hpac a fresh look at fans-final

8/20/2019 HPAC a Fresh Look at Fans-Final

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Public Interest Energy Research (PIER) program.2

The project—Large HVAC Integration3—focuseson the air side of built-up VAV systems, including fan selection, fan staging, and supply-pressure

controls, with the ultimate goal of producing design guides for air-side-system design and control.The project includes detailed sitemonitoring of five built-up VAV

systems. The parameters include fan energy, airflow,fan static pressure, duct pressure, and terminal-unitdemand. In the process of doing the research, which

will conclude in June, the team developed a new simulation model of fan-system efficiency as afunction of flow and pressure. The model will

be presented in a symposium paper during the American Society of Heating, Refrigerating and Air-Conditioning Engineers’ (ASHRAE’s) 2004 Winter Meeting. Details on the model are availableon the project Website.3

This article describes our preliminary findingsat Site 1, a high-tech office building with datacenters located in San Jose, Calif.4 The building isthree stories tall, approximately 105,000 sq ft, and

Fan energy is a large part of mechanical-system energy usage. In California’s new commercial-building stock for built-up

systems, it accounts for 1 terawatt-hour of electric-energy usage per year, representing approximately 50 percent of allHVAC energy usage.1 Standardpractices and operation can lead todesigns that use as much as twicethe energy of optimized designs. This can be attrib-uted in part to the lack of analysis of fan-systemperformance across the full range of operation andthe lack of tools to perform such analysis. Thisarticle reports on a current publicly funded researchproject on optimized design techniques for large

variable-air-volume (VAV) systems and control.Preliminary findings from one of five monitoredsites are presented. These results provide insightinto both methods of design and means to save fanenergy.

BACKGROUND

The authors are part of the team performing theresearch under the California Energy Commission’s

28 May 2003 • HPAC Engineering

Preliminary findings

from California

research project

provide insight into

fan design and

energy savings

By MARK HYDEMAN, PE,and JEFF STEIN, PE,

Taylor Engineering LLC

Mark Hydeman, PE, is a principal and Jeff Stein, PE, a senior engineer for Taylor Engineering LLC in Alameda,

Calif., a mechanical consulting firm specializing in the design and commissioning of HVAC and control systems for commercial buildings. The firm actively participates in the development of building energ y codes andbuilding-science research.

A F r e s h L o o k a t

Fans


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owner-occupied and hasa central air handler and VAV boxes withhot-water reheat. The central air handlerconsists of two 66-in. plenum fans inparallel with barometric backdraftdampers at the fan inlets (Figure 1). Each

of these supply fans was designed toprovide 72,500 cfm at 4-in.-wc staticpressure (145,000 cfm total). The supply fans have 75-hp motors and variable-speed drives (VSDs). The VSDs arecontrolled to maintain a fixed ductstatic-pressure setpoint of 1.5 in. wc, asmeasured at the bottom of the duct riser.This air-handling system runs continu-ously to serve late-night programmersand server rooms.

Figure 5 (Page 35) presents the meas-ured airflow and static pressure. Thestatic-pressure measurements, including

t h epressure

drop acrossthe inlet backdraft

dampers,5 were takenfrom the fan inlet plenum to

the discharge plenum. The red line isthe theoretical sys-tem curve for thefan, assuming a fixedminimum static-

pressure setpoint of 1.5 in. wc and thesystem design staticpressure.6 Figure 2shows the manufac-turer’s fan curve,

while Figure 3 showsthe same fan curvein three dimensions,

with efficiency plot-ted in color on the

Z-axis as a functionof airflow and fanstatic pressure. For

this fan, the “do not select” or “surge” linecoincides with the line of peak efficiency.7

FINDINGS FROM SITE 1

The new fan model and the moni-tored data were used to answer severalquestions:

• What fanshould the engi-neer have selected,and how does it

compare to the onethat was selected?

• What is theoptimal staging of the two fans, tak-ing into accountthe efficiencies of the fans, motors,belts, and VSDs?

• What is theeffect of resetting

the fan static pres-sure according toVAV-box demand?

29HPAC Engineering • May 2003

FIGURE 1. Plan view of Site 1 mechanical

room.


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Fan selection. Although they are moreexpensive and less efficient than housedcentrifugal fans,8 plenum fans take upless space and are quieter in the criticallow-frequency bands.9 The engineer forSite 1 probably selected plenum fans tosave space and provide a hedge for theacoustical design.

While keeping an eye on sound-powerdata and motor size, the engineer likely

It is important to note that fan selec-tion generally applies only to new con-struction, while staging and reset apply toboth new construction and retrofits. Intheory, one could replace an existing fan,but this is unlikely to be cost-effective—unless the existing fan is at the end of itsservice life. Each of these issues will beexamined in detail in the paragraphs thatfollow.


F A N S

FIGURE 2. The manufacturer’s fan curve.

E f f i c i e n c y ,

p e r c e n t 54

36

18

0

D i f f e r e n t i a l p r e s s u r e

4.8

3.6

2.4

1.2

Cu b ic fee t per m inu te 4 0,

0 0 0 8 0, 0 0 0

1 2 0, 0 0 0 1 6 0, 0 0 0

2 0 0, 0 0 0

66%

52%

37%

23%

9%

FIGURE 3. Fan efficiency as a function of airflow and static pressure.


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Circle 184Circle 200

used the manufacturer’s selection pro-gram to pick a fan with a design pointthat was efficient, but not too close to thesurge or do-not-select region.

Given that the manufacturer did notmake a plenum fan large enough tomeet the total design flow efficiently,the engineer decided to use two fans inparallel. This had the benefit of redun-dancy (reduced exposure to fan failure)and improved performance under low-load conditions.10

The engineer probably looked atsomething similar to what is shown inFigure 4 and chose the 66 PL-A because

it provided a good combination ofefficiency, relative cost, acoustics, andmotor size. It is unlikely that any simula-tion or evaluation of part-load operation

was considered.Fan selection for Site 1 was evaluated

by simulating a range of potential selec-tions against the monitored fan-load

profile (total cubic feet per minute anddifferential pressure across the fan).Figure 5 shows the monitored data anddesign point. Figure 6 shows that the

system spends the majority of its time atvery low flows and never came close to thedesign condition during the monitoring period. Figure 6 has the same X-axis scale


FIGURE 4. Typical fan-selection software.

F A N S


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Circle 192

as Figure 5; together, they display the fre-quency of operation for eachregion of operation. As Figure 5 shows,the actual system curve appears to runthrough 1.5 in. at 0 cfm. A consequenceof a high fixed static-pressure setpoint isthat the fan operates in the surge regionat low loads (Figure 2). Because they wereoversized, the fans operate in the surgeregion more than 60 percent of the time.

As described below, aggressively resetting the static-pressure setpoint would reduceor eliminate this problem.

Figure 7 shows the efficiency of theSite 1 fan system along the system curveshown in Figure 5 as a red line. Figure 7shows that the fan efficiency goes upand down as cfm changes and as thesystem stages from single- to dual-fanoperation. According to our simulations,the average fan efficiency during themonitoring period was 57 percent.

Several other fan selections were simu-lated against the actual measured load.These included other sizes of plenumairfoil fans and several sizes of housed air-


F A N S

D i f f e r e n t i a l p r e s s u r e a c r o s s f a n ,

i n c h e s o f w a t e r

4.0

3.5

3.0

2.5

2.0

1.5

1.0

0.5

0.0

Total cubic feet per minute

0 20,000 40,000 60,000 80,000 100,000 120,000 140,000

Actual dataPerfect system curve (i.e., static-pressure reset)Design point (145,000 cfm at 4.0 in.)

Likely actual system curve (no static-pressure reset)

FIGURE 5. Monitored data from Site 1.

R e c o r d s ( 1 5 - m

i n d a t a )

7,000

6,000

5,000

4,000

3,000

2,000

1,000

0

Airflow bin (cubic feet per minute)

0 20,000 40,000 60,000 80,000 100,000 120,000 140,000

The system spends many hours at nightbetween 20,000 and 25,000 cfm

FIGURE 6. Histogram of Site 1 cfm.


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foil fans. Figure 8 shows simulationresults for the base case and alternate fanselections. The first two lines in Figure 8compare the actual fan staging with anoptimal staging using the same fan (seediscussion of optimal staging later in thisarticle). Figure 8 shows that the 66-in.plenum fan (66 PL-A) used less energy than the 49-in. (49 PL-A) fan and aboutthe same as the 54-in. (54 PL-A), 60-in.(60 PL-A), and 73-in. (73 PL-A) fans.

It is interesting to note that the annualenergy ranking from the simulation(Figure 8) does not follow the efficiency ranking from the manufacturer’s selection

program (Figure 4). There are severalreasons for this. One has to do with thevalleys and peaks (or “sweet spots”) inthe efficiency profile of each fan (see, forexample, Figure 7) compared with theload profile. Different fan systems havepeaks and valleys at different spots.

Figure 8 also shows that housed airfoilfans (marked A-DWDI) are consistently more efficient than plenum fans (markedPL-A). Of course, this is not necessarily a fair comparison because of the spacerequirements and acoustical issues associ-ated with housed fans.

Fan staging. In the simulations, themost efficient control sequence for stag-ing the fans was employed. As mentionedpreviously, the first two lines in Figure 8compare the actual fan staging with anoptimal staging using the same fan. Basedon the monitored data, the existingcontrol scheme runs one fan almost all of the time, with the second fan brought

on for a few hours each week, mostlylate in the afternoon. The optimal staging routine in the simulation compared theenergy use of one and two fans at eachoperating condition, selecting the option

with the lowest energy that could meetthe load based on the installed motorhorsepower and fan curve. For this loadprofile, it almost always was better to runone fan because the majority of the hoursare at very low load conditions.11 Thus,

switching from the base case to the opti-mal staging with the same fan (66 PL-A)did not save much energy (Figure 8).

In a real control system, staging logicmust be based on a measured input, suchas supply-fan VSD speed. Also, it needsto have a control differential or otherfeature to avoid short-cycling the fans.The authors sought to develop recom-mendations for staging the fans based onspeed. Figures 9 and 10 present optimalstaging speeds for this building with and

without supply-pressure reset. In thesefigures, the green and blue lines represent

the power consumed by the fans as they run up and down the system curve. Thered and yellow lines represent the speedof these fans at each condition. Thedashed lines show the speed at the opti-mal staging point. For the 66 PL-A fansystem and for a system curve that runsthrough 1.5 in. (i.e., fixed static-pressuresetpoint), the optimal point to stage fromone fan to two is when the fan exceedsabout 77-percent speed. Conversely, the


F a n e f f i c i e n c y , p e r c e n t

7570

65

605550

4540

3530

2520

1510

50

Cubic feet per minute

0 20,000 40,000 60,000 80,000 100,000 120,000 140,000

Stages from one to two fans operating(this is the optimal staging point forthis fan system on this system curve)

FIGURE 7. Fan efficiency.

73 A-DWDI optimal staging

$0 $5,000 $10,000 $15,000 $20,000 $25,000

$20,474

$20,206

$21,666

$20,232

$19,713

$19,993

$17,424

$17,069

$16,835

$17,081

$18,065

66 PL-A base-case staging

66 PL-A optimal staging









Annual electricity cost

FIGURE 8. Simulation results.

F A N S


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optimal point to stage from two fans toone is when the speed drops below about63 percent.

The optimal staging point for this fansystem would be very different if thesystem curve ran through 0 in. at 0 cfm,

which would be the case if zone-basedreset of the static-pressure setpoint were

working perfectly. Figure 10 shows theoptimal speed to stage from one fan totwo is 55 percent. To stage from two fansto one, it is about 35 percent.

Static-pressure-setpoint reset. Static-pressure-setpoint reset by box demand isa control sequence that resets the static-pressure setpoint of the fan VSD basedon the measured demand of the terminalunits. At Site 1, the control systemprovides a fixed setpoint of 1.5 in. wc,regardless of demand. When direct digital

control (DDC) is used to control VAV zones and the zone controller knowsthe damper position (e.g., whether ananalog output or a floating point output

with position feedback is used for dampercontrol), this pressure setpoint couldbe reset from 0 in. wc to 1.5 in. wc asrequired to maintain the near-wide-openposition of the most open VAV damper.12

With perfect implementation of reset,this would provide a system curve that

went from the design condition thoughthe point of 0 in. wc at 0 cfm (i.e., if nozone required air, static pressure would

not be required). Real implementationsof terminal reset fall somewhere inbetween these two extremes. Figure 11shows the relationship between theminimum duct static-pressure setpointand the optimal point for staging fans.The Y-axis represents the optimal fan-staging points (up and down), while the

X-axis represents the minimum pointon the control-system curve.

The impact of static-pressure-setpointreset on both annual energy use and fanselection was evaluated. To simulate reset,a new load profile was developed byreplacing the monitored pressure withthe pressure from the system curve in

Figure 5 (perfect reset line) for the moni-tored airflow. These reset data were usedto compare the performance of the fansevaluated in Figure 8. The results arepresented in Figure 12, which shows thatannual fan energy use can be cut by asmuch as 50 percent if static-pressure resetis implemented successfully (compareFigure 12 with Figure 8). This corrobo-rates the results reported by Hartmand

and others.Figure 12 also shows that annual

energy ranking now follows the efficiency ranking shown in Figure 4. Thisis because a fan operating on a perfect sys-

tem-reset curve has constant efficiency.This is one of the reasons why static-pres-sure reset saves so much energy—notonly is the fan doing less work (maintain-ing lower static pressures), it is doing it ata higher efficiency.13

Another advantage of supply-pressure-setpoint reset is that it dramaticallyreduces the time that fans operate insurge. As can be seen in Figure 2, a systemcurve that tails off horizontally (like theactual system curve in Figure 5) will endup in the surge region more often thana reset curve (Figure 2) will. A perfectsystem curve with reset starting at a pointto the right of surge will never end up inthe surge region.

The results in Figure 12 imply that,in terms of energy cost, bigger fans are


F a n - s y s t e m

K W

140

120

100

80

60

40

20

0

Cubic feet per minute 1 0

, 0 0 0 0

2 0 , 0 0 0

3 0 , 0 0 0

4 0 , 0 0 0

5 0 , 0 0 0

6 0 , 0 0 0

7 0 , 0 0 0

8 0 , 0 0 0

9 0 , 0 0 0

1 0 0 , 0 0

0

1 1 0 , 0 0

0

1 2 0 , 0 0

0

1 3 0 , 0 0

0

1 4 0 , 0 0

0

S p e e d ,

p e r c e n t

100

90

80

70

60

50

40

30

20

10

0

1.5-in. curve—one fan (KW)1.5-in. curve—two fans (KW)

1.5-in. curve—one fan (speed)

1.5-in. curve—two fans (speed)

FIGURE 9. Optimal staging without static-pressure reset.

F a n - s y s t e m

K W

140

120100

80

60

40

20

0

Cubic feet per minute

1 0 , 0 0 0 0

2 0 , 0 0 0

3 0 , 0 0 0

4 0 , 0 0 0

5 0 , 0 0 0

6 0 , 0 0 0

7 0 , 0 0 0

8 0 , 0 0 0

9 0 , 0 0 0

1 0 0 , 0 0

0

1 1 0 , 0 0

0

1 2 0 , 0 0

0

1 3 0 , 0 0

0

1 4 0 , 0 0

0

S p e e d ,

p e r c e n t

100

90

8070

60

50

40

30

20

10

0

0-in. curve—two fans (speed)

0-in. curve—one fan (KW)

0-in. curve—two fans (KW)

0-in. curve—one fan (speed)

FIGURE 10. Optimal staging with static-pressure reset.

F A N S


8/1039HPAC Engineering • May 2003

better for systems with supply-pressure-setpoint reset. Indeed, the approximately $560 in annual energy savings fromselecting the 73-in. plenum fan ratherthan the 66-in. plenum fan pays forthe $1,200 incremental cost increase(Figure 4) with a simple payback ofless than three years; however, theseresults need to be tempered with specialconsiderations. In addition to the firstcost of the fan, other first costs, including the impacts on space and the electricalservice, should be considered. Alsoconsidered should be the increasedrisk that the fan will operate in surgeshould perfect reset not occur. (The mostcommon cause of less-than-perfectreset is a zone that is undersized, haslower-than-design temperature setpoints,

or has consistently high loads, all of which can result in a consistently highdemand for static pressure, even whenthe rest of the system is at low load.)The bigger the fan, the closer the designpoint is to the surge region and thegreater the risk of operating in surge for a less-than-perfect reset curve.

CONCLUSIONS

It is clear that significant energy

savings can result from careful fanselection and control. It also is clear thatfurther work needs to be done to develop

design guidelines and tools for fanselection and system design. Currentsimulation and design tools do notadequately model the part-load perform-ance of the components of a fan system(i.e., fans, motors, drives, and belts). Theauthors’ research project is addressing some of these needs.

Clearly, static-pressure reset by zonedemand is a key to fan-system perform-ance. Not only does it greatly reduce

energy use, it can eliminate the noise,vibration, and instability issues associated

with fans operating in surge. Becausemany DDC systems do not providefeedback on box-damper position,guidelines for applying this reset acrossa range of control-system capabilitiesare required.

Traditional fan-selection techniquesyield reasonably good results. Betterresults are possible with the use of asimulation tool to estimate annual fan-load profile.

Although, in general, larger fans aremore efficient, bigger is not necessarily

better. The design-point-efficiency rank-ings in manufacturers’ software cannotbe extrapolated to annual-energy-costrankings, unless static pressure is resetaggressively.

Optimum fan staging can be deter-mined during design by simulating theproposed fan system against a handful of data points on the expected system curve.The expected system curve and, thus, theoptimum staging point are greatlyaffected by the success of static-pressurereset. Of course, fan staging can be revis-ited after occupancy, based on the actualsuccess of static-pressure reset.

F a n s p e e d ,

p e r c e n t

80

75

70

65

60

55

50

45

40

35

30

Duct static-pressure setpoint

0.0 0.5 1.0 1.5

From one fan to two

From two fans to one

FIGURE 11. Optimal fan-staging point vs. minimum duct static-pressure setpoint.


Base case (66 PL-A, no reset)










$0 $5,000 $10,000 $15,000 $20,000 $25,000

$8,519

$20,474

$10,314

$16,611

$15,562

$10,806

$9,753

$9,720

$9,175

$8,917

$8,685

Annual electricity cost

FIGURE 12. Simulation results with perfect static-pressure reset.

F A N S


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Circle 154

ACKNOWLEDGEMENTS

The authors wish to acknowledge theinput and work of fellow research-teammembers Erik Kolderup and TianzhenHong of Eley Associates, Lynn Qual-mann of SBW Consulting Inc., andRoger Lippman of New Horizon Tech-

nologies. They also wish to recognize

the contributions of members of theirtechnical advisory team. Lastly, theyextend special thanks to buildingengineers Ravish Puri and ChristopherNewbury for putting up with theauthors’ intrusions on their building andfor their significant assistance with theauthors’ work.

FOOTNOTES

1) Data from www.calmac.org .2) Information on the PIER program

is available at www.energy.ca.gov/pier .3) Information on the Large HVAC

Integration project is available at www .newbuildings.org/pier/index.html .

4) The five sites monitored as part of

this research project represent a range of occupancy types, building sizes, andclimates. Criteria for site selectionare presented in a paper by Kolderup,Hydeman, Baker, and Qualmann fromthe 2002 ACEEE Conference on Energy Efficiency.a

5) Note that the pressure measured

was not exactly fan static pressure aspresented in the manufacturer’s fan curve(Figure 2) because it included thebackdraft damper. Fan ratings aredescribed in ANSI/ASHRAE Standard51-1999/AMCA Standard 210-1999,Laboratory Methods for Testing Fans for Aerodynamic Performance Rating .

6) A system curve is a theoretical linethat represents friction losses in a ductsystem as a function of airflow. It istypical to use a quadratic equation (i.e.,pressure drops as the square of the flow);however, in practice, this coefficientis closer to 1.8. The curve in Figure 5

goes to 1.5 in. wc at 0 cfm, as this is thecontrol point maintained at the static-pressure sensor. With perfect demand-based pressure reset, the curve will gothrough 0 in. wc at 0 cfm.

7) Fan surge is a condition of instability that occurs under high-static-pressureand low-flow conditions. Air is pushed


It is clear that significant energy savings can

result from careful fan selection and control.

F A N S


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Circle 189

out by the blades and subsequently pulled back by the high pressure differ-ence between the fan inlet and outlet.Instead of flowing smoothly through thefan, the air pulses in and out of the fanblades. This unstable operation reducescontrollability, increases wear, and cancause severe acoustical and vibrationproblems.

8) The relative cost and efficiency of fan types can be seen using manufactur-ers’ selection programs.

9) This low-frequency noise typically can be mitigated through the reductionof system static pressure and the applica-

tion of lined ductwork.10) A number of issues associated with

isolating fans in parallel are beyond thescope of this article. These will be coveredin detail in the project’s analysis reportand design guidelines.

11) Fan-system efficiency includes thefan, motor, drive (belts), and VSD. For a

given operating point, the efficiency of the motor, drive, and VSD increases asthe load increases. Unless fan efficiency improves appreciably to offset theselosses, operating one fan generally isbetter than operating two. Two fansrunning in parallel often operate in surge,

with a fixed pressure setpoint, unlike a single fan.

12) When box-damper position is notknown, a “trim-and-respond” algorithmcan be employed for a similar effect.

A promising non-traditional techniqueis to directly control fan speed by zonedemand without the use of PID loops.b,c

This technique reduces fan hunting atlow loads.

13) This is true only when fan speedis controlled with a VSD. If a fan iscontrolled using inlet guide vanes or ridesits curve, the savings will be significantly less, and the fan may enter the surgeregion.

REFERENCES

a) Kolderup, E., Hydeman, M., Baker,M., & Qualmann, R.L. (2002, August). Measured performance and design guide-lines for large commercial HVAC systems .

ACEEE Conference on Energy Effi-ciency.

b) Hartman, T. (2001, December).Ultra-efficient cooling with demand-based control. HPAC Engineering , pp.29-32, 34, 35.

c) Hartman, T. (1995, June). Global optimization strategies for high-perform-ance controls . Paper presented at ASHRAE

Annual Meeting, San Diego, CA.

d) Hartman, T. (1993, January). Ter-minal regulated air volume (TRAV)systems . Paper presented at ASHRAE

Winter Meeting, Chicago, IL.

For HPAC Engineering feature articles dating back to January 1992, visit

www.hpac.com.

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hpac a fresh look at fans-final

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