hpac primary secondary loop vs. primary loop only systems

SBy ALEXANDER L. BURD, PHD, PE, and GALINA S. BURD, MS

Advanced Research Technology LLC

Sufeld, Conn.

Subdividing various systems into a primary/secondary

(P/S) loop via a hydraulically dependant interconnection

long has been a standard solution for central chilled-water

plants in the United States and Europe. This achieves, at

a relatively low cost, reasonably good hydraulic separa-

tion of central-water-plant cooling-generation systems

(primary loop) from distribution piping

and terminal units (secondary loop).

Primary-loop flow is relatively con-

stant, while secondary-loop flow varies

based on load demand. Primary- and

secondary-loop water flows are inter-

changeable.

In a primary loop, control is achieved

by maintaining a relatively constant flow

rate; water temperature may be changed

via a reset control. This commonly is

referred to as a qualitative control strat-

egy. In a secondary loop, control typi-

cally is achieved by varying water-flow

rate. This commonly is referred to as a

quantitative control strategy.

Figure 1 depicts a system with a con-

stant- or variable-flow primary loop

and a variable-flow secondary loop.

The dedicated constant-speed Pump 1

maintains practically constant flow in

the primary loop (the pump does not

have variable-frequency-drive [VFD]

control), even if flow in the secondary loop (which has its

own pump, variable-speed Pump 2) varies significantly.1,2

Pump 1 is sized to maintain water flow between the chiller

evaporators minimum and maximum allowable values.

Typical Control StrategyIn the system in Figure 1, the direction of water flow

in the decoupling pipe is not controllable and may vary,

depending on the ratio of flow in the secondary and

primary loops.

Various modes of operation of the system were investi-

gated.3,4 The major parameters in the evaluation of energy

efficiency were supply- and return-water temperature and

flow rate before and after the decoupling pipe separating

the primary and secondary loops.

36 HPAC ENGINEERING DECEMBER 2010

Primary/Secondary-Loop vs.

Primary-Loop-Only Systems

Alexander L. Burd, PhD, PE, is president of and Galina S. Burd, MS, is a project manager for Advanced Research Technology, an

engineering and research consulting firm with offices in Suffield, Conn., and Green Bay, Wis. Alexander ([email protected])

has 35 years of experience in the design, research, and optimization of HVAC and district energy systems, which includes

publication of more than 35 research and technical papers in American and European journals, while Galina (gburd@energyart

.net) has more than 25 years of design and research experience in the HVAC and architectural-engineering fields. She has

co-authored many technical and research papers published in American journals.

Comparison of operational modes

and performance of two schemes for

optimizing chilled-water plants

Load

P2

1

2

B

A

3F2

F1t2

t1

t4

t3

P1

VFD

Chiller

1 = Optional constant or variable speed with variable-frequency-drive (VFD) primary-loop pump 2 = Variable-speed secondary-loop pump with VFD 3 = Non-controllable bidirectional decoupling pipe (AB or BA direction) between primary and secondary loop F1 = Flow-meter primary-loop flow rate F2 = Flow-meter secondary-loop flow rate t1 = Temperature of water leaving chiller t2 = Temperature of water returning to chiller t3 = Secondary-loop supply-water temperature t4 = Secondary-loop return-water temperatureP1 = Pressure differential for controlling F1 in Pump 1P2 = Pressure differential for controlling F2 in Pump 2

VFD

Secondary loop (distribution-system piping)Primary loop

(generation-system piping)

FIGURE 1. Optimized control strategy for chilled-water plant with primary (constant- or

variable-flow)/secondary (variable-flow) loop.

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Mode 1. When water flow in the secondary loop (F2)

exceeds water flow in the primary loop (F1) because of a

load increase, a portion of the water returning from the

secondary loop recirculates into the supply-distribution

system (A-to-B direction) and mixes with the flow in the

primary loop. This mode of operation is represented

by the following equations: F2 > F1, t3 > t1, t2 = t4, and

(t4 t3) < (t2 t1).

Mode 2. When water flow in the secondary loop (F2)

is less than water flow in the primary loop (F1) because

of a load reduction, flow in the decoupling pipe reverses

(B-to-A direction). Thus, the excessive flow exiting the

cooling-generation system returns to the primary-loop

system and the chiller. This mode of operation is repre-

sented by the following equations: F1 > F2, t1 = t3, t2 < t4,

and (t4 t3) > (t2 t1).

Mode 3: When the flow in the primary loop (F1) equals

the flow in the secondary loop (F2), there is no flow in the

decoupling line. All water from the secondary loop returns

to the primary loop and chiller, while all water exiting the

chiller flows through the secondary loop. This mode of op-

eration is represented by the following equations: F1 = F2,

t1 = t3, t2 = t4, and (t2 t1) = (t4 t3). Obviously, this mode of

operation is the most beneficial from an energy perspective.

Optimized Control StrategyFigure 1 depicts the optimized control strategy (Mode

3). Unlike a system with an optional constant-speed

primary-loop pump, the system has an additional VFD

controlling the speed of the primary-loop pump. The

rate of water flow in the secondary loop via Pump 2 is

dependent on system load. Water-flow rate in the primary

loop is a function of water-flow rate in the secondary loop

and adjusted to maintain equalized flow.

Following are simplified thermal-balance equations ap-

plicable for both primary and secondary loops, assuming

the specific heat of water does not change appreciably:

QPR = QSEC (1)

QPR = F1 (t2 t1) = F1 (tPR) (2)

QSEC = F2 (t4 t3) = F2 (tSEC) (3)

F1 tPR = F2 tSEC (4)

F1 = F2 (5)

tPR = tSEC (6)

where:

QPR = Primary-loop cooling load, British thermal units

per hour

QSEC = Secondary-loop cooling load, British thermal

DECEMBER 2010 HPAC ENGINEERING 37

Circle 166

units per hour

tPR = Primary-loop temperature

differential, degrees Fahrenheit

tSEC = Secondary-loop tempera-

ture differential, degrees Fahrenheit

Equations 5 and 6 essentially rep-

resent the algorithm for the control

of chilled-water plants. For the sys-

tem in Figure 1, control is accom-

plished by varying the speed of

Pump 1. The pumps speed should

not be reduced to the extent water-

flow rate falls below the allowable

low limit or increased to the extent

water-flow rate exceeds the allow-

able high limit.

The building-automation system

would have to limit VFD-turndown

and turn-up ratios to stay within

the range of allowable current fre-

quencies correlated to the range of

allowable primary-loop water-flow

rates. To better match primary- and

secondary-loop water-flow rates,

two-phase control is suggested. The

first phase could consist of quanti-

tative control in both the primary

and secondary loops (Equation

5) while the chilled-water-supply

temperature remained at a given

constant value. The second phase

could consist of qualitative control

in the primary loop and quantita-

tive control in the secondary loop.

(The order in which the control ac-

tions are implemented may vary.)

Reset water-temperature control

(Equation 6) could be realized by

varying water temperature (t1) at

a given fixed (limited) magnitude

of water-flow rate in the primary

loop. The change in water tempera-

ture would impact flow rate in the

secondary loop indirectly; flow via

the primary-loop pump could not

be changed further because of the

aforementioned evaporator-flow

limitations. The P/S-loop system

with variable flow and temperature

control in both loops in Figure 1 is

very versatile, allowing the estab-

lishment of flow-limiting parame-

ters and temperature set points (t1)

over a given time period.

Primary-Loop-Only-Variable-Flow Control System

A primary-loop-only-variable-flow

(PLOVF) control system employs

a single pump to circulate water

through generation- and distribu-

tion-system piping loops (Figure 2).

This arrangement allows uniformly

distributed variable flow throughout

entire systems. The generation- and

distribution-piping systems have a

dependant flow-control arrange-

ment, unlike the generation- and

distribution-piping systems in the

P/S-loop system in Figure 1, which

have an independent flow-control

arrangement with two dedicated

pumps. When flow in a PLOVF dis-

tribution system varies because of

a load change, a VFD control varies

the speed of the pump. If flow in the

distribution system falls below the

chillers low limit, an automatic con-

trol valve modulates to divert flow

from the supply line of the distribu-

tion system back to the chiller (A-to-B

direction). This system, unlike the

P/S-loop system in Figure 1, does not

have the ability to bypass the chiller

if flow in the distribution system

exceeds the chillers high-limit flow.

For this capability to be provided in a

PLOVF system, a second controllable

decoupling pipeline with automatic

control valve and reversed flow di-

rection would have to be added on

the other side of the pump. Other-

wise, another chiller, as well as an-

cillary equipment (i.e., chilled-water

pump, cooling tower, condenser

pump), would have to be placed on-

line, which would result in increased

chiller-plant power demand as soon

as flow in the distribution system

exceeded the high-level limit.

A PLOVF system essentially is

capable of maintaining limited total

flow rate. Typically, control of systems

with flow-rate-limiting devices5 is

achieved by assigning priority status

to one of the two loads. For instance,

in a district-heating-system customer

substation with space-heating and

domestic-hot-water loads, priority is

given to the domestic-hot-water load.

The space-heating load has substan-

tial thermal inertia and the ability to

temporarily accept a lower water-

flow rate without noticeable impact

on air temperature. In a PLOVF sys-

tem, two systems that typically have

low thermal inertia (space cooling) or

no inertia at all (chiller-evaporator-

water flow) share the same level of

priority control.

Control of a PLOVF system is

somewhat more challenging than

control of a P/S-loop system. This


PRIMARY/SECONDARY-LOOP VS. PRIMARY-LOOP-ONLY SYSTEMS

1 = Primary-loop-only variable-speed pump with variable-frequency drive (VFD) 2 = Automatic control valve for converting primary-loop-only system to constant primary-loop and variable secondary-loop flow operation (single AB-direction controllable flow decoupling pipe) F1 = Flow-meter generation-system-piping-loop flow rate F2 = Flow-meter distribution-system-piping-loop flow rate t1 = Temperature of water leaving chiller t2 = Temperature of water returning to chiller t3 = Distribution-system-piping-loop supply-water temperature t4 = Distribution-system-piping-loop return-water temperatureP1 = Pressure differential for controlling F1P2 = Pressure differential for controlling F2

Distribution-system piping loop

Generation-system piping loopLoad

P2

1

2

F2

F1t2

t1

t4

t3

P1

VFD

Chiller

B

A

FIGURE 2. Primary-loop-only-variable-flow control system.

is because a PLOVF system has to

maintain two variables that are con-

tinually fighting: a given set point

of pressure differential (P2) in the

distribution-system loop and a given

set point of pressure differential (P1)

in the generation-system-piping

loop. Thus, overall control-strategy

execution and system operation may

be less accurate and stable, and an

allowance for the possible deviation

from the control-parameter set point

may have to be made. Typically, to

avoid a chiller shutting down because

of insufficient flow, the set point for

low-limit flow via the evaporator is

increased. Because any problems

with the modulating control valve

could impair system performance or

even shut down the entire chilled-

water system, the application of reset

chilled-water-temperature control

during off-design conditions (when

the chilled-water-flow limitation is in

place) is difficult. In comparison with

a P/S-loop system, design chilled-

water temperature may be constant

and elevated in a PLOVF system,

which may lead to increased costs

for distribution piping, pumping,

cooling coils, etc.

Chilled-Water-System Temperature Differential

Chilled-water temperature differ-

ential is critical to P/S and PLOVF

operations in that it determines

chilled-water flow per cooling ton.

Impacting chilled-water temperature

differential most significantly are the

end-users connected to chiller plants.

Take, for instance, an air-handling

unit (AHU) with the following param-

eters:

Air is cooled in a counterflow

chilled-water coil.

The design cooling load is 28.5

tons.

The design sensible and latent

loads are 76 percent and 24 percent

of the design cooling load, respec-

tively.

The cooling load varies in direct

proportion to outdoor dry-bulb air

temperature.

The cooling coil has a two-way

chilled-water control valve to vary

water flow through the coil to satisfy

loads.

The cooling coil is selected for

15F design chilled-water tempera-

ture differential.

The design chilled-water flow via

the cooling coil at 40F inlet water

temperature is 45.7 gpm.

End-user load control. The authors

considered two types of AHU cool-

ing-load-control systems: constant

airflow control (CAFC) and variable

airflow control (VAFC). Maximum

airflow turndown ratio was assumed

to be 3.4 at a relative cooling load

of 0.29 and lower. The top graph

in Figure 3 shows the CAFC sys-

tem requires a substantially greater

change in relative chilled-water flow

to achieve the same level of variation

in cooling-coil load. For instance, for

cooling-coil capacity to be reduced

to 40 percent of the design load, rela-

tive chilled-water flow would have

to be reduced to about 21 percent of

its design magnitude. For the same

reduction in cooling-coil capacity to

be achieved with the VAFC system,

relative chilled-water flow rate would

have to be reduced to about 35 per-

cent of its design magnitude. The bot-

tom graph in Figure 3 indicates the

CAFC system will increase the rela-

tive chilled-water temperature differ-

ential by a factor of about 1.9 when

the relative cooling load is reduced

to 40 percent of its design magnitude.

For the same conditions, the VAFC

system will increase relative chilled-

water temperature differential by a

factor of only about 1.2. Thus, the

same relative reduction in chilled-

water flow via the cooling coil would

result in about 1.6-times-higher rela-

tive chilled-water temperature differ-

ential with the CAFC system than it

would with the VAFC system.

Chiller-evaporator low-limit flow

control. Various control strategies

can be utilized to ensure the sustain-



Chill

ed-w

ater

-coil

rela

tive

ch

illed

-wat

er t

emper

ature

diffe

rential

Cooling-coil relative chilled-water flow rate

Chill

ed-w

ater

-coil

rela

tive

coolin

g lo

ad

1.0

0.8

0.6

0.4

0.2

0.2 0.4 0.6 0.8 1.0

Cooling-coil relative chilled-water flow rate

0.2 0.4 0.6 0.8 1.0

2.0

1.8

1.6

1.4

1.2

1.0

CAFC

CAFC

VAFC

VAFC

FIGURE 3. Air-handling-unit-cooling-coil operational parameters with variable (VAFC) and

constant (CAFC) airflow control.

ability of a systems low-limit chilled-

water flow. Figure 4 shows the reset

water-temperature control required

to maintain a given low-limit water-

flow level when relative cooling load

changes from 0.64 to 0.17 of design.

The temperature of water entering

the cooling coil increases as load

decreases. Limited chilled-water

flow is maintained by increasing the

temperature of water entering the

cooling coil.

The variation in relative cooling

load from 0.67 to 0.17 requires the

temperature of water entering the

coil to be increased from 40F to

about 52F. The bottom graph in

Figure 4 indicates chilled-water tem-

perature differential will be reduced

from 15F at the relative cooling load

of 0.67 to 3.7F at the relative cooling

load of 0.17.

Other factors impacting chilled-

water-system temperature differ-

ential. Chilled-water temperature

differential also can be impacted by

deposits on the inside and outside

surfaces of cooling coils. Proper and

timely cleaning of heat-exchanger

heat-transfer surfaces, thus, is

important,6 as is cleaning of chilled-

water cooling coils and chiller evapo-

rators and condensers.

Because of the complexity of

real-life conditions, it is unrealistic

to expect an increase in cooling-coil

chilled-water temperature differen-

tial coinciding with a reduction in

cooling load. If a system is balanced

and well-maintained, temperature

differential will remain relatively

close to its design value during off-

design conditions.

Distribution-piping-system design

temperature differential. Distribution-

piping-system design temperature

differential impacts the installed and

operating costs of central chilled-

water systems. For retrofit projects,

such as conversion from direct-

expansion cooling coils to central

chilled-water cooling coils, the cost of

distribution piping could run as high

as 30 percent of the entire system.3,4

Design temperature differential for

the example chilled-water primary

piping system was assumed to vary

from about 2.1F to 22.4F (Table

1). The magnitude of the tempera-

ture differential at the chillers

design cooling load was limited by

t h e a l l o w a b l e m i n i m u m a n d

maximum flow at a given number of

evaporator passes. Low-level flow

rate was specified to ensure evapo-

rator operation with sufficient heat-

exchanger-tube water velocity. High-

level flow rate was specified to ensure

evaporator operation with allowable

heat-exchanger-tube velocity and

avoid unstable heat transfer and tube



Chill

ed-w

ater

-coil

rela

tive

coolin

g lo

adCooling-coil inlet chilled-water temperature, degrees Fahrenheit

Coolin

g-c

oil

chill

ed-w

ater

tem

per

ature

diffe

rential

, deg

rees

Fah

renhei

t

0.0

16

40 41 42 43 44 45 46 47 48 49 50 51 52 53

Cooling-coil inlet chilled-water temperature, degrees Fahrenheit

40 41 42 43 44 45 46 47 48 49 50 51 52 53

14

12

10

8

6

4

2

0

0.2

0.4

0.6

0.8

FIGURE 4. Cooling-coil inlet chilled-water temperature, relative cooling load, and

temperature differential with low-limit chilled-water-flow control.

Number of evaporator passes 1 2 3

Low-limit evaporator-water flow rate, gallons per minute 964 482 321

Maximum evaporator chilled-water temperature differential, degrees Fahrenheit

7.5 14.9 22.4

Minimum evaporator specific chilled-water flow rate per cooling ton, gallons per minute per ton

3.2 1.6 1.1

High-limit evaporator flow rate, gallons per minute 3,473 1,737 1,158

Minimum evaporator chilled-water temperature differential, degrees Fahrenheit

2.1 4.1 6.2

Maximum evaporator specific chilled-water flow rate per cooling ton, gallons per minute per ton

11.6 5.8 3.9

Relative evaporator-water-flow-rate variation 3.6 3.6 3.6

Notes:1. Relative evaporator-water flow rate shows the ratio of maximum to minimum specific

chilled-water flow rate at the given fixed number of evaporator passes.2. Minimum allowable water-flow rate (gallons per minute) via evaporator at a water velocity of

3.3 fps at a given number of passes could be calculated as minimum specific flow rate (gallons per minute per ton) multiplied by design cooling load (tons).

3. Maximum allowable water-flow rate (gallons per minute) via evaporator at water velocity of 12 fps at a given number of passes could be calculated as maximum specific flow rate (gallons per minute per ton) multiplied by design cooling load (tons).

4. Parameters based on the performance of a 300-ton centrifugal chiller.

TABLE 1. Chiller-evaporator design operational parameters.

erosion. Lower water-flow rates

relate to higher chilled-water tem-

perature differentials and vice versa.

Table 1 indicates temperature

differential is dependent on num-

ber of evaporator passes. It shows

evaporator relative water-flow rate

remains the same (i.e., 3.6the ratio

of maximum to minimum tube water

velocities) for all considered number

of evaporator passes.

Allowable minimum chiller rela-

tive cooling load (MCRCL) in Table

2 was calculated using the following

equation:

MCRCL = [1 (TDEMAX TDEMIN)]

SOSF 100, %

where:

TDEMAX = maximum chiller-evapo-

rator design temperature differential

TDEMIN = minimum chiller-evapo-

rator design temperature differential

SOSF = system operational safety

factor, which increases minimum

chilled-water flow via a chillers

evaporator (for PLOVF) to prevent

chiller shutdown on low flow

The higher water-flow turndown

ratios (WFTDRs) in Table 2 are

advantageous because they allow

optimal operation of a system to

satisfy cooling loads.

Chilled-Water Turndown Ratio and Energy Savings

Figure 5 shows WFTDR patterns

for various control strategies. The

system was assumed to use mechani-

cal free cooling at outdoor dry-bulb

air temperatures of 55F and below.

Also, system cooling load was as-

sumed to change linearly in direct

proportion to outdoor dry-bulb air

temperature.

The following options, shown in

Table 2, were considered:

Option 1

TDEMIN = 15F, TDEMAX = 22.4F,

WFTDR = 1.5

This means re la t ive chi l ler-

evaporator chilled-water flow could

be reduced to 67 percent of its

design value.

Relative chilled-water flow via the

pump serving the PLOVF system

in Figure 2 will follow two straight

lines outlined by the triangle ACB in

Figure 5 and remain constant at the

level designated by straight line CB.

Although relative water flow via the

distribution-piping system will fol-

low the load change, the bypass valve

will be open to ensure the required

low-limit chiller-evaporator flow is

maintained. This means the pump

will run at the constant flow rate

indicated by the straight line CB.

P/S-loop systems with variable-

primary and secondary-loop pump-

ing could provide energy savings

(compared with the PLOVF-system

operation indicated by straight line

CB in Figure 5) while dry-bulb out-

door-air temperature ranged from



primary/secondary and primary-loop-only systems.

Maximum chiller-evapora-tor design temperature differential, TDEMAX, degrees Fahrenheit

Minimum chiller-evaporator design temperature differential, TDEMIN, degrees Fahrenheit

Allowable reduction in water-flow turn-down ratio (TDEMAXto TDEMIN), times

Allowable MCRCL, percent

22.4 15.0 1.5 73.6

22.4 10.0 2.2 49.0

22.4 6.2 3.6 30.5

Notes:1. Allowable reduction in chilled-water flow rate in distribution piping system the ratio of TDEMAX

to TDEMIN.2. Data adapted from Table 1 (chiller with three-pass evaporator).3. Minimum allowable flow rate via chiller evaporator increased by 10 percent to ensure chiller

in PLOVF system will not be turned down on low-level evaporator flow rate.4. MCRCL = Minimum chiller relative cooling load, at which Bypass Control Valve 2 in Figure 2

will remain closed.

F

C

D

B

H

E

G

A

AC = Secondary-loop relative flow-rate value in systems with constant- and variable-flow primary-loop pumps (T

DEMIN = 15F, T

DEMAX = 22.4F)

CB = Minimum primary-loop relative flow-rate value in systems with constant- and variable-flow primary-loop pumps (T

DEMIN = 15F, T

DEMAX = 22.4F)

ACB = PLOVF-system relative flow-rate variation (TDEMIN

= 15F, TDEMAX

= 22.4F)AD = Secondary-loop relative flow-rate value in systems with constant- and variable-flow primary-loop pumps (T

DEMIN = 10F, T

DEMAX = 22.4F)

DH = Minimum primary-loop relative flow-rate value in systems with constant- and variable-flow primary-loop pump (T

DEMIN = 10F, T

DEMAX = 22.4F)

ADH = PLOVF-system relative flow-rate variation (TDEMIN

= 10F, TDEMAX

= 22.4F)AF = Secondary-loop relative flow-rate value in systems with constant- and variable-flow primary-loop pumps (T

DEMIN = 6.2F, T

DEMAX = 22.4F)

FG = Minimum primary-loop relative flow-rate value in systems with constant- and variable-flow primary-loop pumps (T

DEMIN = 6.2F, T

DEMAX = 22.4F)

AFG = PLOVF-system relative flow-rate variation (TDEMIN

= 6.2F, TDEMAX

= 22.4F)AE = Relative flow-rate variations for secondary-loop pumps for all considered P/S-loop systems and T

DEMIN and T

DEMAX values

Pri

mar

y/se

condar

y an

d p

rim

ary-

loop-o

nly

re

lative

flow

rat

e 0.75

1.00

0.50

0.25

0.0055 65 10575 85 95

Outdoor dry-bulb air temperature, degrees Fahrenheit

FIGURE 5. Primary/secondary and primary-loop-only relative chilled-water-flow variation.

TABLE 2. Considered design temperature differentials for chiller evaporators serving

55F to 83F because of the variable

secondary-loop pumping (Pump 2 in

Figure 1) depicted by straight line CE

in Figure 5.

Option 2

TDEMIN = 10F, TDEMAX = 22.4F,

WFTDR = 2.2

This means relative chiller-evap-

orator chilled-water flow could be

reduced to 44.6 percent of its design

value.




lines outlined by the triangle ADH in


level designated by straight line DH.

Although water flow via the distri-

bution-piping system will follow the

load change, the bypass valve (2) will

be open to ensure the required low-

limit chiller-evaporator flow is main-

tained. This means the pump will run

at the constant flow rate indicated by

straight line DH.






DH in Figure 5) while dry-bulb out-




Figure 1) depicted by straight line DE

in Figure 5.

Option 3

TDEMIN = 6.2F, TDEMAX = 22.4F,

WFTDR = 3.6

This means relative chiller-evap-

orator chilled-water flow could be

reduced to 27.7 percent of its design

value.




lines outlined by the triangle AFG in


level designated by straight line FG.

Although water flow via the distri-

bution-piping system will follow the

load change, the bypass valve (2) will

be open to ensure the required low-

limit chiller-evaporator flow is main-

tained. This means the pump will run

at the constant flow rate indicated by

straight line FG.






DH in Figure 5) while dry-bulb out-




Figure 1) depicted by straight line FE

in Figure 5.

Cooling-Load ProfileCooling-load profile depends on

the ratio of constant load (e.g., inter-

nal heat gain from lights, equipment,

people, etc.) to variable load (e.g.,

ventilation, heat gain from building

envelope, etc.).

Cumulative relative cooling load

and time-duration factor for eight

constant- and variable-load compo-

nents in New England are given in

Figure 6. The design cooling load for

the process area of a manufacturing

facility is close to 57 tons. The actual

constant-cooling-load component

was near 35 tons, or 62 percent of the

design load. The actual variable-load

component was close to 22 tons, or

38 percent of the design load.

Applied in energy analysis, Figure

6 is illustrative of relative cooling-

load variations when the percentage

of constant-cooling-load component

(PCCLC) changes from 0 percent to

100 percent of total cooling load.

P/S- and PLOVF-System Electrical-Energy Usage

Chilled-water-pumping electrical-

energy savings. The top graph in

Figure 7 shows potential annual

chilled-water-pumping electrical-

energy savings for a P/S-loop sys-

tem with variable-flow-rate control

in both loops. The electrical energy

consumed by pump motors was

assumed to vary in direct propor-

tion to changes in motor speed by a

power of 2.5. This is lower than the

theoretical value of 3 recommended

for centrifugal pumps. These savings

will be realized when distribution-

system water flow is equal to or lower

than the allowable chiller-evaporator

low limit (Figure 5). In calculations

of annual electrical-energy use, the

pump serving the PLOVF system was

assumed to have a design flow rate



Systems with 62-percent PCCLC








Notes:1. Constant load percentage based on ratio of constant-cooling-load component that does not change during mechanical cooling season to design cooling load.2. Relative cooling load shows ratio of current load to design cooling load.3. PCCLC = Percentage of constant-cooling-load component.

Cum

ula

tive

load

tim

e-dura

tion f

acto

r, p

erce

nt

(at

load

equal

to o

r lo

wer

than

show

n)

Relative cooling load, percent

100

90

80

70

60

50

40

30

20

10

10 20 30 40 50 60 70 80 90 1000

0

FIGURE 6. System relative cooling load and cumulative time-duration factor.

equal to the design flow rate of the

P/S-loop pumps. The pump serving

the PLOVF system also was assumed

to have a design pressure head equal

to the cumulative design pressure

head of the P/S-loop pumps.

With a MCRCL of 73.6 percent

(Table 2), P/S-loop-system annual

electrical-energy savings vary from

5.8 percent (calculated PCCLC of 62

percent) to 60 percent (calculated PC-

CLC of 0 percent). Annual pumping

electrical-energy usage for the con-

sidered P/S and PLOVF systems will

be equalized at a PCCLC greater than

or equal to 75 percent (Figure 7).

With a MCRCL of 49 percent

(Table 2), P/S-loop-system annual


10.4 percent (PCCLC of 26 percent)

to 41 percent (PCCLC of 0 percent).

Annual pumping electrical-energy

usage for the considered P/S and

PLOVF systems will be equalized

at a PCCLC greater than or equal to

35 percent (Figure 7).

With a MCRCL of 30.5 percent

(Table 2), P/S-loop system annual


2.8 percent (PCCLC of 15 percent)

to 9.5 percent (PCCLC of 0 percent).

Annual electrical-energy usage for

the considered P/S and PLOVF sys-

tems will be equalized at a PCCLC

greater than or equal to 21.5 percent

(Figure 7).

With an MCRCL of 73.6 percent

and a PCCLC varying from 0 percent

to 40 percent, the P/S-loop system

with constant-primary-loop and

variable-secondary-loop pumping

control consumes less electrical

energy annually than the PLOVF

system. These savings are identified

in Figure 7 as the difference in Y-axis

values to the left of the intersection of

the dashed lines. The savings could

be as high as 22 percent at a PCCLC

of 0 percent.

In calculations of electrical-energy

savings, primary- and secondary-

pump usage was assumed to be

about 6 percent of chiller-plant an-

nual electrical-energy consumption.

The data in the top graph of Figure

7 are related to single-chiller plants.

The use of multiple chillers will in-

crease WFTDR and reduce MCRCL.

This will reduce annual pumping

electrical-energy savings. Correc-

tion factors for two- and three-chiller

plants are given in Table 3.

Reset-chilled-water-temperature

electrical-energy savings. The P/S-

loop system with variable-flow



.

Linear (approximated MCRCL = 73.6 percent)Linear (approximated MCRCL = 49 percent)Linear (approximated MCRCL = 30.5 percent)Linear (approximated MCRCL = 73.6 percent)

Calculated MCRCL = 73.6 percentCalculated MCRCL = 49 percentCalculated MCRCL = 30.5 percentCalculated MCRCL = 73.6 percent

Two-chillers MCRCL = 36.7 percentSingle-chiller MCRCL = 30.5 percentTwo-chillers MCRCL = 22.9 percent

Single-chiller MCRCL = 73.6 percentTwo-chillers MCRCL = 55.2 percentSingle-chiller MCRCL = 49 percent

Notes:1. Savings calculated for chiller plant operating 24 hr a day, seven days a week in Hartford, Conn.2. MCRCL = minimum chiller relative cooling load (Table 2), at which Bypass Control Valve 2 in the primary-loop-only-variable-flow system (Figure 2) remains closed.3. Single-chiller operation considered in top graph only.4. Dashed red line in top graph indicates pumping savings over primary (constant-flow)/secondary (variable-flow) system.5. Dashed lines indicate savings over the primary-loop-only system.6. Annual pumping savings in top graph compared with chiller-plant annual pumping electrical energy only.7. In bottom graph, pumping and reset-chilled-water-temperature-control annual electrical-energy savings compared with annual electrical energy of entire chiller plant.

Percentage of constant-cooling-load component in relation to overall design cooling load, percent706050403020100

0

1

2

3

4

5

6

Percentage of constant-cooling-load component in relation to overall design cooling load, percent70605040302010

10

20

30

40

50

60

70

00

P/S

-loop-s

yste

m p

um

pin

g a

nd r

eset

ch

illed

-wat

er-t

emper

ature

-contr

ol a

nnual

el

ectr

ical

-ener

gy

savi

ngs,

per

cent

Pri

mar

y/se

condar

y-lo

op-s

yste

m p

um

pin

g

annual

ele

ctri

cal-

ener

gy

savi

ngs,

per

cent

FIGURE 7. Potential annual electrical-energy savings for primary/secondary-loop system with

variable flow.

multiple-chiller operations.

Single-chiller plant Two-chiller plant Three-chiller plant

1.00 0.75 0.50

Notes:1. Single-chiller-plant loading factor assumed to be 100 percent.2. Multiple-chiller-plant loading factor assumed to be 75 percent.3. Minimum allowable chiller relative cooling load increased by 10 percent.

TABLE 3. Correction factors for adjusting pumping annual energy savings to account for

control also has savings associated

with reset chilled-water-temperature

control. While considering these

savings, the authors assumed, based

on manufacturer data, that approxi-

mately 2 percent of the chillers input

energy would be saved per degree-

Fahrenheit increase in chilled-water

temperature. The increase in chilled-

water temperature also would in-

crease secondary-loop water flow,

which, in turn, would reduce the

electrical-energy savings associated

with the secondary-loop pump em-

ploying variable-flow control. The

authors incorporated that reduction

in their calculations. The chillers

power demand was assumed to be

about 78 percent of the total plant

power demand.

Cumulative annual electrical-

energy savings. Cumulative annual

electrical-energy savings for pump-

ing and reset chilled-water-temper-

ature control expressed as a per-

centage of total chiller-plant annual

energy consumption are shown in the

bottom graph of Figure 7. The annual

savings of a single-chiller plant with

an MCRCL of 30.5 percent and a

PCCLC of 62 percent and a single-

chiller plant with an MCRCL of 73.6

percent and a PCCLC of 0 percent

vary from 0.5 percent to 5.2 percent.

The annual savings of a two-chiller

plant with an MCRCL of 22.9 percent

and a PCCLC of 62 percent and a two-

chiller plant with an MCRCL of 55.2

percent and a PCCLC of 0 percent

vary from 0.4 percent to 4.1 percent.

Converting From PLOVF to P/S-Loop-With-VFD-Control System

A central chilled-water system

with 2,500-ton chillers serving mul-

tiple AHUs at a New England manu-

facturing facility was converted from

PLOVF (similar to that shown in Fig-

ure 2) to P/S-loop operation (nearly

identical to that shown in Figure 1).

The system had a design cooling load

of approximately 930 tons and a de-

sign secondary-loop chilled-water

flow of about 1,600 gpm. About 50

percent of the cooling load was con-

stant (PCCLC of 50 percent). The

P/S-loop system had two dedicated

water-flow meters to separately

measure the flows in the primary and

secondary loops, as well as in the

distribution- and generation-piping

loops of the PLOVF system. To com-

pare the hourly operational modes of

the systems, the authors selected a

two-day sample with approximately

the same dry-bulb outdoor-air tem-

perature for both systems. The reset

chilled-water-temperature control

for the P/S system was not activated

during the two days. The outdoor

dry-bulb air temperature ranged

from 57F to 83F. The average daily

dry-bulb air temperature was close to

69F, while the cooling load ranged

from 250 tons to 450 tons. Only one

chiller was running. The high-limit

chiller-evaporator flow was 1,200

gpm (i.e., TDEMIN was 10F). The low-

level chiller-evaporator flow was 670

gpm (i.e., TDEMAX was 17.9F). Thus,

WFTDR was 1.79.



Hourl

y el

ectr

ical

-ener

gy

usa

ge

by

pum

ps

in p

rim

ary/

seco

ndar

y an

d p

rim

ary-

loop-o

nly

sys

tem

s, k

ilow

att-

hours

Outdoor dry-bulb air temperature, degrees Fahrenheit55 60 65 7570 8580



Outdoor dry-bulb air temperature, degrees Fahrenheit55

50

40

30

20

10

0

900

800

700

600

500

40060 65 7570 8580

900

800

700

600

500

400

3

4

5

6

7

Pri

mar

y-lo

op-o

nly

-flow

-rat

e ch

illed

-w

ater

sys

tem

, gal

lons

per

min

ute

Primary-loop-only-system total water-flow rate (F1 in Figure 2)Primary-loop-only-system distribution-piping flow (F2 in Figure 2)

Pri

mar

y-lo

op-o

nly

-chill

ed-

wat

er-s

yste

m r

elat

ive

flow

ra

te v

ia b

ypas

s va

lve,

per

cent

Pri

mar

y/se

condar

y-ch

illed

-wat

er-s

yste

m-

loop fl

ow

rat

e, g

allo

ns

per

min

ute

Secondary-loop water-flow rate (F2 in Figure 1)Primary-loop water-flow rate (F1 in Figure 1)

Primary-loop-only-system pump hourly electrical-energy usagePrimary/secondary-pump cumulative hourly electrical-energy usage

FIGURE 8. Primary/secondary- and primary-loop-only-system hourly trend data.

Low-limit chiller-evaporator flow

was set at 700 gpm during P/S-loop

operation; during PLOVF-loop op-

eration, it was set at 770 gpm. An

attempt to lower low-limit flow

during PLOVF-loop operation was

unsuccessful because of evapora-

tor-induced chiller shutdown on

low flow.

Total chilled-water flow via the

pump in the PLOVF system (Point

F1 in Figure 2) was relatively sta-

ble, varying from about 720 to 860

gpm (Figure 8). Meanwhile, chilled-

water flow via the distribution-

piping system varied from 439 to

847 gpm. Chilled-water flow via the

chillers evaporator was maintained

at its allowable low-limit magnitude

through bypass-control-valve mod-

ulation. Chilled-water flow via the

bypass control valve, which repre-

sents the difference between flow

via the pump and flow via the distri-

bution-piping system, varied from

about 10 to 285 gpm. Flow via the

bypass valve, expressed as a per-

centage of total flow via the pump,

varied from about 2 percent (at

higher cooling loads) to 39 percent

(at lower cooling loads). Any time

and to any extentthe bypass valve

(2) was open, VFD speed and pump

flow remained constant, wasting

electrical energy.

Chilled-water flow in the P/S-loop

system varied from 700 to 816 gpm

(Figure 8). Flow in the secondary

loop varied from 420 to 820 gpm.

Chilled-water flow via the distribu-

tion-piping system varied indepen-

dently of water flow via the chiller

evaporator pump. Chilled-water

flow via the distribution system was

reduced following the load reduction

in the cooling system. This led to

reduced electrical-energy use by

Pump 2. Chilled-water flow via the

primary-loop pump and chil ler

evaporator remained relatively con-

stant, satisfying low-limit water-flow

requirements via the evaporator.

The same 100-hp pump (design

water flow 2,400 gpm) was used for

the PLOVF system and the secondary

loop of the P/S-loop system. The pri-

mary loop of the P/S-loop system had

its own 15-hp pump (design water

flow 1,200 gpm). The long operational

hours of the PLOVF system, with

nearly constant flow via the pump

necessary to maintain low-limit flow

via the chiller evaporator, resulted

in wasted electrical-energy power

demand (kilowatts) and usage (kilo-

watt-hours). Secondary-loop pump

electrical-energy usage decreased

and increased following cooling-load

variations. Electrical-energy savings

for the P/S-loop system varied from

about 3 percent to 50 percent, aver-

aging around 20 percent per day.

The authors have observed similar

electrical-energy savings following

other chiller plants conversion from

PLOVF to P/S-loop operation.

SummaryThis article compared the opera-

tional modes and performance of

P/S-loop and PLOVF systems.

The operating efficiency of chilled-

water systems with a primary (con-

stant-flow)/secondary (variable-

flow)-loop arrangement can be

improved by applying a two-phase

optimized control strategy consist-

ing of variable flow and temperature

control for primary-loop pumps and

variable-flow control for secondary-

loop pumps.

PLOVF-with-VFD control wastes

energy when primary-loop flow is

constrained by the chiller-evaporator

low limit and chilled water is diverted

via decoupling line from the supply

pipe to the return pipe of the gen-

eration system. The magnitude of

annual energy savings (pumping

and reset chilled-water temperature)

of P/S-loop control depends on

multiple variables, such as chiller-

evaporator-flow turndown ratio,

minimum chiller relative cooling load,

load composition, number of chill-

ers, etc. For New England weather

conditions, annual electrical-energy

savings could be 0.5 to 5.2 percent

for a single-chiller plant and 0.4 to 4.1

percent for a two-chiller plant.

P/S-loop-with-variable-flow con-

trol reduces chiller-plant power

demand by resetting chilled-water

temperature when flow in the dis-

tribution piping exceeds a high-limit

value for the chiller evaporator and

the system diverts water via decou-

pling line from the return pipe to the

supply pipe of the distribution sys-

tem. This prevents or delays the addi-

tion of a chiller and associated ancil-

lary equipment and saves energy.

Compared with PLOVF systems,

P/S-loop systems with variable flow

also have the advantage of more-

accurate, flexible, and stable control

and operational conditions.

References1) Burd, A.L. (1994). Optimizing

customers heating systems to reduce

capital and operating cost in district

heating with cogeneration. Proceed-

ings of International District Heat-

ing and Cooling Association Annual

Conference, pp. 237-252.

2) Burd, A.L. (1993). Computer

design of terminal heating substa-

tions for district heating. ASHRAE

Transactions, 99 (2), 245-265.

3) Burd, A.L., Burd, G.S., & De

Maio, M. (2005, March). Smith &

Wesson: The story of a chilled-water

retrofit (part 1). HPAC Engineering,

pp. 38-47.

4) Burd, A.L., Burd, G.S., & De

Maio, M. (2005, July) . Smith &

Wesson: The story of a chilled-water

retrofit (part 2). HPAC Engineering,

pp. 24-37.

5) Burd, A.L. (1997). Deferred

heat supply for space heating using

a capacity-limiting device A ben-

eficial approach for district heating.

ASHRAE Transactions, 103 (2), 23-31.

6) Burd, A.L., & Burd, G.S. (1996).

Optimal cycle for plate heat exchanger

cleaning in customer substation.

Proceedings of International District

Energy Association Annual Confer-

ence, pp. 177-192.

Did you find this article useful? Send

comments and suggestions to Scott

Arnold at [email protected].



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