When leakage was not included and the fan was off, the supply air flow was the same as the primary air flow. When leakage was included and the fan was on, the supply air flow was the primary air flow plus the induced air flow minus the leakage flow.

[Q.sub.s] = [Q.sub.sen]/[]([T.sub.sp]-[T.sub.p])(4)

The properties for the return air were evaluated by performing an energy and mass balance on the return air duct. In all cases, the return air flow rate, [Q.sub.r], was the same as the primary air flow rate, [Q.sub.p]. Equation 5 was used to calculate the mixed return air temperature and was based on an energy balance of the return air stream.

[T.sub.r]= [T.sub.sp] + [L.sub.p] ([ .sub.p] - [T.sub.sp]) (5)

If the primary air leakage rate, [L.sub.p], was zero, the return air temperature, [T.sub.r], was the same as the set point temperature, [T.sub.sp], for the space; otherwise, the return air temperature was lower than the setpoint temperature.

The latent load of the space was combined with the humidity ratio of the supply air to calculate the humidity ratio of the return air. The zone return air humidity ratio, was calculated using Equation 6.

[[omega].sub.r] = [[omega].sub.p] + []/[][Q.sub.p](6)

Calculations for the Parallel FPTU

When the Fan Is On

For the case where the primary air flow rate was at or below the minimum, the analysis of the operation of FPTU proceeded along the path in Figure 3 as if the unit were not in cooling mode. If the sensible load for the zone was not a cooling load, then the terminal unit was in heating mode, the terminal unit fan was turned on, and the analysis followed the "fan-on" operation.

Once the fan was turned on and the primary flow delivered to the space was set to the minimum, the downstream static pressure was calculated using Equation 7.

[P.sub.dwn] = K[([Q.sub.s]/1000).sup.2](7)

The "K" factor in Equation 7 is the loss coefficient for the duct system that connects the supply port of the FPTU to the conditioned space. The "K" for the zone was calculated with a user-defined downstream static pressure at design air flow with Equation 7 for each zone. The calculated "K" was then used at part load conditions to calculate the downstream static pressure, [P.sub.dwn].

Equation 8, developed by Furr et al. (2008), was used to calculate the flow rate of the fan. The coefficients for the SCR controlled motors were from Furr et al. (2008). The coefficients for the ECM motors were from Edmonson et al. (2011).

[Q.sub.f] = [C.sub.1]+[C.sub.2][V.sup.2]+[C.sub.3]V+[C.sub.4][P.sub.dwn]+[C.sub.5][P.sub.dwn] (8)

After the downstream static pressure was determined, the fan power was calculated using Equation 9 which was developed by Furr et al. (2008). The coefficients for the SCR controlled motors were from Furr et al. (2008). Coefficients for the ECM motors were from Edmondson et al. (2011).

[W.sub.f] = [C.sub.1]+[C.sub.2][V.sup.2]+[C.sub.3]V+[C.sub.4][P.sub.dwn]+[C.sub.5] [P.sub.iav] (9)

Equation 10 was developed by Furr et al. (2008) and was used to calculate the value of [P.sub.iav] in Equation 9. As before, coefficients for the SCR parallel FPTUs were from Furr et al. (2008) and the coefficients for the ECM parallel FPTUs were from Edmondson et al. [(2011).

[P.sub.iav]= [C.sub.1] + [C.sub.2][Q.sub.p] (10)

In Equations 8 and 9 the term "V" refers to the "voltage" applied to control the terminal unit fan. In the case of the SCR controlled motors, "V" refers to the SCR output voltage which ranged from 0 - 277 VAC. For the ECM controlled motors "V" meant either a control voltage of 0 - 10 VDC or a controller setting of 0 - 100%. Edmondson et al (2011) detailed the variations in the use of the voltage (V) as it related to the equations. For the model, the user inputs a flow rate of 0 - 100% in the FPTU setup portion of the user interface and the software automatically converts to the proper "V" range for the calculations.

The fan power, [W.sub.f], in Watts was added to the return air inducted into the fan and the temperature of the air supplied to the mixing chamber by the fan, [T.sub.f], was computed:

[T.sub.f] = [T.sub.r]+3.413[W.sub.f]/[C.sub.s][Q.sub.f] (11)

where 3.413 was the factor used to convert from Watts to BTU/hr. The supply air temperature, [T.sub.sest], calculated from an energy balance was:

[T.sub.sest] = [Q.sub.p][T.sub.p]+[Q.sub.f][T.sub.f]/[Q.sub.p]+[Q.sub.f] (12)

Equation 13 was used to estimate the temperature of the return air, for the zone. The temperature of the supply air and the temperature of the return air were estimated using a procedure that iterated through equations 12 and 13 until both the supply temperature and the return air temperature changed by less than 0.01[degrees]F between iterations.

[T.sub.r] = [Q.sub.1][T.sub.m]+[Q.sub.s][T.sub.sp]/[Q.sub.p]+[Q.sub.f] (13)

Once the estimated supply air temperature, [T.sub.sest], was known it was compared to the supply air temperature required to maintain space conditions, [T.sub.sreq] which was calculated using Equation 1. Here, the sensible heat transfer was the summation of the space load and the cooling delivered to the space by the minimum primary air setting supplied by the user.

If the estimated temperature differed from the required temperature, then the primary air flow rate was increased so that the extra heat energy provided by the terminal unit blower motor or heat energy was added to increase the supply temperature to the required value.

Static Pressure Calculations--Minimum Static Pressure

Equation 14, developed by Furr et al. (2008), was used to calculate the flow rate for the primary air for both the SCR and ECM FPTUs for a given damper setting and differential pressure. The coefficients for SCR parallel FPTUs were from Furr et al. (2008) and the coefficients for ECM parallel FPTUs were from Edmondson et al. (2011).

[Q.sub.p] = [C.sub.1] (1+[C.sub.2]S+[C.sub.3][S.sup.2]) [square root of ([increment of P])] (14)

A binary search algorithm determined the differential pressure, required to produce a primary air flow [Q.sub.p], needed to maintain zone conditions. The minimum static pressure required that the damper be in the fully open position so the value of "S" was zero (0) degrees for this calculation. Once [increment of P] was determined, the downstream static pressure was calculated using Equation 7. The upstream static pressure was determined by adding [increment of P]o the downstream static pressure.

Figure 4 shows that the upstream static pressure for the parallel terminal unit drops with the space loads but that it jumps to almost half of the full-load levels when the terminal unit fan is in operation. The increase in the upstream static pressure at low-load conditions causes a system-wide increase in the available upstream static pressure for all zones.

Calculations for the SDVAV System with Parallel FPTU

From Figure 1, the components of the SDVAV system consisted of FPTUs that control the zones, return air ducts, exhaust and fresh air ducts, a pre-heat coil (PHC), the primary air fan (Fan), the primary air cooling coil (CC), and the primary air distribution ducts. Figure 5 shows a diagram that identifies the calculation sequence used in the simulation program.


Zone Calculations

The system control algorithm started at the first hour of the year and on an hourly basis it executed the zone simulation procedure for the parallel FPTU illustrated in Figure 3. The hourly procedures provided information such as the required primary flow, return air flow, minimum static pressure, heat added, and FPTU fan power that resulted from the operation of the zone. After all of the zone calculations were completed, the system return air properties were calculated.


Return Air

Equations 15 and 16 were used to calculate the return air mixed air temperature and humidity ratio.

[T.sub.ra] = [summation over ([.sub.i.sup.n])][Q.sub.i]x[T.sub.i]/[summation over ([.sub.i])][V.sub.i] (15)

[[omega].sub.ra] = [summation over ([.sub.i.sup.n])][[omega].sub.i][Q.sub.i]/[summation over ([.sub.i])][V.sub.i] (16)

Exhaust/Fresh Air

The temperature and humidity ratio after the exhaust/fresh air intake was calculated using Equations 17 and 18 where X is the fraction of fresh/exhaust air introduced into the return air stream which was 20% (0.20).

[T.sub.rapf] = X[T.sub.oa]+(1-X)[T.sub.ra] (17)

[[omega].sub.rapf] = X[[omega].sub.oa]+(1-X)[[omega].sub.ra] (18)

Primary Fan

Equation 19 was used to calculate the primary air flow, [Q.sub.p], as a function of the fan speed, S, in revolutions per minute and static pressure P, in inches of water (in. w.g.). Equation 20 was used to calculate the fan power as a function of fan speed, S, in rpm. Equation 19 was a correlation developed using published data for a fan with an operating range that covered the simulations included in this study. Equation 20 was included in the published data provided by the fan manufacturer and was modified to include the efficiency of the fan motor, fan, which was 85%.

[Q.sub.p] = [a.sub.1]+[a.sub.2]P+[a.sub.3][P.sup.2]+[a.sub.4]S+[a.sub.5][S.sup.2]+[a.sub.6]PS+[a.sub.7][P.sup.2][S.sup.2] (19)

[W.sub.f] = 746[(S/1631).sup.3](1/fan) (20)

where [a.sub.1]... [a.sub.7] were constants that were inputs by the user and fan was the user defined efficiency of the fan motor in percent. The binary search algorithm held the static pressure constant and searched for the fan speed that would produce the required amount of flow. The static pressure used in Equation 19 was the maximum value of the minimum required static pressure for all of the zones.

Pre-Heat Coil

The return air entered the pre-heat coil after the fresh air was introduced into the system (Figure 1). If the temperature of the pre-fan return air, [T.sub.rapf], was below the pre-fan mini-mum temperature, [T.sub.pfm], then heat energy was added so that the pre-fan return air temperature was at the minimum. Equation 21 was used to calculate the amount heat energy, [], required to warm the return air to the minimum pre-fan temperature.

[] = [][Q.sub.ra]([T.sub.pfm]-[T.sub.rapf]) (21)

The minimum entering fan temperature was calculated by subtracting the fan temperature rise from the cooling coil leaving air temperature. If the temperature of the return air was above the minimum entering air temperature for the fan, then no heat was added.

The increase in the temperature of the air, [[increment of T].sub.f], as it passed through the fan was calculated with Equation 22 using the power consumed by the primary fan, [W.sub.f].

[[increment of T].sub.f] = 3.412[W.sub.f]/[C.sub.s][Q.sub.ra] (22)

Primary Cooling Coil

The return air cooling coil entering temperature, [T.sub.raec], was determined by adding the temperature rise across the primary fan to the pre-fan return air temperature, [T.sub.rapf]. The sensible cooling load, [Q.sub.ccsen], handled by the cooling coil was found using Equation 23.

[Q.sub.ccsen] = [C.sub.s][Q.sub.p] ([T.sub.raec] - [T.sub.p]) (23)

The latent load, [Q.sub.cclatt], on the primary cooling coil was calculated using Equation 24.

[Q.sub.cclatt] = [C.sub.l][Q.sub.p] ([[omega].sub.raec] - [[omega].sub.p]) (24)


The system model was used to predict the operation of a commercial building for one year at five different geographical locations around the United States using TMY2 (NREL 1995) weather data. The base case operating conditions used Houston weather data for the analysis. Following the base case operation, a sensitivity analysis was done that investigated changes in the air leakage and location.

The building model consisted of five zones with exterior exposures covering a range of loads resulting from weather and solar effects while the interior, or core zone, was dominated by internal thermal loads. The hourly space loads were normalized to the peak cooling capacity of the FPTUs and were based on the hourly loads generated by modeling a single story rectangular structure with four perimeter zones and a single core zone.

This technique allowed modeling of the operation of the facility at various geographic weather locations while maintaining the peak cooling loads within the capacity of the selected FPTU. This method also eliminated any bias in the simulation results if the VAV terminal units were either over or undersized when moved to different geographic locations. The procedure used to generate the normalized loads at each of the five weather locations was described by Bryant et al. (2009).

Base Case

The base case settings were intended to mimic the theoretical best case operating patterns of buildings that used fan powered terminal units with no leakage. Table 1 shows the simulation results for the base case for systems with ECM_P12A, and SCR_P12A terminal units.
Table 1. Simulation Results Summary for the Base Case Houston
Location (No Leakage Case)

Item                  SCRP12A    ECMP12A   % Difference

Total Plant Energy         156        156             0

Total Cooling Plant         56         56             0

Primary Fan Energy           3          3             4

Terminal Unit Fan            9          3           -66
Energy (MWh)

Heat Added (MWh)            87         93             7

Fan + Heat Energy           96         96             0

Maximum Static       0.25 (65)  0.37 (92)     0.16 (41)
Pressure in. w.g.

Minimum Static       0.024 (6)  0.036 (9)     0.15 (38)
Pressure in. w.g.

The total plant energy was the summation of the cooling plant, primary fan energy, terminal unit fan energy and heat added. The total cooling energy was the cooling energy supplied by the primary cooling coil over the entire year. Because the model developed was intended to be used within a building simulation program, neither detailed cooling nor heating plant simulations were done and the energy quantities used by the mechanical systems were not estimated by the simulation.

The cooling plant energy was estimated using an annual average Energy Efficiency Ratio (EER) of 11.0 which was taken from the minimum efficiency standards for package rooftop air-conditioning units (Energy Star 2010). The heat energy added was from the heating plant.

The ECM_P12A FPTU used 0.2% less total plant energy than the SCR_P12A unit. The annual operating hours were the same for both the SCR_P12A and the ECM_P12A but the ECM FPTUs used 65% less unit fan energy as a result of the more efficient ECM motor. However, because the terminal unit fan motors are in the air stream and parallel terminal unit fans operated only during heating mode, the fan heat energy not added to the air stream from the more efficient ECM motor had to be offset by added space heating.

Case 1--Sensitivity to Unit Manufacturer
Table 2. Simulation Results for Three ECM Parallel FPTUs
(No Leakage Case)

                              Parallel FPTU
Item                  ECMP12A    ECMP12B     ECMP12C

Total Plant Energy         156        156         156

Total Cooling Plant         57         56          57

Primary Fan Energy           3          3           3

Terminal Unit Fan            3          4           3
Energy (MWh)

Heat Added (MWh)            93         92          93

Fan + Heat Energy           96         96          96

Maximum Static       0.37 (92)  0.27 (68)  0.41 (103)
Pressure in. w.g.

Minimum Static       0.036 (9)  0.028 (7)  0.040 (10)
Pressure in. w.g.

Table 2 shows the simulation summary results for the ideal case simulation using parallel ECM FPTUs P12A, P12B and P12C. The replacement of the ECM_P12A terminal unit with either of the other two manufacturer's units had no significant impact on the total cooling plant energy, total cooling energy or primary fan energy. ECM_P12B used more terminal unit fan energy than ECM_P12A or ECM_P12C but the additional energy was used during heating mode and was offset by the reduced amount of added heat required by the zone so there was no net increase in the energy consumed by the system as shown by the value of the "fan + heat" which was 96 MWh for all three units.

The sensitivity of the simulation to unit manufacturer for the SCR parallel units was previously documented in the report for ASHRAE RP 1292 (Davis et al. 2007) and showed no difference in the results based on the selection of the manufacturer.

Case 2--Base Case Operation at Five Locations in the United States
Table 3. Energy Usage of SCR_P12A and ECM_P12A by Location for
24 Hour Operation (No Leakage Case)

City       Total Plant        Total        Primary Fan
             Energy          Cooling         Energy
              (MWh)           Plant           (MWh)

                   SCR  ECM      SCR  ECM          SCR  ECM

Houston            156  156       56   56            3    3

Phoenix            140  140       60   60            3    3

Chicago            152  151       44   44            3    3

New York           143  143       46   46            3    3

San                134  134       45   45            2    2

City       Terminal       Heat Added       Heat Added
           Unit Fan         (MWh)            + Fan
            Energy                             Energy
             (MWH)                               (MWh)
                SCR  ECM         SCR  ECM         SCR  ECM

Houston           9    3          87   93          96   96

Phoenix           9    3          68   74          77   76

Chicago          11    4          95  102         105  105

New York         10    4          84   91          95   94

San              10    4          76   83          87   87

Table 3 shows the energy used by each system component at each weather location for SCR_P12A and ECM_P12A. Table 4 shows a summary of the change in energy consumption for all five locations when SCR_P12A was replaced with ECM_P12A. Both tables show that there was no significant difference in the operation of the ECM_P12A and the SCR_P12A for the no leakage case.

Case 3--Impact of Leakage Rates of 5%, 10%, and 20%
Table 4. Percentage Change in Energy Usage for ECM_P12A Compared to
SCR_P12A by Location for 24 Hour Operations (No Leakage Case)

City       Total    Total   Primary  Terminal  Heat     Heat
           Plant   Cooling    Fan    Unit Fan  Added  Added +
           Energy   Plant   Energy    Energy            Fan

Houston      -0.1     -0.1      3.6     -65.7    6.7     -0.3

Phoenix      -0.2     -0.1      3.5     -66.7    8.2     -0.4

Chicago      -0.1      0.0      3.7     -66.7    7.1     -0.3

New York     -.02      0.0      3.7     -65.7    7.6     -0.3

San           0.0     -0.1      3.8     -65.7    8.8      0.0

Table 5. Simulation Results for P12A for 24 Hour Operation with 5%,
10%, and 20% Leakage Rates

                                Leakage Rates
Item          Base (0%)        5%           10%          20%
              SCR    ECM    SCR    ECM    SCR   ECM    SCR    ECM

Total Plant    156    156    161    161    165   165    174    175

Total           56     56     58     58     59    59     61     61

Primary Fan      3      3      4      4      4     4      5      5

Terminal         9      3      9      3      9     3      9      3
Unit Fan

Heat Added      87     93     90     96     93    99     98    105

Fan + Heat      96     96     99     99    102   102    108    108

Max Static    0.26   0.37   0.26   0.38   0.27  0.40   0.28   0.43

Pressure      (65)   (92)   (66)   (96)   (67)  (99)   (69)  (108)
in. w.g.

Min Static   0.024  0.036  0.024  0.036  0.024  0.04  0.028   0.04

Pressure       (6)    (9)    (6)    (9)    (6)  (10)    (7)   (10)
in. w.g.

Table 5 shows the results of the simulation when leakage was added to the base case (Houston) operation for the SCR_P12A and ECM_P12A. The total plant energy increased by 3.0%, 5.8% and 11.7% over the base case when leakage was 5%, 10% and 20%, respectively, for both the ECM_P12A and SCR_P12A units. For both the ECM_P12A and SCR_P12A units, the cooling plant energy increased by 2.1%, 4.1% and 8.7% over the base case when leakage was 5%, 10% and 20%, respectively and the primary fan energy increased by approximately 13.0%, 25% and 55%, respectively. For the SCR_P12A unit, the heat added increased by 3.3%, 6.4% and 13.0% for 5%, 10% and 20% leakage rates, respectively. For the ECM_P12A unit, the heat added increased by 3.2%, 6.4% and 13.0% for 5%, 10% and 20% leakage rates, respectively.


A system model was developed that used the performance characteristics of ECM and SCR FPTUs that were measured by Furr et al. (2008), Cramlet (2008) and Edmondson et al. (2011). The measurements from laboratory experiments combined with the sensitivity analysis indicated that leakage can have a substantial impact on the system performance as well as the overall energy consumption.

Consideration should also be given to what happens to the SDVAV system during low-load conditions when one of the terminal unit fans turns on. The activation of the terminal unit fan resulted in the increase in the pressure of the primary fan which increased the overall operating cost of the system. This was shown clearly in Figure 4. The large change in static pressure is quite evident upon fan activation resulting in an increase in static pressure at the primary fan with a coincident increase in operating energy.

The results from the simulation showed:

1. Terminal unit fan energy was reduced by approximately two thirds when ECM motors were used compared to SCR controlled motors.

2. Because the parallel unit fan motors operated only in heating mode, the heat energy that was added to the air stream from the fan motor was reduced and had to be made up by supplemental heating. If the supplemental heat energy is supplied by electric resistance heating then this is a one-for-one trade which resulted in no net energy savings.

The results of this study show that paying a premium cost for an ECM fan motor in a parallel terminal unit will provide no significant cost reduction in energy use. The energy comparisons show virtually identical energy consumption for these two motor types regardless of operation or load. Current conventional wisdom is to install this technology for both parallel and series FPTU applications. These results indicate that for parallel FPTU application, the additional cost is not warranted. This study shows that paying a premium for an ECM motor in a parallel FPTU would not be cost effective.


The authors wish to thank the following individuals and organizations that provided support and advice for this project: Gus Faris of Nailor Industries, David John of Metal Industries, Inc., Dan Int-Hout of Krueger Manufacturing Co., Ron Jordan and Doug Fetters of A.O. Smith Electrical Products Company, Floyd Blackwell of Regal Beloit Corporation, Gaylon Richardson of Engineered Air Balance Co., Inc., and Jack Stegall of Energistics Laboratory. The authors would like to acknowledge the Qatar National Research Fund, which helped support the research for this project.


[] = latent heating/cooling specific heat

[] = sensible heating/cooling specific heat

[C.sub.1]...[C.sub.n] = coefficients of an equation

K = constant relating [increment of P] to [Q.sup.2] (in w.g.-mi[n.sup.2] / f[t.sub.6])

[L.sub.p] = primary air leakage as decimal percent of primary flow

[P.sub.dwn] = down stream static pressure

[P.sub.iav] = primary air inlet valve entering air velocity differential pressure

[increment of P] = pressure drop

Q = volumetric air flow rate

Q = heat energy transferred

[Q.sub.aux] = auxiliary sensible heat added

[Q.sub.f] = fan power

[Q.sub.f] = flow rate of the terminal unit fan

[Q.sub.ind] = induced air flow rate

[Q.sub.L] = leakage flow rate

[] = latent space load

[Q.sub.min] = minimum sensible cooling supplied to a zone

[Q.sub.p] = primary air flow rate

[Q.sub.pmin] = primary minimum air flow rate

[Q.sub.r] = return air flow rate

[Q.sub.s] = supply air flow rate

[Q.sub.sen] = sensible space load

R[H.sub.p] = primary air relative humidity (%)

S = damper position in degrees

[T.sub.L] = leaked air temperature

[T.sub.p] = primary air temperature

[T.sub.pfest] = estimated pre-fan mixed air temperature

[T.sub.ref] = required pre-fan mixed air temperature

[T.sub.r] = return air temperature

[T.sub.s] = supply air temperature

[T.sub.sest] = supply temperature based on current conditions

[T.sub.sreq] = supply temperature required to maintain

[T.sub.sp] = set point temperature

V = FPTU fan control voltage

[W.sub.f] = fan power (Watts)

[[omega].sub.r] = return air humidity ratio

[[omega].sub.p] = primary air humidity ratio

[rho] = density of the air


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Davis, M. A., Bryant, J. A., O'Neal, D. A., Cramlet, A. 2007. ASHRAE Project 1292-RP, Phase II Final Report. Energy Systems Laboratory, Texas A&M University - College Station.

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Michael A. Davis, PhD


John A. Bryant, PhD, PE


Dennis L. O'Neal, PhD, PE


Michael A. Davis is director of laboratories at New York University Abu Dhabi, Abu Dhabi, United Arab Emirates. John A. Bryant is an associate professor at Texas A&M University at Qatar, Doha, Qatar. Dennis L. O'Neal is Holdredge/Paul Professor and associate dean of engineering at Dwight Look College of Engineering, Texas A&M University, College Station, TX.
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Author:Davis, Michael A.; Bryant, John A.; O'Neal, Dennis L.
Publication:ASHRAE Transactions
Article Type:Report
Geographic Code:1USA
Date:Jan 1, 2012
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