MAIN PAGE
TABLE OF CONTENTS
4222
Analysis of Transfer Pumping Interfaces
for Stratified Chilled Water Thermal Storage
Systems—Part 1: Model Development
William P. Bahnfleth, Ph.D., P.E.
Christopher G. Kirchner, P.E.
Member ASHRAE
Associate Member ASHRAE
ABSTRACT
This paper describes methods for modeling the performance of transfer pumping interfaces used to connect open
chilled water thermal energy storage tanks to closed chilled
water systems. Indirect (heat exchanger) interfaces and direct
interfaces are considered. Treatment of direct interfaces
includes the use of energy recovery with centrifugal pumps
serving as hydraulic turbines. Of the seven interface types
considered, only three (constant-speed direct, constant-speed
direct with energy recovery, and constant-speed indirect) have
been previously discussed in the literature. Characteristics of
interfaces with variable-speed pumping and a novel reversible
direct interface with energy recovery are described quantitatively for the first time. Models of interface components are
derived and assembled to simulate the performance of interfaces operating at both constant and variable speed. Representative examples of performance for moderate and large
adverse static pressure differentials are presented and
discussed. It is shown that the power vs. flow characteristics
of the interface types investigated differ greatly and that these
differences are affected by the value of the static pressure
differential at the interface. The models presented in this paper
are of practical value in the design of pumping interfaces for
all types of open systems.
INTRODUCTION
Stratified chilled water thermal storage reduces peak
cooling plant chiller capacity requirements, generates operating cost savings, and improves the operational flexibility of
large district and campus chilled water systems. A critical
feature of a chilled water thermal storage system is the interface between the storage tank, typically an open vessel, and
the closed, pressurized chilled water system it serves. The
location of the storage tank can create a significant adverse
static pressure differential, which must be overcome by transfer pumps. The energy cost of pumping against a static pressure differential may substantially reduce thermal storage
operating cost savings.
Several options for interface design, with varying ranges
of capital cost and energy consumption, are available to the
design engineer. In order to select the best method for an application, modeling of interface performance is essential. This
paper surveys transfer pumping interface types, some of
which have not been documented previously, and describes
practical techniques for modeling the performance of interfaces. Modeled power consumption characteristics of a variety of interface types are illustrated under low and high fixed
head conditions. The use of these models in economic
comparisons of interface alternatives is described in part 2 of
this paper (Bahnfleth and Kirchner 1998). Transfer pumping
interfaces are also found in other storage applications with
open tanks including ice harvesters, containerized ice
systems, and external melt ice-on-pipe systems. The analysis
described in this paper is adaptable to these systems as well.
TRANSFER PUMPING INTERFACE TYPES
Transfer pumping interfaces control pressure, modulate
flow rate, and regulate flow direction in chilled water storage
systems. Interfaces can be divided into two classes: indirect
and direct. These classifications indicate the method
employed by the pumping interface to control pressure differential and transfer thermal energy between the open tank and
the closed system. An indirect pumping (IP) interface physically isolates the water stored in the tank from the rest of the
William P. Bahnfleth is an assistant professor in the Department of Architectural Engineering at Penn State University, University Park, Pa.
Christopher G. Kirchner is a project mechanical engineer at SHG, Inc., Detroit, Mich.
THIS PREPRINT IS FOR DISCUSSION PURPOSES ONLY, FOR INCLUSION IN ASHRAE TRANSACTIONS 1999, V. 105, Pt. 1. Not to be reprinted in whole or in
part without written permission of the American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc., 1791 Tullie Circle, NE, Atlanta, GA 30329.
Opinions, findings, conclusions, or recommendations expressed in this paper are those of the author(s) and do not necessarily reflect the views of ASHRAE. Written
questions and comments regarding this paper should be received at ASHRAE no later than February 13, 1999.
BACK TO PAGE ONE
chilled water system by use of a heat exchanger. A direct
pumping (DP) interface permits the flow of water between the
storage tank and chilled water system. Direct interfaces may
be further classified on the basis of whether they incorporate
hydraulic energy recovery. Direct interfaces with energy
recovery may be unidirectional or reversible. Each of these
types and subtypes is described below.
Indirect Interface
Indirect pumping in chilled water thermal energy storage
(TES) systems provides hydraulic isolation from high staticpressure loads that occur in high-rise buildings and in district
cooling systems with large variations in site elevation. A heat
exchanger placed between tank and system creates independent hydraulic circuits, each with its own pumps, piping, and
controls (Figure 1). This arrangement minimizes pressure
disturbances, simplifies system design, reduces pump motor
size, and eliminates the need for pressure accommodation
between the storage tank and chilled water system. It also
eliminates the pump energy consumption that would be
needed to overcome the static pressure differential between
the tank and system. The design temperature differential
between the cool inlet and warm outlet temperatures of the
interface heat exchanger is typically no less than 2°F (1.1°C).
This temperature differential imposes thermodynamic penalties on the storage system and chilled water system that detract
from the energy savings created by hydraulic separation and
may significantly increase capital cost relative to a system
having a direct interface.
Direct Interface
As illustrated in Figure 2, direct pumping allows chilled
water from the thermal storage tank to be pumped directly into
the chilled water supply system while return water flows from
the system to the tank. Transfer pumps boost the pressure of
water leaving the tank to match the system level, and pressuresustaining valves (PSVs) throttle away excess pressure of
water flowing into the tank. A direct interface maximizes the
temperature differential across the storage tank by utilizing the
lowest supply and highest return temperatures possible. This
results in higher storage density and smaller tank size than are
possible when an indirect interface is used. The first cost of a
direct interface is less than that of an indirect interface because
it requires less equipment, fewer controls, and fewer reversing
valves. However, when the static pressure differential
becomes large, the economic impact of factors such as pump
size, PSV size, PSV reliability, and pumping energy can
become significant (Gatley and Mackie 1995). In extreme
cases, the energy required to pump against a static differential
may account for 20% or more of the cost to produce and
deliver a unit of thermal energy from storage (Bahnfleth
1995).
Figure 1 Indirect pumping interface.
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Figure 2 Direct pumping interface.
Direct Interface with Energy Recovery
Direct pumping with hydraulic turbine energy recovery
(DPT) is an alternative to indirect pumping for high static pressure applications. In a DPT interface, an energy recovery
turbine (ERT) in the return line to the tank is connected to a
transfer pump through a double extended motor shaft (Figure
3). The turbine operates in series with a pressure-sustaining
valve to absorb excess system pressure as water flows back to
the tank, converting energy associated with static pressure
differential into shaft power that directly reduces pump motor
input power. This allows the storage tank to be connected
directly to the chilled water system without paying the full
energy penalty of an adverse static pressure differential. As
flow through the turbine decreases from its design value, the
fraction of the available energy recovered decreases. When it
can no longer absorb energy from the system, a clutch disengages the turbine to prevent it from becoming an additional
load on the motor. In typical applications, a turbine recovers
35% to 50% of the peak pump motor power (Bahnfleth 1995;
Taylor 1989). Although relatively uncommon, DPT interfaces
have been successfully applied in several high-rise buildings
and district energy installations (Bahnfleth 1995; Taylor 1989;
Tackett 1988)
Reversible Direct Interface with Energy Recovery
Reversible direct pumping with hydraulic turbine energy
recovery (RDPT) is a special case of the DPT interface. Instead
of reversing flow through the tank by changing the position of
the flow control valves, the reversible interface changes flow
direction by reversing the rotation of the pump/turbine set
(Figure 4). The charge mode pump becomes the discharge
mode turbine and vice versa. A voltage inverter changes the
direction of rotation of the motor driving the set to reverse the
flow direction. A variable-speed RDPT transfer pumping
system is currently in use at a semiconductor manufacturing
facility (Fiorino 1994). The turbines in this system unload
each motor by approximately 40%. In an RDPT interface, each
pump/turbine must be identical. Selection for pump performance takes precedence over turbine performance with the
consequence that turbine performance is less than ideal.
Interface Pump, Turbine, and System Curves
Transfer pumps must overcome the sum of the static pressure differential and frictional resistance to flow imposed on
them. In an indirect interface, the transfer pumps on both sides
of the heat exchanger work only against frictional resistances.
In a direct interface, transfer pumps must overcome adverse
static pressure differential plus the frictional resistance of the
path from the outlet diffuser to the point of connection with the
chilled water system. The path from the other point of connection to the inlet diffuser is not part of the system for the direct
interface transfer pump. However, for interfaces with energy
recovery, this return path is the system for the hydraulic
turbine and affects the pressure available to it.
Figure 5 shows typical shapes of pump and turbine characteristic curves and system curves for a direct interface. The
pump curve is taken from a manufacturer’s data. The turbine
characteristic was generated using performance prediction
methods described in the following section. Pump and turbine
system curves in a direct interface with energy recovery typically mirror one another because static pressure is identical
and frictional pressure drop is similar. The pump pressure
differential required increases as flow increases, while the
turbine available pressure differential decreases. Pump and
turbine pressure-flow characteristics also have a qualitatively
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Figure 3 Direct pumping interface with hydraulic energy recovery turbine.
Figure 4 Reversible direct pumping interface.
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Figure 5 Typical pump, turbine, and system curves for a direct interface.
mirror-image relationship. The pump pressure differential
falls with increasing flow, while the turbine pressure differential absorbed increases. When a turbine operates at constant
speed, it will generally consume less pressure differential than
is available from the system, as shown in Figure 5. This excess
system head must be throttled away by a PSV, as also indicated
in the figure. This is analogous to the behavior of a constantspeed pump under variable-flow conditions, for which excess
pump head must be wasted by a control valve.
Sharma 1985; Stepanoff 1957). The pump affinity laws are
valid for centrifugal pumps operating as turbines and can be
used to predict performance once a point on the turbine curve
is known (Garay 1993; Ingersoll 1976). The literature
describes several methods for estimating the BEP performance of centrifugal pumps operating as turbines on the basis
of pump performance characteristics (Buse 1981; Sharma
1985; Stepanoff 1957). Two distinct approaches have been
used, which for convenience will be identified as the “best
efficiency point method” and the “specific speed method.”
TURBINE PERFORMANCE OF REVERSE
RUNNING CENTRIFUGAL PUMPS
Best Efficiency Point Method
Specially built hydraulic turbines are prohibitively
expensive for thermal storage applications. However, standard centrifugal pumps running backwards have acceptable
turbine characteristics and are substantially less expensive.
Reverse running pumps have been used successfully as ERTs
in cool storage DPT and RDPT interfaces serving high-rise
buildings and district energy systems (Bahnfleth 1995;
Fiorino 1994; Tackett 1988; Taylor 1989).
The fundamental assumption of the best efficiency point
method is that the BEP efficiencies of a centrifugal pump operating as a turbine (ηt,BEP) and as a pump (ηp,BEP) are equal
(Sharma 1985; Stepanoff 1957). Taking this assumption as a
starting point, Sharma developed the following relations for
converting pump BEP volume flow rate (Qp,BEP) and pressure
(Hp,BEP) to the corresponding values of turbine BEP flow rate
(Qt,BEP) and pressure (Ht,BEP):
The concept of using reverse-running centrifugal pumps
as turbines is well documented in the literature (Buse 1981;
Garay 1993; Nelik and Cooper 1984; Sharma 1985; Stepanoff
1957). Testing has shown that a properly selected centrifugal
pump, when acting as a turbine, has a peak hydraulic efficiency within a few percentage points of its best performance
as a pump (Buse 1981; Ingersoll Rand 1976). This difference
is typically 2% to 3% at the best efficiency point (BEP). For
practical purposes, turbine and pump peak efficiency can be
considered equal, according to several sources (Garay 1993;
Q t , BEP =
H t , BEP =
Q p , BEP
η p , BEP
0 .8
H p , BEP
η p , BEP
1 .2
≡ C q ⋅ Q p , BEP
(1)
≡ C h ⋅ H p , BEP
(2)
where Cq and Ch are conversion factors relating turbine BEP
flow and pressure to pump BEP values of the same parameters.
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Specific Speed Method
An alternative approach to the BEP method is based on
conversion factors that are functions of turbine specific speed
(Buse 1981). Turbine specific speed is a similarity parameter
combining turbine BEP speed, flow rate, and pressure:
Ns =
N BEP × QBEP
H BEP
0.75
(3)
where Ns is specific speed, NBEP is BEP speed, QBEP is BEP
flow rate, and H B E P is BEP pressure. BEP speed is
expressed in rpm. In the I-P system, flow has units of gpm
and pressure is given in feet of head. In the SI system, flow
is given in m 3 /s and pressure is given in meters of head.
Consequently, the numerical value of Ns in I-P units is 51.65
times the value in SI units.
Buse presents data on values of conversion factors Cq, Ch,
and Ce, which relate, respectively, turbine BEP flow rate
(Qt,BEP), pressure (Ht,BEP), and efficiency (ηt,BEP) to pump
BEP values. Given a desired turbine selection point (flow rate
and pressure), these conversion factors permit rapid calculation of required pump characteristics:
Q p , BEP =
Qt , BEP
H p , BEP =
H t , BEP
Cq
Ch
η t , BEP = η p , BEP ⋅ C e
(4)
(5)
(6)
According to Buse, Cq and Ch vary from 2.2 to 1.1 and Ce
varies from 0.92 to 0.99 over a specific speed range of 500 to
2800 (9.7 to 54.2). Considering these values, it is evident that
(1) peak turbine efficiency is less than pump peak efficiency
and occurs at a higher flow rate, and (2) the pressure drop in
turbine duty is greater than the pressure developed by the
pump. Consequently, the maximum BHP is higher in turbine
duty.
Comparison of BEP Performance Models
Williams (1994) compared the BEP performance predictions of eight pump-to-turbine prediction methods to the
measured turbine performance of 35 different types of pumps.
Predicted and measured BEP values for turbine flow rate and
head were compared via a prediction coefficient having values
varying between zero and one when the sum of square errors
in predicted flow and head is within acceptable tolerances.
Prediction coefficient values greater than one indicate unacceptably large errors and a value of zero indicates a perfect
prediction. Sharma’s best efficiency point method yielded the
lowest average prediction coefficient for all of the pumps
tested, and only 20% of tested pumps had prediction coefficients greater than one. In contrast, 81% of predictions using
the least accurate method had prediction coefficients greater
than one.
Based on Williams’ results, Sharma’s best efficiency
point method is the best available procedure for preliminary
determination of turbine BEP performance from peak pump
efficiency and has acceptable accuracy for that purpose.
However, both Williams (1994) and Buse (1981) emphasize
that testing should be undertaken to determine actual turbine
performance once a preliminary selection has been made. The
specific speed method has been utilized in the design of
several successful DPT and RPDT systems for chilled water
TES applications (Bahnfleth 1995; Tackett 1988; Taylor
1989). It has proved reasonably reliable for the selection and
performance prediction of centrifugal pumps used as turbines
when the manufacturer’s conversion factors are used.
However, manufacturers frequently consider conversion
factors proprietary, which makes the independent application
and testing of this method difficult. For this reason, the
specific speed method was not tested by Williams.
Turbine Performance Curves
The best efficiency point and specific speed methods
predict turbine performance only at the best efficiency point.
Other means must be employed to predict turbine behavior at
other operating points. Buse (1981) published representative
curves showing constant-speed turbine pressure and efficiency ratio variation as a function of percentage of turbine
BEP flow rate for a range of specific speeds. The shape of the
performance curve for a turbine can be developed from these
curves given the turbine specific speed and BEP performance
estimated by the BEP or specific speed method (Buse 1981).
Typical constant-speed pressure and efficiency characteristics for a centrifugal pump (from manufacturer’s data) and
the corresponding turbine performance curves generated
using Sharma’s best efficiency point method and Buse’s
curves are shown in Figure 6. The best efficiency point for the
pump is 2070 gpm (131 L/s) with a head of 217 ft (67 m) of
water, at which it attains a hydraulic efficiency of 90%. The
corresponding BEP for the pump operating as a turbine determined by using the best efficiency point method is 2252 gpm
(142 L/s) and 246 ft (75 m) of water, at which efficiency is also
assumed to be 90%. These curves can be generalized to the
variable-speed case through the application of affinity laws, as
is described in the next section.
TRANSFER PUMPING INTERFACE
COMPONENT MODELS
Heat Exchanger
A simple heat exchanger model can be developed using
LMTD analysis. Pure counterflow has been assumed in the
example presented here; other flow arrangements can be
modeled using correction functions available in the literature.
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Figure 6 Predicted turbine performance of a centrifugal pump: BEP method.
In the LMTD formulation, the load, q, on the heat exchanger,
is given by
q = U ⋅ A ⋅ LMTD
(7)
where U is the overall heat transfer coefficient, A is the heat
exchange area of the heat exchanger, and LMTD is the log
mean temperature difference. The LMTD is defined as
follows:
( T hi – T co ) – ( T ho – T ci )
LMTD = ----------------------------------------------------------T hi – T co
ln  ---------------------
 T ho – T ci
(8)
where Tci and Tco are the cold fluid inlet and outlet temperatures, respectively, and Thi and Tho are the hot fluid inlet and
outlet temperatures.
Inlet and outlet temperatures on the storage tank side of
the heat exchanger and inlet temperature on the system side
were assumed to be known and constant. In practice, this could
be achieved by controlling flow rate to regulate the leaving
temperature. On the system side of the heat exchanger, the
inlet temperature was assumed to be known and constant for
both charge and discharge modes. The charge and discharge
outlet temperatures on the system side of the heat exchanger
do not remain constant under varying loads and must, therefore, be calculated. Design temperatures representative of
conditions in actual chilled water storage systems with indirect interfaces are shown in Figure 7. Of particular significance is the design heat exchanger temperature differential of
2°F (1.1°C), which reduces the maximum temperature differential across the storage tank by 4°F (2.2°C) over a complete
cycle. This reduces storage density by nearly 20% for the
assumed system charge and return temperatures and commensurately increases the size of the tank and peak flow rate for a
given capacity.
The unknown system-side charge and discharge outlet
temperatures can be expressed in terms of the fraction of full
design load (φ). As a consequence of Equation 7,
φ=
q
U ⋅ LMTD
=
q d U d ⋅ LMTD d
(9)
where qd, Ud, and LMTDd are the load, overall heat transfer
coefficient, and log mean temperature difference at design
conditions, respectively. In general, the overall heat transfer
coefficient decreases as flow decreases because of reductions
in convective heat transfer coefficients. In some instances,
however, it is an acceptable approximation to assume that U is
constant.
In an indirect transfer pumping interface, heat exchanger
performance determines flow rate, which governs pumping
power. In the present study, temperature differential is specified on one side of the heat exchanger; therefore, flow variation due to variation in U will affect only the other side.
Additionally, the indirect interfaces considered in this study
were assumed to have four heat exchangers in parallel, and
these heat exchangers were staged to maintain the highest
possible velocity for any operating configuration. Although
the maximum error in flow on one side of the heat exchanger
could exceed 10%, the average error in total pumping power
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Figure 7 Heat exchanger parameters: (a) discharge cycle, and (b) charge cycle.
for both sides of the heat exchanger was estimated to be less
than 5%. On this basis, it was assumed that U was constant.
For constant U, it follows from Equation 9 that the LMTD
under an arbitrary loading is φ (LMTDd). Forms of Equation 9
particular to the charge and discharge modes can be solved for
their respective system-side outlet temperatures. For the
charge mode,
(10)
( T ho – T ci ) – ( ( T co – T hi ) – ( T ci – T ho ) )
T co = T hi – EXP ln ------------------------------------------------------------------------------------------------φ ⋅ ( LMTD d )
and for discharge mode,
(11)
( T hi – T co ) – ( ( T hi – T co ) – ( T ho – T ci ) )
T ho = T ci – EXP ln ------------------------------------------------------------------------------------------------φ ⋅ ( LMTD d )
Equation 10 is implicit in the unknown Tco and Equation
11 is similarly implicit in Tho with all other quantities being
known for a specified value of the part-load fraction. These
equations are easily solved by successive substitution or other
iterative techniques. Figure 8 shows outlet charge and
discharge temperatures as a function of the load fraction (φ)
obtained by solving Equations 10 and 11 for the design condi
tions shown in Figure 7. If all temperatures other than the
system-side outlet are assumed constant, it is convenient to
use polynomial approximations of the functions shown in
Figure 8 in place of repeated iterative solutions of Equations
10 and 11 in the indirect interface model. For example, for
charging,
Tco = W1φ 3 + W 2φ 2 + W 3φ + W 4
(12)
where the Wi are constant regression coefficients. Equation 12
can be evaluated explicitly; therefore, its use reduces computational time. This is an advantage when the interface model
is used in hourly annual simulation requiring solutions for
thousands of operating points.
After the system-side outlet temperature has been determined, the system-side flow rates needed for calculation of
pumping energy can be obtained. The charge cycle flow rate,
Qs,c, is
Q s ,c =
q
ρ c p (T co − T ci )
(13)
where cp and ρ are the fluid specific heat and density. Similarly, for discharge,
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Figure 8 Heat exchanger system-side outlet temperature.
Q s ,d =
q
ρ c p (T hi − T ho ) .
Variable-Speed Pump and Turbine
(14)
Constant-Speed Pump and Turbine
The pressure and efficiency characteristics of centrifugal
pumps and centrifugal pumps operating as turbines are functions of their respective specific speeds (Ns). Pressure (H) and
efficiency (η) data for pump performance as functions of flow
rate (Q) are easily obtained from manufacturer’s literature and
selection software. Some pump manufacturers can furnish
limited turbine performance information upon special request,
but this information is not easily available. Application of
methods for estimating turbine performance, such as those
described previously, are generally more useful for conducting preliminary studies.
Typical constant-speed pressure and efficiency characteristics, based on the horizontal split case centrifugal pumps of
a major manufacturer, are shown in Figures 9a and 9b, respectively. For each specific speed, these characteristics were
modeled by second-order polynomial regressions. The regression coefficients themselves were then modeled as secondorder functions of specific speed. Thus, the model for H(Q, Ns)
for either a pump or turbine has the form
H = C 1Q 2 + C 2 Q + C 3
(15)
2
(16)
where E n = F1, n N s + F2, n N s + F 3, n . The specific speed (Ns)
utilized in equations Cn and En refers to pump or turbine
specific speed as appropriate.
H = C 1Q
2
 N 
 N 
 + C 3 

+ C 2 Q 
N
 0 
 N0 
2
(17)
A derivation of Equation 17 is given in the appendix.
When Equation 17 is used to model a variable-speed
pump operating to provide system flow, Qsys, at system pressure, Hsys, the required speed can be determined by means of
the quadratic formula:
(18)
[(
)
 − C Q + C − 4C C ⋅ Q
2 Sys
1 3
N =
2C 3

2
2
2
Sys
+ 4C 3 H Sys
]
0.5

 ⋅ N0

The variable-speed efficiency map for a pump or turbine
is developed by applying affinity relations to the constantspeed efficiency model (Equation 16) in a manner similar to
that used in the derivation of Equation 17:
2
2
where C n = D 1, n Ns + D 2, n N s + D 3, n . Similarly, pump or
turbine efficiency may be expressed as
η = E 1Q 2 + E 2 Q + E 3
Performance models for variable-speed pumps and
turbines can be developed by applying pump affinity laws to
constant-speed head and efficiency characteristics (Equations
15 and 16). For a pump or turbine with pressure characteristic
H0 defined as a function of flow Q0 for operation at reference
speed N0, the characteristic at an arbitrary speed N is
 N
 N 
 Q 2 + E 2 
η = E 1 
 N0
 N0 

 Q + E 3 .

(19)
When energy recovery is employed in a direct pumping
interface, the turbine typically is connected to the pump
through a double extended motor shaft. Therefore, pump and
turbine must operate at the same speed. Because the speed of
the pump and turbine set is determined by the requirements of
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Figure 9 Effect of specific speed on constant-speed pump performance characteristics: (a) pressure vs. flow rate, (b)
efficiency vs. flow rate. Ns is given in IP units: flow in gpm, head in ft, speed in rpm.
the pump, it is the pump system characteristics that control
motor speed and the motor acts as a governor for the turbine.
Q ⋅ H ⋅η t
W& t =
K
(21)
The shaft power input required by a pump is
W& p
Q⋅H ,
=
K ⋅η p
and the shaft power output of a turbine is
(20)
where ηp and ηt are pump and turbine efficiencies, respectively, and K is a constant required for consistency of units. In
I-P units, K has a value of 3960 for water at 68°F (20°C) when
shaft power is given in horsepower, flow in gallons per
minute, and pressure in feet of water. In the SI system, K is
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1000 when power is expressed in kW, flow in L/s, and pressure in kPa. Because the variation in water density is negligible over the range of temperatures found in chilled water
systems, these common engineering expressions may be used
in place of variable property forms with little error.
Electric Motor and Variable-Speed Drive Models
The power required to drive a centrifugal pump in the
absence of energy recovery can be expressed in terms of
pump, motor, and drive efficiencies. Motor power is given by
W& m =
W& p
(22)
ηm
where ηm is the motor efficiency. The efficiency of an electric
motor varies as a function of the fraction of its full-load power
rating. A generic high-efficiency motor performance curve
from the literature (ASHRAE 1996), approximated by a
piece-wise continuous polynomial, was used in this investigation.
·
The total electric power input, W total , required to operate
a variable-speed pump is given by
W&
W& total = m
ηd
(23)
where ηd is the efficiency of the variable-speed drive controlling the pump motor. If the pump has no variable-speed drive,
ηd is equal to one. In the present study, a generic high-performance variable-speed drive efficiency curve vs. fraction of full
speed was obtained from the literature (ASHRAE 1996) and
approximated as a quadratic function of the ratio of operating
speed to full speed.
The shaft power produced by an energy recovery turbine
is deducted directly from the output that must be produced by
the pump motor to which it is connected. The total power
required by an interface with energy recovery is, therefore,
W& total =
W& p − W& t
η m ⋅ηd
=
Q (H p − H tη pη t ),
Kη pη mη d
(24)
where Hp and Ht are the pump and turbine heads, respectively.
Because motor efficiency depends upon the fraction of nameplate power being consumed, this reduction in motor power
alters the motor efficiency; however, because pump speed is
determined solely by pumping requirements, drive efficiency
is not affected by the addition of the turbine.
TYPICAL PERFORMANCE CHARACTERISTICS
Seven interface energy consumption models were developed from the component models described above. These
included indirect interfaces (Figure 1) operated at both
constant (IC) and variable (IV) speed, direct interfaces without energy recovery (Figure 2) operated at constant (DC) and
variable (DV) speed, direct interfaces with energy recovery
(Figure 3) operated at constant (DCT) and variable (DVT)
speed, and a reversible, variable-speed direct interface with
energy recovery (RDVT, Figure 4). These interface types and
their acronyms are summarized for reference in Table 1.
TABLE 1
Summary of Interface Types
Description
Acronym
Indirect, Constant Speed
IC
Indirect, Variable Speed
IV
Direct, Constant Speed
DC
Direct, Variable Speed
DV
Direct with Energy Recovery, Constant Speed
DCT
Direct with Energy Recovery, Variable Speed
Reversible Direct with Energy Recovery, Variable Speed
DVT
RDVT
To illustrate the differing performance characteristics of
various interface types, total power required as a function of
flow rate was calculated for two levels of adverse static pressure differential between the chilled water system and thermal
storage tank (see Kirchner [1997] for full details regarding
component selection, interface design, and model coefficients). Frictional head in each case was determined by takeoff
from the piping layout particular to the interface. The lower
fixed-head value, 130 ft (40 m), is relatively small but has
potential for application of energy recovery in a direct interface. The higher value, 260 ft (79 m), represents the static
differential of a twenty-story-tall building and may be sufficiently large to justify the use of an indirect interface. Actual
thermal storage interfaces typically have two to three pump
sets, so three parallel sets were assumed in this example.
Pumps were selected with best efficiency points at 2,070 gpm
(131 L/s) for direct interfaces and 2,530 gpm (159 L/s) for the
indirect interfaces. Peak efficiencies varied from 80% to 88%.
Motor nameplate and variable-speed drive, full-speed efficiencies were assumed to be 95%. Chilled water supply and
return temperatures and nominal heat exchanger performance
were as shown in Figure 7.
Figures 10 and 11 show electric power required by the
seven interface types as a function of flow rate for the lower
and higher fixed head cases, respectively. Transitions between
operation with one pump and two pumps and between two
pumps and three pumps are indicated by the discontinuities in
the power vs. flow curves. An important difference between
direct and indirect interface performance, apparent in both
Figure 10 and Figure 11, is that the maximum flow rate for
indirect alternatives is approximately 20% higher than for
direct alternatives due to the reduced temperature differential
associated with the use of the heat exchanger. Nevertheless,
both constant- and variable-speed indirect interfaces consume
substantially less pumping power than direct interfaces in the
260 ft (79 m) case (Figure 11). This is because the heat
BACK TO PAGE ONE
Figure 10 Interface power vs. flow rate, 130 ft (40 m) of static pressure: (a) alternatives without energy recovery, (b)
alternatives with energy recovery (IV and DC also shown for comparison—see Table 1 for acronym definitions).
BACK TO PAGE ONE
Figure 11 Interface pumping power vs. flow rate, 260 ft (79 m) of static pressure: (a) alternatives without energy recovery,
(b) alternatives with energy recovery (IV and DC also shown for comparison—see Table 1 for acronym
definitions).
BACK TO PAGE ONE
exchanger eliminates the head associated with moving flow
across the interface against the static pressure differential by
separating the system into two hydraulically independent
zones. This is not the case for the 130 ft (40 m) example, as
indicated by Figure 10. The variable-speed indirect system
requires less power than the direct interfaces at virtually all
flow rates, but the constant-speed indirect system requires
total power comparable to direct interfaces without energy
recovery.
Turbines in DCT and DVT interfaces recover approximately half of the pumping power required at the peak flow
rate of each system configuration (i.e., with one, two, or three
sets fully loaded). Figures 10b and 11b indicate that the turbine
in the DCT interface reduces the peak pumping power by
approximately 40% at 130 ft (40 m) of fixed head and by 54%
at 260 ft (79 m) of fixed head relative to the DC system. The
higher recovery rate in the 260 ft (79 m) case is due to the
larger component of static pressure in the system curve.
Generally speaking, direct interfaces consume more power
than direct interfaces with energy recovery and still more than
indirect interfaces. However, because pumping power
required to overcome friction cannot be recovered, the
percentage of total pumping power reclaimed decreases as the
design ratio of static head to total pumping head decreases. At
low static head, the differences in performance between the
three interface types are small, but they increase greatly as
static head increases.
Turbine power output in the RDVT case and the corresponding peak demand reduction are significantly less than for
either the DCT or DVT systems as a result of the suboptimal
turbine performance of the full-size pump. Net power
consumption for the RDVT and DCT interfaces is similar.
However, RDVT energy recovery falls off more rapidly with
decreasing flow rate. At the peak flow rate of 6,846 gpm (440
L/s), the turbine reduces peak electrical demand by 43%, relative to the DV system. However, when the flow rate is reduced
by 26% to 5,052 gpm (320 L/s), the reduction in power is only
25%. Therefore, the turbine in the RDVT system recovers a
very small amount of energy at low flow and tends to operate
more like a DV system near the changeover points. By
comparison, the DVT system operating with an optimum
turbine for the same conditions reduces peak demand by 54%
at full flow and by 45% at the reduced flow rate.
Typical turbine performance characteristics are shown in
Figure 12 for an interface with three parallel pump and turbine
sets operating in a system with a static pressure of 260 ft (79
m) of water. The turbine system curve is shown in Figure 12a.
It is typical of system characteristics well suited to the application of energy recovery turbines because of its large fixedhead component. Figure 12a shows that the variable-speed
turbine (DVT) has a lower operating pressure for a given flow
rate than either of the other alternatives. The reversible pump/
turbine, which is not optimized for turbine performance,
requires the greatest pressure and operates over a more
restricted range of flow rate. Power recovered by turbines as
a function of flow rate is shown in Figure 12b. The suboptimal
reversible turbine (RDVT) recovers less power than either of
the other alternatives. The difference in power recovered by
the optimized DVT and DCT turbines is small.
Note that the maximum power of a given DCT interface
configuration does not occur at its peak flow rate (Figures 10
and 11). This is because power recovered by the turbine in a
DCT interface decreases more rapidly than does the power
required by the pump as flow decreases. This is not the case for
variable-speed operation. In this case, peak power demand
occurs at full flow rate but remains relatively constant as flow
decreases. Figure 12b show that recovered power for constantand variable-speed turbines is quite similar; therefore, the
difference between the DCT and DVT cases in Figures 10b
and 11b are due primarily to the effect of speed variation on
pumping power.
CONCLUSIONS
Empirical models suitable for simulating transfer pumping interface performance have been derived and utilized to
illustrate the energy consumption characteristics of a variety
of chilled water thermal storage interface types. These models
can be employed to evaluate the performance of interface
alternatives in the preliminary stages of system design.
Performance models and selection methods for the application of centrifugal pumps as hydraulic turbines can be developed using published information and pump/turbine affinity
laws. Such models are suitable for preliminary design only.
For final design, a pump manufacturer should be consulted for
specific equipment selections and performance data.
Despite statements in the published literature (Taylor
1989), the present study identified no technical considerations
precluding the use of energy recovery turbines with variablespeed pumping. It is not necessary to sacrifice the advantages
of one to obtain the other. Variable-speed operation reduces
the amount of energy recovered by the turbine slightly because
turbine head is reduced relative to constant-speed operation;
however, it substantially improves the performance of the
pump in such an interface.
The various interface alternatives discussed have greatly
varying power consumption characteristics. They also differ
substantially in construction cost, as may be inferred from the
relative complexity and equipment quantities indicated in
Figures 1 through 4. As a result, significant trade-offs exist
between capital and operating costs. In order to determine the
best alternative for an application, economic analysis relating
annual operating cost to capital cost must be performed. The
second part of this paper will demonstrate the application of
detailed interface models in a parametric study of interface
energy consumption, energy cost, simple payback, and lifecycle cost as a function of adverse static pressure differential.
BACK TO PAGE ONE
Figure 12 Typical energy recovery turbine performance, 260 ft (79 m) of static pressure: (a) pressure vs. flow rate, (b) output
shaft power vs. flow rate (see Table 1 for acronym definitions).
ACKNOWLEDGMENTS
tional District Energy Association. 86th Annual Conference.
The authors wish to extend their appreciation to Mr. Steve
Gavlick of the PACO Pump Corporation for his assistance in
making the pump selections needed in this research. Partial
support of this work through an ASHRAE Graduate Grant-inAid is also gratefully acknowledged.
Bahnfleth, W.P., and C.G. Kirchner. 1998. Analysis of transfer pumping interfaces for stratified chilled water thermal storage systems—Part 2: Parametric study.
ASHRAE Transactions 105(1).
REFERENCES
Buse, F. 1981. Using centrifugal pumps as hydraulic turbines. Chemical Engineering, Jan. 26, pp. 113-117.
ASHRAE. 1996. 1996 ASHRAE handbook—HVAC systems
and equipment. Atlanta: American Society of Heating,
Refrigerating and Air-Conditioning Engineers, Inc.
Bahnfleth, W.P. 1995. Hydraulic issues in the design of
chilled water storage systems. Proceedings, Interna
Fiorino, D. 1994. District cooling re-invented. ASHRAE
Journal, 36(5): pp. 20-28. May.
Garay, P.N. 1993. Standard pumps used as hydraulic turbines
for power production. Pump Application Deskbook, pp.
377 - 396. New Jersey: Fairmount Press.
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Gatley, D.P., and I. Mackie 1995. Cool storage open
hydronic systems design guide. EPRI, TR-104906.
Electric Power Research Institute.
Ingersoll Rand. 1976. Hydraulic turbines. Cameron Pump
Manual, M-7006.10. New Jersey: Ingersoll-Rand Company.
Kirchner, C.G. 1997. Analysis and modeling of transfer
pumping for stratified chilled water thermal storage systems. M.S. thesis, Department of Architectural Engineering, The Pennsylvania State University.
Nelik, L., and P. Cooper 1984. Performance of multi-stage
radial-inflow hydraulic power recovery turbines. ASME
Winter Meeting Conference Paper. New York: American Society of Mechanical Engineers.
Sharma, K.R. 1985. Small hydroelectric projects—Use of
centrifugal pumps as turbines. Kirloskar Electric Company, Bangalore, India.
Stepanoff, A.J. 1957. Centrifugal and axial flow pumps, pp.
276 - 277. New York: John Wiley and Sons.
Tackett, R.K. 1988. The use of direct pumping and hydraulic
turbines in thermal storage systems. ASHRAE Transactions 94(1): 1989 - 2007.
Taylor, R.M. 1989. Chilled water storage tank requires special consideration when designing a piping distribution
system. IDHCA Cooling Conference.
Williams, A.A. 1994. The turbine performance of centrifugal
pumps: A comparison of prediction methods. Journal of
Power and Energy, Part A, pp. 59 - 66.
DERIVATION OF VARIABLE-SPEED
PUMP CHARACTERISTIC
(A2)
where H is head. A quadratic approximation to the pump characteristic at the reference speed N0 may be written in the form
H 0 = C 1 Q 02 + C 2 Q 0 + C 3
(A3)
where C1, C2, and C3 are regression coefficients. From Equation A2,
 N 

H = H 0 
 N0 
2
(A4)
Therefore, substituting for H0 from Equation A3,
(
H = C 1Q
2
0
+ C 2Q 0 + C 3
 N 

⋅ 
 N0 
)
2
(A5)
From Equation A1,
 N 
Q0 = Q ⋅ 0 
 N 
(A6)
Substituting from Equation A6 for Q0 in Equation A5
gives
(A7)
which simplifies to Equation 17:
According to affinity laws for homologous pumps,
N 2
H
------ =  ------
 N0
H0
2
N0
N 2
2 N0
H =  C 1 Q  ------ + C 2 Q  ------ + C3 ⋅  ------ ,

  N 0
 N
 N
APPENDIX A
Q
N
=
Q0 N 0
where Q is head, N is speed, and the subscript 0 denotes a
reference or nominal condition.
(A1)
 N
 N 
 + C 3 
H = C 1 Q 2 + C 2 Q 
 N0
 N0 



2
(A8)
Equation 19 may be derived by a similar procedure.