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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

Available o

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journal homepage: www.elsevier .com/locate/ i j refr ig

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: frickeba@ornl.gov (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

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 988

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

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 990

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

urnalofrefrig

eratio

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.

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