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This paper presents an analysis of various CO2 transcritical and cascade/secondary looprefrigeration systems that are becoming popular in supermarket applications with theobjective of optimizing the operating parameters of these systems. In addition, the performanceof selected CO2-based refrigeration systems is compared to the baseline R404Amultiplex direct expansion system using bin analyses in the eight climate zones of theUnited States. For the refrigeration systems investigated, it was found that the TranscriticalBooster System with Bypass Compressor (TBS-BC) had the lowest energy consumptionfor ambient temperatures (Tamb) less than 8 C, and for higher ambienttemperatures the R404A direct expansion system was found to have the lowest energyconsumption. Also, the TBS-BC performs equivalent to or better than the R404A directexpansion system in the northern two-thirds of the US. For the southern portion of the US,the R404A multiplex DX system performs better than CO2 systems.
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i n t e r n a t i o n a l j o u r n a l o f r e f r i g e r a t i o n 4 6 ( 2 0 1 4 ) 8 6e9 9
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Comparative analysis of various CO2 configurationsin supermarket refrigeration systems*
Vishaldeep Sharma, Brian Fricke*, Pradeep Bansal
Oak Ridge National Laboratory, PO Box 2008, Oak Ridge, TN 37831-6070, USA
a r t i c l e i n f o
Article history:
Received 11 April 2013
Received in revised form
10 June 2014
Accepted 2 July 2014
Available online 9 July 2014
Keywords:
Refrigeration
Transcritical booster system
Cascade system
Secondary loop system
Carbon dioxide
Thermodynamics
* Notice: This manuscript has been authoreEnergy. The United States Government retainStates Government retains a non-exclusive,manuscript, or allow others to do so, for Un* Corresponding author. Tel.: þ1 865 576 082E-mail address: [email protected] (B. Fric
http://dx.doi.org/10.1016/j.ijrefrig.2014.07.0010140-7007/© 2014 Elsevier Ltd and IIR. All rig
a b s t r a c t
This paper presents an analysis of various CO2 transcritical and cascade/secondary loop
refrigeration systems that are becoming popular in supermarket applications with the
objective of optimizing the operating parameters of these systems. In addition, the per-
formance of selected CO2-based refrigeration systems is compared to the baseline R404A
multiplex direct expansion system using bin analyses in the eight climate zones of the
United States. For the refrigeration systems investigated, it was found that the Tran-
scritical Booster System with Bypass Compressor (TBS-BC) had the lowest energy con-
sumption for ambient temperatures (Tamb) less than 8 �C, and for higher ambient
temperatures the R404A direct expansion system was found to have the lowest energy
consumption. Also, the TBS-BC performs equivalent to or better than the R404A direct
expansion system in the northern two-thirds of the US. For the southern portion of the US,
the R404A multiplex DX system performs better than CO2 systems.
© 2014 Elsevier Ltd and IIR. All rights reserved.
Analyse comparative de diverses configurations au CO2 dansdes syst�emes frigorifiques de grandes surfaces
Mots cl�es : R�efrig�eration ; Syst�eme de suralimentation transcritique ; Syst�eme en cascade ; Syst�eme �a boucle secondaire ; Dioxyde de
carbone ; Thermodynamique
1. Introduction
Supermarkets and other large food retail stores commonly
utilize multiplex direct expansion (DX) refrigeration
d by UT-Battelle, LLC, uns and the publisher, by apaid-up, irrevocable, worited States Government p2; fax: þ1 865 574 9332.ke).
hts reserved.
systems in conjunction with synthetic refrigerants such as
R22, R404A and R507. The estimated annual refrigerant leak
rate for these systems ranges from 3% to 35% for in-use
equipment, with the higher annual leak rates (>25%) being
der Contract No. DE-AC05-00OR22725 with the U.S. Department ofccepting the article for publication, acknowledges that the Unitedld-wide license to publish or reproduce the published form of thisurposes.
Nomenclature
CR circulation ratio
h enthalpy (kJ kg�1)_m mass flow rate (kg s�1)
MR mass flow ratio
P pressure (MPa)_Q rate of heat transfer (W)
T temperature (�C)DT evaporator superheat_W power (W)
WR work ratio
Greek Symbols
Ɛ heat exchanger effectiveness
Subscripts
amb ambient
app approach
BP bypass
circ refrigerant circuit
comp compressor
cond condensing
GC gas cooler
HP high-pressure
int intermediate
LP low-pressure
LT low-temperature
MT medium-temperature
pump pump
ref refrigerating capacity
SLHX suction-liquid line heat exchanger
total sum around refrigeration cycle
Abbreviations
CFC chlorofluorocarbon
COP coefficient of performance
CSC combined CO2 secondary/cascade system
CSC-G combined glycol secondary/CO2 cascade system
DEC direct expansion cascade system
DX direct expansion
GWP global warming potential
HC hydrocarbon
HCFC hydrochlorofluorocarbon
HFC hydrofluorocarbon
HTC high temperature circuit
LTC low temperature circuit
ODP ozone depletion potential
SC secondary coolant system
SLHX suction-liquid line heat exchanger
STBS standard transcritical booster system
TBS-BC transcritical booster system with bypass
compressor
TBS-UX transcritical booster system with upstream
expansion valve
i n t e rn a t i o n a l j o u r n a l o f r e f r i g e r a t i o n 4 6 ( 2 0 1 4 ) 8 6e9 9 87
more characteristic of older equipment and the lower rates
(<15%) being more characteristic of newer equipment (ICF
Consulting, 2005). The high Global Warming Potential
(GWP) of the hydrofluorocarbon (HFC) refrigerants
commonly used in these systems, coupled with the large
refrigerant charge and the high refrigerant leakage rates
leads to significant direct emissions of greenhouse gases
into the atmosphere. Hence, these multiplex refrigeration
systems can directly contribute to the increase in global
warming.
Furthermore, the operation of refrigeration systems con-
tributes to global warming indirectly. For a typical super-
market in the US, which has an average size of 4200 square
meters, the annual energy consumption is roughly 2 million
kWh, and approximately half of that is for refrigeration
(Westphalen et al., 1996; Zhang, 2006). Thus, the indirect
impact on the environment results from the release of
greenhouse gases (mainly CO2) associatedwith the generation
and transmission of the electrical energy used by the refrig-
eration system.
The direct environmental impact of the refrigeration sys-
tem can be reduced by using refrigerants with lower GWP.
Refrigerants such as R32, R134a, R717, R744, R290, R600a and
R1234yf could be potential alternative refrigerants. However,
due to toxicity and/or flammability, some of these refrigerant
options may not be permissible under various municipal
safety codes. Cascade systems and secondary loop systems
(discussed later in Section 2) using CO2 as a refrigerant can be
used to reduce the direct impact on the environment due to
their lower HFC refrigerant charge.
The indirect environmental impact of the refrigeration
system can be reduced by increasing the energy efficiency or
decreasing the energy consumption of the system. One option
for decreasing energy consumption is to reduce the load on
the refrigeration system. For example, this can be done by
replacing open display cases with doored display cases.
Several studies have shown that doored display cases can
reduce refrigeration system energy consumption by up to 50%
when combined with high efficiency display case components
such as LED lighting, demand defrost, electronically commu-
tated evaporator fan motors and humidity controlled anti-
sweat heaters (Rauss et al., 2008; Fricke and Becker, 2010).
Other energy efficiency measures that can be utilized include
variable speed drives for compressors and condenser fan
motors, as well as floating condensing and suction pressure
controls.
Carbon dioxide has recently received considerable atten-
tion as an alternative to the commonly used synthetic re-
frigerants in supermarket refrigeration systems, in an effort to
develop systems with lower environmental impact (Bansal,
2012; Getu and Bansal, 2008). Although CO2 has a high crit-
ical pressure (7.38 MPa) and a low critical temperature
(30.97 �C), its high operating pressure leads to a high vapor
density and thus a high volumetric refrigerating capacity. The
volumetric refrigerating capacity of CO2 (22,545 kJ m�3 at 0 �C)is 3e10 times larger than CFC, HCFC, HFC and HC refrigerants
(Kim et al., 2004). In addition, carbon dioxide has no Ozone
Depletion Potential (ODP); a GWP of one; and is nontoxic,
nonflammable and inexpensivee all attractive characteristics
when compared to synthetic refrigerants.
Table 1 e Schematics and P-h diagrams of various CO2-based refrigeration systems.
Schematic P-h diagram Features
� R404A is used as refrigerant
in the system
� System has a rack of
compressors and an SLHX
� CO2 low-temperature circuit
� R404A high-temperature circuit
� CO2 receiver
� Liquid CO2 is pumped to
either LT or MT evaporators
� R404A circuit has a rack
of compressors
� CO2 low-temperature circuit
� R404A high-temperature circuit
� Both circuits have an SLHX and
a rack of compressors
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Table 1 e (continued )
Schematic P-h diagram Features
� CO2 low-temperature circuit
� R404A high-temperature circuit
� CO2 receiver
� LT load is cooled with direct
expansion CO2 while MT load is
cooled with pumped liquid CO2
� Both cascade circuits have a
rack of compressors and an SLHX
� CO2 low-temperature circuit
� R404A high-temperature circuit
� CO2 receiver
� LT load is cooled with direct
expansion CO2 while MT load is
cooled with chilled glycol-water
solution
� Both cascade circuits have a
rack of compressors and an SLHX
� CO2 is used as refrigerant
throughout the system
� CO2 receiver
� LT and MT loads are cooled via
direct expansion
� The system has high-pressure
and low-pressure compressors
and two SLHXs
� Refrigerant is expanded after
exiting SLHX1
(continued on next page)
i n t e rn a t i o n a l j o u r n a l o f r e f r i g e r a t i o n 4 6 ( 2 0 1 4 ) 8 6e9 9 89
Table 1 e (continued )
Schematic P-h diagram Features
� CO2 is used as refrigerant
throughout the system
� CO2 receiver
� LT and MT loads are cooled
via direct expansion
� The system has high-pressure
and low-pressure compressors
and two SLHXs
� Refrigerant is expanded before
entering SLHX1
� CO2 is used as refrigerant
throughout the system
� CO2 receiver
� LT and MT loads are cooled
via direct expansion
� The system has high-pressure,
low-pressure and bypass
compressors and two SLHXs
� Bypass refrigerant is
compressed before entering
the Gas Cooler
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Carbon dioxide has successfully been used as a refrigerant
in the low-temperature circuit of cascade systems, in second-
ary loop systems, and in transcritical systems (Bansal, 2012;
Girotto et al., 2004; Hinde and Zha, 2009). However, tran-
scritical CO2 systems tend to be more popular in moderate
climates such as Northern Europe where the refrigeration
system operates a majority of the time in the more efficient
subcriticalmode (Deneckeet al., 2012; SawalhaandPalm, 2003).
In an effort to increase the efficiency of the transcritical CO2
system and to make it applicable to warmer climates, several
researchers have investigated the energy efficiency of various
configurations of the transcritical refrigeration system (Bell,
2004; Ferrandi and Orlandi, 2012; Ge and Tassou, 2009, 2010;
Mazzola et al., 2012; Sarkar and Agrawal, 2010; Sawalha,
2007; Winter and Murin, 2012). However, these studies
focused on a particular system only and lacked any system
performance comparison with various other possible CO2
systemconfigurations. Also, these studies did not optimize the
design parameters such as the evaporator superheat (DT) and
the suction line heat exchanger (SLHX) effectiveness, in order
to maximize the system coefficient of performance (COP). In
this study, a comprehensive analysis of seven CO2-based
refrigeration system configurations that are currently being
used in the supermarket refrigeration industry around the
world is performed. The paper presents a systematic analysis
of each configuration for the same operating conditions. In
addition, the performance of the more energy-efficient CO2-
based refrigeration systems is compared with that of a base-
line R404A multiplex DX system using bin analyses in sixteen
cities from eight climate zones of the United States.
2. Description of CO2 refrigeration systemsevaluated
The CO2-based refrigeration systems investigated in this
study include cascade and secondary loop systems as well as
transcritical systems, as shown in Table 1. These systems are
briefly described in the following section, along with the per-
formance comparison of the CO2-based refrigeration systems
with the baseline R404A multiplex DX system (System 1).
2.1. Secondary loop and cascade systems
A secondary loop system is comprised of two circuits, a pri-
mary DX circuit and a secondary pumped loop circuit, coupled
through a secondary fluid heat exchanger as shown in System
2 of Table 1. The primary DX circuit typically utilizes R404A
while the secondary loop uses a pumped liquid such as pro-
pylene glycol (single-phase) or CO2 (two-phase).
Table 2 e Refrigeration capacity and required power.
System _Qref_Wtotal
DX _Qref ¼ _QLT ¼ _m3ðh4 � h3Þ or _Qref ¼ _QMT ¼ _m3ðh4 � h3Þ _Wtotal ¼ _m5ðh6 � h5Þ;SC _Qref ¼ _QLT ¼ _m10ðh11 � h10Þ or _Qref ¼ _QMT ¼ _m10ðh11 � h10Þ _Wtotal ¼ _Wpump þ _Wcomp
_Wcomp ¼ _m5ðh6 � h5Þ;_Wpump ¼ _m9ðh10 � h9Þ
DEC _Qref ¼ _QLT ¼ _m9ðh10 � h9Þ _Wtotal ¼ _WLP þ _WHP
_WLP ¼ _m11ðh12 � h11Þ; _WHP ¼ _m5ðh6 � h5ÞCombined System 1 (LT SC þ MT SC) _Qref ¼ _QLT;SC þ _QMT;SC
_QLT;SC ¼ ½ _m10ðh11 � h10Þ�LT_QMT;SC ¼ ½ _m10ðh11 � h10Þ�MT
_Wtotal ¼ _Wpump;LT;SC þ _Wpump;MT;SC þ _Wcomp;LT;SC þ _Wcomp;MT;SC
_Wcomp;LT;SC ¼ ½ _m5ðh6 � h5Þ�LT_Wcomp;MT;SC ¼ ½ _m5ðh6 � h5Þ�MT_Wpump;LT;SC ¼ ½ _m9ðh10 � h9Þ�LT_Wpump;MT;SC ¼ ½ _m9ðh10 � h9Þ�MT
Combined System 2 (LT DEC þ MT SC) _Qref ¼ _QLT;DEC þ _QMT;SC_QLT;DEC ¼ ½ _m9ðh10 � h9Þ�LT_QMT;SC ¼ ½ _m10ðh11 � h10Þ�MT
_Wtotal ¼ _Wpump;MT;SC þ _Wcomp;LT;DEC þ _Wcomp;MT;SC
_Wcomp;LT;DEC ¼ ½ _m11ðh12 � h11Þ þ _m5ðh6 � h5Þ�LT_Wcomp;MT;SC ¼ ½ _m5ðh6 � h5Þ�MT_Wpump;MT;SC ¼ ½ _m9ðh10 � h9Þ�MT
CSC _Qref ¼ _QLT þ _QMT_QLT ¼ _m14ðh15 � h14Þ; _QMT ¼ _m12ðh13 � h12Þ
_Wtotal ¼ _WLP þ _WHP þ _Wpump
_WLP ¼ _m16ðh17 � h16Þ; _WHP ¼ _m5ðh6 � h5Þ_Wpump ¼ _m9ðh10 � h9Þ
CSC-G _Qref ¼ _QLT þ _QMT_QLT ¼ _m14ðh15 � h14Þ; _QMT ¼ _m19ðh20 � h19Þ
_Wtotal ¼ _WLP þ _WHP þ _Wpump
_WLP ¼ _m16ðh17 � h16Þ; _WHP ¼ _m10ðh11 � h10Þ_Wpump ¼ _m18ðh19 � h18Þ
STBS _Qref ¼ _QLT þ _QMT_QLT ¼ _m10ðh11 � h10Þ; _QMT ¼ _m7ðh8 � h7Þ
_Wtotal ¼ _WLP þ _WHP
_WLP ¼ _m12ðh13 � h12Þ; _WHP ¼ _m16ðh17 � h16ÞTBS-UX _Qref ¼ _QLT þ _QMT
_QLT ¼ _m10ðh11 � h10Þ; _QMT ¼ _m7ðh8 � h7Þ_Wtotal ¼ _WLP þ _WHP
_WLP ¼ _m12ðh13 � h12Þ; _WHP ¼ _m16ðh17 � h16ÞTBS-BC _Qref ¼ _QLT þ _QMT
_QLT ¼ _m10ðh11 � h10Þ; _QMT ¼ _m7ðh8 � h7Þ_WTotal ¼ _WLP þ _WHP þ _WBP
_WLP ¼ _m12ðh13 � h12Þ; _WHP ¼ _m15ðh16 � h15Þand_WBP ¼ _m4ðh5 � h4Þ
internatio
naljo
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n46
(2014)86e99
91
Table 3 e Parameters and their ranges.
Parameter Range
Tamb 0e40 �CPump Circulation Ratio (CR) 1.5, 2.5
Heat Exchanger Effectiveness (εSLHX) 0, 0.4, 0.7
Medium-Temperature Superheat (DTMT) 10, 15 K
Low-Temperature Superheat (DTLT) 10, 15 K
Fig. 1 e COP and work ratio vs. ambient temperature with
CR ¼ 1.5 for Combined System 1.
i n t e r n a t i o n a l j o u r n a l o f r e f r i g e r a t i o n 4 6 ( 2 0 1 4 ) 8 6e9 992
A cascade system is comprised of separate high-
temperature and low-temperature circuits, coupled through
a heat exchanger called the cascade condenser. The cascade
condenser functions as an evaporator for the high-
temperature circuit and a condenser for the low-
temperature circuit. Generally, the high-temperature circuit
is a single-stage direct expansion system but the low-
temperature circuit can either be a direct expansion system
or a secondary loop system. For a cascade system with high-
and low-temperature DX circuits, one or both circuits may
have a SLHX as shown in System 3 of Table 1. The refrigerant
in the high temperature circuit is typically an HFC (R404A in
this case), while CO2 is used in the low-temperature circuit.
2.1.1. CO2 Secondary Coolant (SC) SystemIn the SC system (System 2, Table 1), a high temperature
R404A circuit provides cooling to a low-temperature pumped
CO2 secondary loop circuit. The pumped CO2 circuit in turn,
provides cooling for the medium- or low-temperature loads.
In the pressure-enthalpy (P-h) diagram, the DX R404A (1-2-3-4-
5-6) and the pumped CO2 (7-8-9-10-11) circuits are represented
by the inner and the outer domes, respectively.
Refrigeration systems with separate SC systems for the
low- and medium-temperature loads are commonly used in
supermarkets. In this paper, such a paired SC system is called
the Combined System 1. This systemwould be represented by
two separate System 2 cycles as shown in Table 1.
2.1.2. CO2 direct expansion cascade (DEC) systemIn the DEC system (System 3, Table 1), the low-temperature
circuit and the high-temperature circuit are single-stage
Table 4 e Baseline parameters of refrigeration systems.
Baseline parameters
Refrigeration Load
Baseline System Condensing Temperature
Cascade Systems Condensing Temperature
Cascade Condenser Approach Temperatu
Pump Power, Glycol
Pump Power, CO2
Booster Systems Intermediate Pressure in Receiver
Condenser/Gas Cooler Outlet Temperatur
Condenser/Gas Cooler Pressure (for Tamb �
direct expansion systems, each having their own SLHXs (if
necessary), coupled through the cascade condenser. In the P-h
diagram, the DX R404A (1-2-3-4-5-6) and the DX CO2 (7-8-9-10-
11-12) circuits are represented by the inner and the outer
domes, respectively.
A refrigeration system using a DEC system for the low-
temperature (LT) loads and a separate SC system for the
medium-temperature (MT) loads is also studied in this paper.
Such a refrigeration system will be called the Combined Sys-
tem 2. This system would be represented by two separate cy-
cles; System 2 for MT and System 3 for LT, as shown in Table 1.
2.1.3. Combined CO2 secondary/cascade (CSC) systemIn the CSC system (System 4, Table 1), the low-temperature
and medium-temperature loads are managed by a single-
stage direct expansion system and a CO2 secondary loop
system, respectively, coupled with the high-temperature cir-
cuit through the cascade condenser. In the P-h diagram, the
DX R404A (1-2-3-4-5-6) and the CO2 (7-8-9-10-11-12-13-14-15-
16-17-18) circuits are represented by the inner and the outer
domes, respectively.
Value
_QMT ¼ 120 kW at �5 �C_QLT ¼ 65 kW at �30 �CTcond ¼ 21 �C for Tamb � 8 �CTcond ¼ Tamb þ 10 �C for Tamb > 8 �CTcond ¼ 21 �C for Tamb � 8 �CTcond ¼ Tamb þ 10 �C for Tamb > 8 �C
re Tapp ¼ 3.3 �C_Wpump ¼ 0.05 _Wcomp
_Wpump ¼ 0.01 _Wcomp
Pint ¼ 3.5 MPa
e TGC ¼ 10 �C for Tamb � 0 �CTGC ¼ Tamb þ10 �C for 0 �C < Tamb < 18 �CTGC ¼ 28 at P ¼ 6.9 MPa for 18 �C � Tamb < 22 �CTGC ¼ 28 at P ¼ 7.5 MPa for 22 �C � Tamb < 25 �CTGC ¼ Tamb þ 3 �C for Tamb � 25 �C
27 �C) PSTBS ¼ 0:00254*T2amb þ 0:10003*Tamb þ 2:9211
PTBS-UX ¼ 0:0021*T2amb þ 0:13516*Tamb þ 2:2487
PTBS-BC ¼ 0:0004*T2amb þ 0:25237*Tamb þ 0:362
Fig. 2 e COP vs. ambient temperature with DTLT ¼ 10 K for
the CSC system (System 4).
i n t e rn a t i o n a l j o u r n a l o f r e f r i g e r a t i o n 4 6 ( 2 0 1 4 ) 8 6e9 9 93
2.1.4. Combined glycol secondary/CO2 cascade (CSC-G)systemIn the CSC-G System (System 5, Table 1), the low-temperature
loads are served by a single-stage direct expansionCO2 system
coupled with the high-temperature circuit through the
cascade condenser, while the medium-temperature loads are
managed through a propylene glycol secondary loop (instead
of liquid CO2 as in System 4). In the P-h diagram, the DX R404A
(1-2-3-4-5-6-7-8-9-10-11) and the CO2 (12-13-14-15-16-17) cir-
cuits are represented by the inner and the outer domes,
respectively.
2.2. Transcritical booster (CO2) systems
A transcritical booster system (Systems 6, 7 and 8 in Table 1) is
divided into four pressure levels. At the highest pressure level,
the refrigerant (CO2) from the high pressure compressors en-
ters a gas cooler/condenser and rejects heat to the surround-
ings. The high pressure CO2 is further cooled in a suction line
heat exchanger (SLHX1) and expanded before being collected
in the receiver. Depending upon the system configuration,
refrigerant is expanded to an intermediate pressure level
either before or after passing through the SLHX1. The satu-
rated refrigerant liquid from the receiver is expanded to the
Fig. 3 e Performance of the STBS system (System 6) for ƐSLHX ¼ 0
flow ratios.
saturation pressure corresponding to the medium-
temperature and low-temperature loads, respectively. After
absorbing heat from the low-temperature loads, the low-
pressure refrigerant is further superheated in the SLHX2 and
compressed in the low-pressure compressors. The discharge
from the low-pressure compressors, the medium-
temperature loads and the bypass valve combines before
entering the high-pressure compressors.
2.2.1. Standard transcritical booster system (STBS)In the STBS (System 6, Table 1), the high pressure refrigerant
from the outlet of the gas cooler/condenser passes through
the SLHX1 before expanding to an intermediate pressure level
through the expansion device between locations 2 and 3. In
the P-h diagram, state points 1, 2 and 17 are at the high pres-
sure level, state points 3, 4, 6 and 9 are at the intermediate
pressure level, state points 5, 7, 8, 13, 14, 15 and 16 are at the
medium-temperature level and state points 10, 11 and 12 are
at the low-temperature level.
2.2.2. Transcritical booster system with upstream expansionvalve (TBS-UX)The TBS-UX (System 7 in Table 1) differs from the STBS in that
the high pressure refrigerant from the gas cooler is expanded
between 2 and 3, prior to passing through SLHX1 and stored in
the receiver. In the P-h diagram, state points 1 and 17 are at the
high pressure level, state points 2, 3, 4, 6 and 9 are at the in-
termediate pressure level, state points 5, 7, 8, 13, 14, 15 and 16
are at themediumpressure level and state points 10, 11 and 12
are at the low pressure level.
2.2.3. Transcritical booster system with bypass compressor(TBS-BC)In the TBS-BC (System 8, Table 1), is similar to the STBS,
however, the refrigerant exiting the receiver bypass at loca-
tion 4 is compressed by an additional set of compressors
(bypass compressors), which then combines with the
discharge of the high-pressure compressors at location 17,
before finally entering the gas cooler/condenser. In the P-h
diagram, state points 1, 2, 5, 16 and 17 are at the high pressure
level, state points 3, 4, 6 and 9 are at the intermediate pressure
level, state points 7, 8, 13, 14, and 15 are at the medium
.4 and DT ¼ 10 K. (a) System COP and work ratios, (b) Mass
Fig. 4 e Bypass mass flow ratio vs. ambient temperature
with DT ¼ 10 K for the STBS (System 6).
i n t e r n a t i o n a l j o u r n a l o f r e f r i g e r a t i o n 4 6 ( 2 0 1 4 ) 8 6e9 994
pressure level and state points 10, 11 and 12 are at the low
pressure level.
3. System analysis
The overall coefficient of performance, COP, of a refrigeration
system is determined by:
COP ¼_Qref
_Wtotal
(1)
where _Qref is the total refrigerating capacity of the system and_Wtotal is the total power required by the compressors and
pumps (the condenser/gas cooler fan power is excluded in
these analyses). For the eight refrigeration systems shown in
Table 1, the corresponding equations for refrigerating capacity
and total power input are given in Table 2. These equations
were used with the definition of COP (Eq. 1) to determine the
performance of each of the eight refrigeration systems.
Note that for booster and cascade refrigeration systems, it
is difficult to separate cooling energy and work input into
medium-temperature and low-temperature contributions.
Thus an overall COP was used in this study to quantify the
performance of the various refrigeration systems.
Fig. 5 e System COP vs. ambient temperature at ƐSLHX ¼ 0.4, DT
Cascade/secondary systems.
4. Methodology
4.1. Baseline parameters
In order to assess themerits and the advantages offered by the
CO2-based refrigeration systems shown in Table 1, a para-
metric studywas performed to determine the effect of various
operating parameters on the energy efficiency of the systems.
Ambient temperature ranging between 0 �C and 40 �C was
used to determine the performance of the systems in all eight
climate zones of the US. SLHX effectiveness values of either
0 or 0.7 were used to represent a systemwithout a SLHX or the
upper range of a realistic heat exchanger (where 1.0 repre-
sents an ideal 100% efficient heat exchanger), respectively.
Evaporator superheat is always required to ensure that no
liquid enters the compressor. Thus, the performance of the
systems was studied using an evaporator superheat of either
10 K or 15 K for both the low- andmedium-temperature loads.
A summary of the parameters used in the analysis is given in
Table 3. The thermodynamic properties of the refrigerants
were calculated using the National Institute of Standards and
Technology (NIST) Reference Fluid Thermodynamic and
Transport Properties Database (REFPROP) Version 8.0
(Lemmon et al., 2007).
4.2. System operating assumptions
For an average supermarket (4200 m2), the typical refrigera-
tion loads and temperatures of the low-temperature and the
medium-temperature loads were assumed to be 65 kW at
�30 �C and 120 kW at �5 �C, respectively. It has been noted
that for typical supermarkets, the MT load is approximately
two to five times more than that of the LT load (Girotto et al.,
2004; Getu and Bansal, 2008). In this study, it was assumed
that the MT load is twice that of the LT load.
For the proper operation of the expansion valves and effi-
cient heat transfer through the condenser in the DX and
cascade system options, the minimum condensing tempera-
ture was fixed at Tcond ¼ 21 �C and the condensing tempera-
ture was set to be 10 �C greater than the ambient temperature
(Tamb). In addition, no liquid sub-cooling was assumed at the
exit of condenser. The approach temperature, which is the
¼ 10 K and CR ¼ 1.5. (a) Transcritical booster systems, (b)
Table 6 e Impact of increasing ƐSLHX and DT on thesystem COPs.
System ƐSLHX DT
Combined System 1 Increase NA
Combined System 2 Increase Negligible*
CSC Decrease for Tamb < 15 �C Negligible*
Increase for Tamb > 15 �CCSC-G Decrease Negligible*
STBS Negligible* Negligible*
TBS-UX Negligible* Negligible*
TBS-BC Negligible* Negligible*
* Less than 1%.
Fig. 6 e System COP vs. ambient temperature at
ƐSLHX ¼ 0.4, DT ¼ 10 K and CR ¼ 1.5 for R404A DX System
(System 1), CSC (System 4), STBS (System 6), and TBS-BC
(System 8).
i n t e rn a t i o n a l j o u r n a l o f r e f r i g e r a t i o n 4 6 ( 2 0 1 4 ) 8 6e9 9 95
difference between the temperature at the outlet of the low-
temperature circuit and the inlet of the high-temperature
circuit of the cascade condenser, was Tapp ¼ 3.3 �C.Generally, in transcritical booster systems, the receiver
pressure is approximately 0.5 MPa higher than the saturated
pressure of the medium-temperature loads, and thus, the
intermediate pressure in the receiver was fixed at
Pint¼ 3.5MPa for this study. In addition, since transcritical CO2
booster systems operate at higher pressures in comparison to
the HFC based DX and cascade system options, the expansion
valves can function properly at lower ambient temperatures.
Thus, the minimum condensing temperature of the tran-
scritical CO2 booster systems can be lower than that of the
HFC-based DX and cascade system options. In this study it
was assumed that the gas cooler/condenser temperature (TGC)
was 10 �C greater than the ambient temperature during
subcritical operation. In the intermediate stage,
18 �C < Tamb < 22 �C, TGC and condensing pressure were fixed
to 28 �C and 6.9 MPa; 22 �C < Tamb < 25 �C, TGC and condensing
pressure were fixed to 28 �C and 7.5 MPa, respectively. In
Table 5 e US climate zones, cities and annual average COP for
Climate zone City Annual average temperature
1 Miami, FL 24.9
2A Houston, TX 20.7
2B Phoenix, AZ 23.8
3A Atlanta, GA 17.0
3B Los Angeles, CA 17.3
3B Las Vegas, NV 20.2
3C San Francisco, CA 14.4
4A Baltimore, MD 13.3
4B Albuquerque, NM 14.2
4C Seattle, WA 11.4
5A Chicago, IL 10.0
5B Boulder, CO 10.3
6A Minneapolis, MN 8.0
6B Helena, MN 7.2
7 Duluth, MN 4.3
8 Fairbanks, AK �2.1
supercritical operation, Tamb � 27 �C, TGC is set to be 3 �Cgreater than the ambient temperature. Also, in supercritical
operation, an optimum gas cooler pressure exists which
maximizes the COP of the system (Kauf, 1999). These param-
eters, along with the calculated optimum pressure equations
for each transcritical booster system as functions of ambient
temperature, are given in Table 4. In addition, the following
assumptions were made for the parametric analysis:
� The pressure drop and heat loss/gain in the suction lines
were ignored.
� Only saturated liquid and saturated vapor exit the receiver.
� The isentropic efficiency of the low-pressure, high-pres-
sure and bypass compressors was 0.65.
� The expansion valves were isenthalpic.
� Condenser/gas cooler fan power consumption was
assumed to be roughly equal for all the systems, so fan
power was not included in the COP computations.
5. Results and discussion
Altogether, seven CO2-based refrigeration system configura-
tions were analyzed and compared, including the Combined
System 1 (MT system 2, LT System 2), Combined System 2 (MT
several CO2-based refrigeration systems.
(�C) Annual average COP
R404A DX CSC and combined system 2 TBS-BC
2.60 2.54 2.36
2.90 2.80 2.68
2.70 2.62 2.40
3.14 3.02 3.04
3.22 3.10 2.75
2.96 2.86 2.78
3.51 3.35 3.09
3.32 3.18 3.44
3.31 3.17 3.40
3.60 3.42 3.52
3.42 3.27 3.78
3.46 3.30 3.75
3.47 3.31 3.98
3.58 3.41 4.08
3.63 3.46 4.28
3.70 3.52 4.54
Fig. 7 e Comparative advantage of a specific system in various climate zones of the US.
i n t e r n a t i o n a l j o u r n a l o f r e f r i g e r a t i o n 4 6 ( 2 0 1 4 ) 8 6e9 996
System 2, LT System 3), CSC system (System 4), CSC-G system
(System 5), STBS (System 6), TBS-UX (System 7) and TBS-BC
(System 8). The effects of ambient temperature, evaporator
superheat and SLHX effectiveness on the performance of each
of the seven configurations were investigated. In addition, the
performance of the seven CO2-based refrigeration systems
was compared to the baseline R404A multiplex DX system
(System 1).
5.1. Cascade and secondary loop systems
In general for the cascade and secondary loop systems, an
increase in the ambient temperature leads to an increase in
the work required by the high-temperature circuit with no
impact on the work required by the low-temperature circuit.
This is due to the fact that the pressure ratio of high-
temperature circuit compressors changes with change in the
ambient temperature, while the low-temperature circuit
compressors and pumps are decoupled from the ambient
conditions through the cascade condenser. Thus, the system
COP decreases with an increase in the ambient temperature.
5.1.1. Combined System 1 (MT System 2, LT System 2)Fig. 1 shows the system COP and the work ratio (WR) vs.
ambient temperature for a given pump circulation ratio of 1.5
for the Combined System 1 (CO2 Secondary Coolant System
for low- and medium-temperature applications, two separate
System 2 cycles as shown in Table 1). The work ratio of the
compressors is defined to be the ratio of the compressor
power, _W, to the total power required by the system,_W þ _Wpump, as follows:
WR ¼_W
_W þ _Wpump
(2)
Since the SLHX is in the high-temperature circuit, it has no
impact on the performance of the low-temperature circuit.
However, with an increase in the effectiveness of the SLHX,
thework required by the compressors decreases, as illustrated
by a decrease in the compressor work ratio shown in Fig. 1,
resulting in an increase in the system COP. For example, at an
ambient temperature of 20 �C, the system COP increases by
5.8% with an increase in SLHX effectiveness from 0 to 0.7.
5.1.2. Combined System 2 (MT System 2, LT System 3)In the Combined System 2 (DEC for low-temperature (System
3), SC for medium-temperature (System 2)), an increase in the
effectiveness of the SLHX in the high-temperature circuit re-
sults in a decrease in the mass flow rate which dominates the
increase in enthalpy change occurring across the high-
pressure compressors. Thus, the power required by the
high-pressure compressors decreases. However, an increase
in the effectiveness of the SLHX in the low-temperature circuit
results in an increase in the enthalpy change across the low-
pressure compressors which is greater than the decrease in
the mass flow rate across the compressor. This leads to in-
crease in the power required by the low-pressure compres-
sors. Thus, the impact of SLHX effectiveness on the high-
temperature and low-temperature circuits counteracts each
other. The result is an increase in COP, but not as great as that
of Combined System 1. For example, at an ambient tempera-
ture of 20 �C, the system COP increases by 2%with an increase
in SLHX effectiveness from 0 to 0.7. Finally, the effect of
increasing the LT superheat by 5 K from 10 K to 15 K is insig-
nificant, resulting in a decrease in COP of only 0.02%.
5.1.3. Combined CO2 secondary/cascade (CSC)For the Combined CO2 Secondary/Cascade System (CSC, Sys-
tem 4, Table 1), Fig. 2 shows the system COP vs. ambient
temperature for a LT evaporator superheat of 10 K.
With an increase in the effectiveness of SLHX1 and SLHX2,
the mass flow through the system decreases and the enthalpy
at the inlet of compressors increases. In comparison to the
rate of increase in enthalpy change across the compressors,
the rate of decrease in the refrigerant mass flow rate is greater
in the high-temperature circuit than in the low-temperature
circuit. This leads to an increase in the work required by the
i n t e rn a t i o n a l j o u r n a l o f r e f r i g e r a t i o n 4 6 ( 2 0 1 4 ) 8 6e9 9 97
low-pressure compressors and a decrease in the work
required by the high-pressure compressors.
As the SLHX effectiveness increases at low ambient tem-
perature, the increase in work required by the low-pressure
compressors dominates the decrease in that of the high-
pressure compressors, thereby, increasing the total work
required by the system and decreasing the system COP.
However, as the SLHX effectiveness increases at high ambient
temperatures, the total work required by the system de-
creases and the COP increases. At an ambient temperature of
5 �C, an increase in effectiveness of the SLHX from 0 to 0.7
leads to the system COP decreasing by about 1.3%. On the
other hand, at an ambient temperature of 20 �C and 40 �C, with
change in the effectiveness of the SLHX from 0 to 0.7, the
system COP increases by 0.9% and 9.7%, respectively. The ef-
fect of increasing the LT superheat by 5 K, from 10 K to 15 K, is
insignificant, resulting in a decrease in COP of only 0.01%.
5.1.4. Combined glycol secondary/CO2 cascade system (CSC-G)As with the CSC System, low-temperature evaporator super-
heat has a negligible impact on the COP of the CSC-G System
(System 5, Table 1). However, the work required by the low-
temperature circuit and high-temperature circuit compres-
sors of the CSC-G System increases with an increase in the
effectiveness of all the suction line heat exchangers, leading
to a decrease in the systemCOP. At an ambient temperature of
20 �C, an increase in effectiveness of the SLHX from 0 to 0.7
leads to a 12.7% decrease in the system COP.
5.2. Transcritical Booster Systems
For the Standard Transcritical Booster System (STBS), Fig. 3
shows the system COP, the normalized work ratios (WRLP,
WRHP) and the normalized mass flow ratios (MRBP), vs.
ambient temperature at constant SLHX effectiveness
(ƐSLHX ¼ 0.4). The normalized work ratio is defined as the ratio
of the work required by either the low-pressure or high-
pressure compressor to the work required by low-pressure
compressor.
WRcomp ¼ _wcomp
_wLP(3)
where WRcomp is either WRLP or WRHP, and _wcomp is either _wLP
or _wHP, respectively. The normalizedmass flow ratio is defined
as the ratio of either the low-temperature, medium-temper-
ature, or bypass mass flow rate to the mass flow rate through
the low-temperature circuit.
MRcirc ¼_mcirc
_mLT(4)
where MRcirc is either MRLT, MRMT or MRBP, and mcirc is either
mLT, mMT or mBP, respectively.
Since the loads in the low-temperature circuit and the
medium-temperature circuit are fixed in this analysis, the
refrigerant flow rate through the LT and MT circuits are con-
stant, and ambient temperature has no impact on these flow
rates. However, an increase in ambient temperature leads to
an increase in vapor entering the receiver. As a result, the
bypass flow rate and consequently, the bypass flow ratio
(MRBP) increases, as shown in Fig. 3(b).
The bypass mass flow ratio is constant for
18 �C � Tamb < 22 �C and 22 �C � Tamb � 25 �C due to fixed
condensing pressure of 6.9 MPa and 7.5 MPa, respectively,
during the transition from subcritical to transcritical opera-
tion. The change in bypassmass flow ratio has a direct impact
on the work required by the high-pressure compressors. This
can be seen in Fig. 3(a) where the work ratio of the high-
pressure compressor increases, and consequently, the sys-
tem COP decreases.
An increase in evaporator superheat results in a decrease
in the total refrigerant mass flow rate. In general, increasing
the evaporator superheat leads to an increase in the refrig-
erant enthalpy throughout the system between the low-
pressure compressor and the gas cooler inlet. However, the
increase in enthalpy associated with superheat is nearly
counterbalanced by the decrease in the total refrigerant flow
rate. As a result, the rate of change of COP with an increase in
superheat is insignificant (<0.5%). Similarly, in the Tran-
scritical Booster Systemwith UpstreamExpansion Valve (TBS-
UX) and the Transcritical Booster System with Bypass
Compressor (TBS-BC), the evaporator superheat has insignif-
icant impact on the system COP.
Fig. 4 shows the bypass mass flow ratio (MRBP) vs. ambient
temperature for a given LT and MT evaporator superheat of
10 K for the STBS (System 6). With an increase in the effec-
tiveness of SLHX1 and SLHX2, the quality of refrigerant at the
inlet of both the receiver and the low-temperature evaporator
decreases, thereby decreasing the bypass mass flow ratio
(MRBP) and the low-temperature mass flow ratio (MRLT).
In addition, for the STBS (System 6), an increase in the
SLHX effectiveness increases the enthalpy of the refrigerant at
the inlet of the low- and high-pressure compressors. In
subcritical operation, the rate of increase in the enthalpy
difference across the compressors and the rate of decrease in
the mass flow through the low-temperature evaporator and
the bypass, leads to changes in the relative work input to the
low-pressure and high-pressure compressors. This results in a
decrease in the system COP during subcritical operation but
an increase during transcritical operation. For example, as the
SLHX effectiveness increases from 0 to 0.7, the system COP
decreases by 1.4% at an ambient temperature of 10 �C; how-
ever, the COP increases by 1.9% at an ambient temperature of
35 �C. Similarly, in the TBS-UX (System 7) and the TBS-BC
(System 8), as the SLHX effectiveness increases from 0 to 0.7,
the systemCOP decreases by 1.2% and 1.3%, respectively, at an
ambient temperature of 10 �C, whereas the system COP in-
creases by 0.2% and 1.9%, respectively, at an ambient tem-
perature of 35 �C.
5.3. Comparison of CO2-based refrigeration systemswith the baseline system
Fig. 5 shows the system COP vs. ambient temperature for all
the Transcritical Booster Systems and the Combined Cascade/
Secondary Loop Systems for a constant SLHX effectiveness
(ƐSLHX ¼ 0.4) and evaporator superheat (DT ¼ 10 K). It can be
seen that the Transcritical Booster System with Bypass
Compressor (TBS-BC) and both the Combined System 2 and
i n t e r n a t i o n a l j o u r n a l o f r e f r i g e r a t i o n 4 6 ( 2 0 1 4 ) 8 6e9 998
the Combined CO2 Secondary/Cascade (CSC) System are the
most efficient systems among the transcritical booster and
the cascade/secondary loop systems, respectively.
Fig. 6 shows the system COP vs. ambient temperature for a
baseline system consisting of separate R404A multiplex DX
circuits for the MT and LT loads, the TBS-BC (System 8), the
STBS (System 6), the Combined System 2, and the Combined
CO2 Secondary/Cascade (CSC) (System 4), for a constant SLHX
effectiveness (ƐSLHX ¼ 0.4) and evaporator superheat
(DT ¼ 10 K). At ambient temperatures between 0 �C and 40 �C,the COP of the R404A multiplex DX system is higher than the
COP of the Combined System 2 and the Combined CO2 Sec-
ondary/Cascade (CSC) System by up to 6.8%, with an average
difference in COP of approximately 5.8%.
At ambient temperature less than 8 �C, the COP of the TBS-
BC (System 8) is higher than the COP of R404A multiplex DX
system (System 1), but at higher temperatures, the COP of the
R404A multiplex DX system is higher than the COP of TBS-BC
(System 8). For example, at an ambient temperature of 5 �C,the system COP of TBS-BC (System 8) is 14.3% higher than that
of R404A multiplex DX system (System 1). Whereas, at an
ambient temperature of 35 �C, the system COP of R404A
multiplex DX system (System 1) is 15% higher than TBS-BC
(System 8). The STBS behaves similar to the TBS-BC, howev-
er, over the temperature range of 0 �Ce40 �C, the average COP
of the TBS-BC is 3.3% greater than that of the STBS. Also, the
average COP's over the range 0 �Ce40 �C for the R404A multi-
plex DX system (System 1) and the TBS-BC (System 8) are
nearly equal.
5.4. Climate zones
Bin analyses were performed to determine the annual average
COP of the TBS-BC (System 8), the Combined System 2, Com-
bined CO2 Secondary/Cascade (CSC) System and the R404A
multiplex DX system (System 1) in sixteen cities selected from
the eight climate zones of the United States (ICC, 2009). The 16
cities selected for the bin analyses are shown in Table 5, along
with their annual average temperature and annual average
COP of the four systems. The hourly weather data from these
sixteen cities were used to determine the annual average COP
for the four refrigeration systems. It can be seen that the TBS-
BC (System 8) is the most efficient system in Zones 5, 6, 7 and
8. In these zones, the annual average COP of the TBS-BC
(System 8) is higher than that of the R404A DX system (Sys-
tem 1) by up to 13%. In Zones 1, 2 and 3, the R404A DX system
performs the best. The annual average COP of the R404A DX
system (System 1) is up to 8.3% higher than that of the TBS-BC.
However, in zone 4, the R404A DX system (System 1) and the
TBS-BC (System 8) perform similarly, within 4% of each other.
Also, in all climate zones, the average performance of the
Combined System 2 and the CSC system are 4.2% lower than
that of the R404A DX system.
6. Conclusions
The comparative analysis of the CO2-based refrigeration sys-
tems with the baseline R404A system has revealed the
following conclusions:
The qualitative impact of an increase in the SLHX effec-
tiveness (ƐSLHX) and evaporator superheat (DT) on the perfor-
mance of the transcritical booster systems and the combined
cascade/secondary loop systems is summarized in Table 6.
Generally, the evaporator superheat has negligible impact on
the performance of the CO2-based refrigeration systems. An
increase in SLHX effectiveness has negligible impact on the
performance of the transcritical booster systems. However, an
increase in SLHX effectiveness leads to an increase in the
performance of the Combined Systems 1 and 2 and a decrease
in the performance of the CSC-G (System 5). Also, an increase
in the SLHX effectiveness in the CSC (System 4) leads to an
increase in the system performance for Tamb < 15 �C but a
decrease in performance for Tamb > 15 �C.The TBS-BC and both the Combined System 2 and the
Combined CO2 Secondary/Cascade System (CSC) are the most
efficient systems among the transcritical booster and cascade/
secondary loop systems, respectively. Fig. 7 shows the refrig-
eration systems which perform most efficiently for each
climate zone in the US. In Zones 5, 6 and 7, shown in the green
shaded regions on the map, the TBS-BC is the most efficient
system. In Zones 1, 2, and 3, shown in the red regions, the
R404A DX system (System 1) is most efficient. Finally, in Zone
4, shown in yellow, the R404A DX system (System 1) and the
TBS-BC (System 8) perform similarly.
Implementation of transcritical booster systems or
cascade/secondary loop systems using optimized operating
conditions will lead to reduced direct greenhouse gas emis-
sions while achieving comparable energy consumption as
compared to current HFC-based multiplex DX systems.
r e f e r e n c e s
Bansal, P.K., 2012. A review e status of CO2 as a low temperaturerefrigerant: fundamentals and R&D opportunities. Appl.Therm. Eng. 41, 18e29.
Bell, I., 2004. Performance increase of carbon dioxide refrigerationcycle with the addition of parallel compressioneconomization. In: Proceedings of the 6th IIR GustavLorentzen Conference on Natural Working Fluids, Glasgow,UK, 29 August e 1 September 2004.
Denecke, J., Hafner, A., Eikevik, T., Ladam, Y., 2012. Heat recoverysolutions for R744 booster commercial refrigeration systems.In: Proceedings of the 10th IIR Gustav Lorentzen Conferenceon Natural Working Fluids, Delft, The Netherlands, 25e27 June2012.
Ferrandi, C., Orlandi, M., 2012. Theoretical analysis of cold storagedevice effects on the performance and regulation of a CO2
supermarket refrigeration plant. In: Proceedings of the 10th IIRGustav Lorentzen Conference on Natural Working Fluids,Delft, The Netherlands, 25e27 June 2012.
Fricke, B.A., Becker, B.R., 2010. Energy use of doored and openvertical refrigerated display cases. In: Proceedings of the 13thInternational Refrigeration and Air Conditioning Conference.Purdue University, West Lafayette, IN, 12e15 July 2010.
Ge, Y.T., Tassou, S.A., 2009. Control optimization of CO2 cycles formedium-temperature retail food refrigeration systems. Int. J.Refrigeration 32, 1376e1388.
Ge, Y.T., Tassou, S.A., 2010. Performance evaluation and controloptimization of a CO2 booster refrigeration system insupermarket. In: Proceedings of the 1st IIR International
i n t e rn a t i o n a l j o u r n a l o f r e f r i g e r a t i o n 4 6 ( 2 0 1 4 ) 8 6e9 9 99
Conference on Sustainability and the Cold Chain, Cambridge,UK, 29e31 March 2010.
Getu, H.M., Bansal, P.K., 2008. Thermodynamic analysis of anR744-R717 cascade refrigeration system. Int. J. Refrigeration31, 45e54.
Girotto, S., Minetto, S., Neksa, P., 2004. Commercial refrigerationsystem using CO2 as the refrigerant. Int. J. Refrigeration 27,717e723.
Hinde, D., Zha, S., 2009. Natural refrigerant applications in NorthAmerican supermarkets. In: Proceedings of the IIAR IndustrialRefrigeration Conference and Exhibition, Dallas, TX, 22e25March 2009.
ICC, 2009. International Energy Conservation Code. InternationalCode Council, Washington, D.C.
ICF Consulting, 2005. Revised Draft Analysis of U.S. CommercialSupermarket Refrigeration Systems.
Kauf, F., 1999. Determination of the optimum high pressure fortranscritical CO2 refrigeration cycles. Int. J. Therm. Sci. 38,325e330.
Kim, M.H., Pettersen, J., Bullard, C.W., 2004. Fundamental processand system design issues in CO2 vapor compression systems.Prog. Energy Combust. Sci. 30, 119e174.
Lemmon, E., McLinden, M., Huber, M., 2007. NIST StandardReference Database 23: Reference Fluid Thermodynamic andTransport Properties-REFPROP, Version 8.0, StandardReference Data Program. National Institute of Standards andTechnology, Gaithersburg, MD.
Mazzola, D., Toffolo, A., Orlandi, M., 2012. Supermarketapplication. CO2 system with groundwater sink. Modelsimulation. In: Proceedings of the 10th IIR Gustav Lorentzen
Conference on Natural Working Fluids, Delft, TheNetherlands, 25e27 June 2012.
Rauss, D., Mitchell, S., Faramarzi, R., 2008. Cool retrofit solutionsin refrigerated display cases. In: Proceedings of the 2008 ACEEESummer Study on Energy Efficiency in Buildings, Pacific Grove,CA, 17e22 August 2008.
Sarkar, J., Agrawal, N., 2010. Performance optimization oftranscritical CO2 cycle with parallel compressioneconomization. Int. J. Therm. Sci. 49, 838e843.
Sawalha, S., 2007. Theoretical evaluation of trans-critical CO2
systems in supermarket refrigeration. Part 1: modeling,simulation and optimization of two system solutions. Int. J.Refrigeration 31, 516e524.
Sawalha, S., Palm, B., 2003. Energy consumption evaluation ofindirect systems with CO2 as secondary refrigerant insupermarket refrigeration. In: Proceedings of the 21st IIRInternational Congress of Refrigeration, Washington, D.C.,17e22 August 2003.
Westphalen, D., Zogg, R.A., Varone, A.F., Foran, M.A., 1996. Energysavings Potential for Commercial Refrigeration Equipment.Building Equipment Division, Office of Building Technologies,U.S. Department of Energy, Washington, D.C.
Winter, J., Murin, S., 2012. Energy saving and increasing reliabilityat CO2 transcritical boosters. A case study. In: Proceedings ofthe 10th IIR Gustav Lorentzen Conference on Natural WorkingFluids, Delft, The Netherlands, 25e27 June 2012.
Zhang, M., 2006. Energy analysis of various supermarketrefrigeration systems. In: Proceedings of the 11thInternational Refrigeration and Air Conditioning Conference.Purdue University, West Lafayette, IN, 17e20 July 2006.