Field measurements and simulations of supermarkets with CO .../Menu/general/column... · Master of Science Thesis EGI 2009/ETT:490 Field measurements and simulations of supermarkets

Field measurements and

simulations of supermarkets

with CO2 refrigeration systems

D A V I D F R E L E C H O X

Master of Science Thesis

Stockholm, Sweden 2009

Field measurements and simulations of supermarkets

with CO2 refrigeration systems

David Freléchox

Master of Science Thesis Energy Technology 2009:490

KTH School of Industrial Engineering and Management Division of Applied Thermodynamics and Refrigeration

SE-100 44 STOCKHOLM

Master of Science Thesis EGI 2009/ETT:490

Field measurements and

simulations of supermarkets

with CO 2 refrigeration systems

David Freléchox

Approved

Date Examiner

Björn Palm Supervisor

Samer Sawalha Commissioner

Contact person

Master Student: David Freléchox David Freléchox

Polhemsgatan 34 La Pran 9

11230 Stockholm CH - 2824 Vicques

Registration Number: 821109-A692 (KTH Stockholm)

Registration Number: 650540 (HfT Stuttgart)

Departement: Energy Technology (KTH Stockholm)

Degree Programme: SENCE (HfT Stuttgart)

Examiner at EGI: Prof. Dr. Björn Palm

Supervisor at EGI: Dr. Samer Sawalha

Examiner at HfT: Prof. Dr. Ursula Eicker

KTH Stockholm, Sweden Department of Energy Technology

David Freléchox

I

Abstract

This Master Thesis is a part of a project initiated by Sveriges Energi- & Kylcentrum in

Katrineholm and co-financed by the Swedish energy agency. The project aims to evaluate the

potential of refrigeration systems using carbon dioxide in supermarket refrigeration.

This thesis includes the analysis of three supermarkets using different cooling systems such as

CO2 transcritical chiller unit, CO2 transcritical freezer unit, CO2 transcritical booster unit and

R404A/CO2 cascade unit. The supermarkets have complete instrumentations necessary to

measure temperatures and pressures and measure or evaluate the energy consumptions. The

collected data cover a period of 7 to 18 months. The COP of each system has been evaluated.

Opportunities for improving the regulation of the condensation valve have also been proposed.

CO2 fluid offer possibilities to used floating condensation until a low outdoor temperature and

due to the big pressure difference across the expansion valve, it should still offer improvements

possibilities.

In parallel simulation models of each supermarket were developed using the softwares EES.

The good quality of the collected data allowed validating these models which were then used for

the comparison of the different systems. The COP of each cooling system was being evaluated

in a dynamic manner depending on the condensation temperature, outside temperature and

evaporation temperature. The potential of subcooling and free desuperheating was also

demonstrated.

Several simulations have shown less annual energy consumption for CO2 transcritical plant in

cold and mild climates, as of Stockholm and Frankfurt respectively. On the other hand the

advantage of a cascade system using R404A and CO2 under the hot climate of Phoenix is

evident.

The experimental and theoretical studies reported in this thesis prove that CO2 based system

solutions investigated can be efficient solutions for supermarket refrigeration; however,

comparison with traditional systems is needed and will be presented in following publications in

the ongoing project.


David Freléchox

II

Acknowledgements

The ‘CO2 in Supermarket Refrigeration’ project was initiated as an agreement between

Sveriges Energi- & Kylcentrum (SEK) and KTH/Applied Thermodynamics and Refrigeration

Division. The project is managed by SEK and supported by the companies AGA, Ahlsell,

Cupori, Green and Cool, Huurre, ICA, Oppund Svets, Cupori and WICA. This project is also

financed by Energimyndigheten (STEM) and the partner companies. This thesis work is

involved in this project.

First of all, I would like to express my sincere gratitude to my supervisor Samer Sawalha since

this thesis was only possible due to your support and advice. Thanks for your receptiveness and

your unpretentiousness. It was really pleasant working with you.

Secondly, I would like to extend my gratitude to Jörgen Rogstam who is the project manager in

SEK. Thanks for your support, comments and your interest in my work. The project meetings

were always a great source of inspiration for new ways around CO2 refrigeration systems.

Furthermore, I would like to thank my professor from the Master Program “Sustainable Energy

Competence” in the University of Applied Sciences in Stuttgart (HfT), Professor Ursula Eicker,

and also Professor Björn Palm from Applied Thermodynamics and Refrigeration Division of

Energy Technology Department in Royal Institute of Technology (KTH), for giving me the

opportunity to achieve my Master Degree with this great experience in the North.

And I also would like to thank Jigme Nidup for being my support, company and friend during the

months we spent in the laboratory together. That is also for you, Claire and Jean-Remi, for all

the chats we had in the department and for the marvellous moments we shared during this

period here in Sweden.

This thesis is especially dedicated to my parents, Maggy and Mario who supported me all the

time and in all the decision I made during my studies. Thanks for the trust you have in me. It

helps me to enjoy life and grow in this world.

Finally I want to thank you, you who supported my choices, who followed me around Europe,

who helped me during the hard time and who makes life so nice. You are a beautiful person in

your head and in your heart. With all my love, thanks Nancy !


David Freléchox

III

Table of contents

ABSTRACT .............................................................................................................................................................................................................I ACKNOWLEDGEMENTS................................................................................................................................................................................... II TABLE OF CONTENTS .....................................................................................................................................................................................III LIST OF FIGURES............................................................................................................................................................................................... V LIST OF TABLES..............................................................................................................................................................................................VII NOMENCLATURE...........................................................................................................................................................................................VIII DEFINITIONS......................................................................................................................................................................................................XI 1 INTRODUCTION......................................................................................................................................................................................... 1

1.1 BACKGROUND................................................................................................................................................ 1 1.2 ENERGY USAGE IN SWEDISH SUPERMARKETS................................................................................................ 2 1.3 REFRIGERANT EMISSIONS............................................................................................................................... 3

2 OBJECTIVES............................................................................................................................................................................................... 4

2.1 BACKGROUND................................................................................................................................................ 4 2.2 PROJECT.........................................................................................................................................................5 2.3 SUMMARY ...................................................................................................................................................... 5 2.4 PROJECT PARTNER......................................................................................................................................... 6

3 CO2 TECHNOLOGY................................................................................................................................................................................... 7

3.1 BACKGROUND................................................................................................................................................ 7 3.2 CO2 AS A REFRIGERANT................................................................................................................................ 8

3.2.1 Properties.............................................................................................................................................. 8 3.2.2 Heat exchange characteristics and high pressure compression.......................................................... 12 3.2.3 Efficiency of CO2 versus synthetic refrigerants .................................................................................. 14

3.3 CO2 SOLUTIONS IN SUPERMARKET REFRIGERATION.................................................................................... 17 3.3.1 Indirect systems................................................................................................................................... 17 3.3.2 Cascade DX systems............................................................................................................................ 18 3.3.3 Transcritical DX systems .................................................................................................................... 19

3.4 SAFETY ISSUES............................................................................................................................................. 20 3.4.1 Concentration levels and safety limits................................................................................................. 20 3.4.2 Case study ........................................................................................................................................... 21

4 MEASUREMENTS AND EVALUATION METHODS ................ .......................................................................................................... 22

4.1 PRESSURE AND TEMPERATURE MEASUREMENTS.......................................................................................... 22 4.2 ELECTRICAL POWER CONSUMPTION; MEASUREMENT OR CALCULATION...................................................... 23 4.3 MASS FLOW EVALUATION............................................................................................................................ 26 4.4 COP CALCULATION ..................................................................................................................................... 31

5 FIELD INSTALLATIONS......................................................................................................................................................................... 33

5.1 SUPERMARKET WITH TRANSCRITICAL SYSTEM TR1..................................................................................... 33 5.2 SUPERMARKET WITH TRANSCRITICAL SYSTEM TR2..................................................................................... 36 5.3 SUPERMARKET WITH CASCADE SYSTEM CC1............................................................................................... 39

6 GENERAL SYSTEM ANALYSIS............................................................................................................................................................. 42

6.1 SUPERMARKET TR1..................................................................................................................................... 42 6.2 SUPERMARKET TR2..................................................................................................................................... 45 6.3 SUPERMARKET CC1..................................................................................................................................... 48 6.4 COMPARISON OF THE THREE SYSTEMS......................................................................................................... 50 6.5 COMPARISON OF THE THREE SYSTEMS WITH A LOAD RATIO OF 3 ................................................................. 53


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7 SPECIFIC SYSTEM ANALYSIS.............................................................................................................................................................. 57

7.1 EFFECTS OF THE INSTALLATION OF A FREQUENCY CONVERTER ON THE COMPRESSOR................................. 57 7.2 DISCHARGE PRESSURE VALVE REGULATION................................................................................................ 62 7.3 INFLUENCE OF THE INTERNAL AND EXTERNAL SUPERHEAT ON THE DX CO2 REFRIGERATION SYSTEMS..... 67 7.4 SUBCOOLING WITH GROUND HEAT SINK....................................................................................................... 68 7.5 ANALYSE OF TR2 SYSTEM UNDER TRANSCRITICAL REGIME......................................................................... 71

8 SIMULATION MODEL ............................................................................................................................................................................ 73

8.1 DATA INPUT AND ASSOMPTIONS.................................................................................................................. 73 8.2 FUNCTION TO SIMULATE THE DEPENDENCE OF THE COOLING CAPACITY TO THE OUTDOOR TEMPERATURE.. 76 8.3 FUNCTION TO SIMULATE THE FLUID COMPRESSION...................................................................................... 77 8.4 LIMIT OF THE CONDENSING TEMPERATURE.................................................................................................. 78 8.5 DAY AND NIGHT INFLUENCE ON THE COOLING CAPACITY............................................................................ 79 8.6 VALIDATION OF THE MODELS....................................................................................................................... 81

9 SYSTEMS SIMULATION......................................................................................................................................................................... 83

9.1 VARIATION OF THE CONDENSING TEMPERATURE......................................................................................... 83 9.2 SIMULATION USING IMPROVEMENT POSSIBILITIES FOR CO2 SYSTEMS......................................................... 85 9.3 ANNUAL SIMULATION – COMPARISON OF THE THREE SYSTEMS IN DIFFERENT CLIMATES IN SWEDEN .......... 86 9.4 ANNUAL SIMULATION – COMPARISON OF THE THREE SYSTEMS IN DIFFERENT CLIMATES IN THE WORLD..... 89

10 SPECIFIC SYSTEM SIMULATION................................................................................................................................................... 93

10.1 IMPACT OF THE EVAPORATION TEMPERATURE ON COP ............................................................................... 93 10.2 OPTIMAL CONDENSATION TEMPERATURE FOR THE SUBCRITICAL / TRANSCRITICAL OPERATION TRANSITION

95 10.3 POTENTIAL OF DESUPERHEATING FOR LOW STAGE....................................................................................... 97 10.4 POTENTIAL OF SUBCOOLING WITH GROUND HEAT SINK IN TR2.................................................................... 99

11 DISCUSSION ....................................................................................................................................................................................... 100 12 CONCLUSIONS AND SUGGESTIONS FOR FUTURE WORKS ................................................................................................. 103 13 REFERENCES..................................................................................................................................................................................... 104 APPENDIX 1: FORMULA COPTOT WITH LOAD RATIO CORRECTI ON ............................................................................................ 106


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V

List of Figures

Figure 1.1: A breakdown of energy usage in a supermarket in Sweden. (Arias, 2005)........................................................................2 Figure 3.1: h-logP diagram for the CO2 (left) and the R134a (right)....................................................................................................8 Figure 3.2: Saturation pressure versus temperature for selected refrigerants (Sawalha, 2008)...........................................................9 Figure 3.3: Latent heat of vaporization / condensation (left), saturated vapour density (right) for selected refrigerants (Sawalha,

2008).........................................................................................................................................................................................9 Figure 3.4: Volumetric refrigeration capacity for selected refrigerants (Sawalha, 2008) ....................................................................10 Figure 3.5: Isobaric specific heat of CO2 (left), Isobaric Prandtl number of CO2 (right) (Kim et al., 2003).........................................12 Figure 3.6: Liquid to vapour density (left), surface tension versus saturated temperature for selected refrigerants (Sawalha, 2008).13 Figure 3.7: Compressor pressure diagrams for R134a and CO2 assuming equal cooling capacity (π: pressure ratio, pm: mean

effective pressure) (Kim et al., 2003) .......................................................................................................................................14 Figure 3.8: Comparison of thermodynamic cycles for R134a and CO2 in temperature-entropy diagrams, showing additional

thermodynamic losses for the CO2 cycle when assuming equal evaporating temperature and equal minimum heat rejection temperature(left) (Kim et al., 2003); Comparison of thermodynamic cycles for R134a and CO2 in temperature-entropy diagrams, when assuming equal evaporating temperature and equal logarithmic mean temperature difference(right) (Sawalha, 2009).......................................................................................................................................................................................15

Figure 3.9: Average monthly COP of CO2 and R404A medium (left) and low (right) temperature unit in the climate of Treviso (Italy). (Girotto et al., 2004) ................................................................................................................................................................15

Figure 3.10: Comparison of R404A and CO2 for energy efficiency, medium temperature refrigeration, single stage compressor, direct expansion, no heat recovery. (Haaf, 2005).....................................................................................................................16

Figure 3.11: Secondary fluid systems with phase change. (Girotto, 2005)........................................................................................17 Figure 3.12: Direct expansion system in cascade. (Girotto, 2005) ....................................................................................................18 Figure 3.13: Direct expansion system and transfer of heat directly into the environment. (Girotto, 2005)..........................................19 Figure 3.14: CO2 concentration limit for many safety levels. (Sawalha, 2009)..................................................................................20 Figure 3.15: CO2 concentration against time in a shopping area for leakage durations of 15 minutes, 30 minutes, 1 hour and 2

hours. (Sawalha ,2008) ...........................................................................................................................................................21 Figure 4.1: Schematic of a CO2 Transcritical Supermarket with the pressure and temperature measurement points. ......................23 Figure 4.2: Compressor electrical power measured for one day in July 2008 (01.07.08) in TR1 Supermarket. .................................24 Figure 4.3: Compressor’s electrical power consumption as a function of the pressure ratio for Bitzer compressors in CC1

supermarket. ...........................................................................................................................................................................24 Figure 4.4: Electrical power consumption, comparison with the two methods for a single stage CO2 system during the whole year

2008, KA1 unit in the TR1 Supermarket. .................................................................................................................................25 Figure 4.5: Volumetric efficiency based on compressor data for three CO2 compressors.................................................................26 Figure 4.6: Mass flow of CO2 in the freezer system FA1 during one day of July 2008 in the TR1 supermarket ................................27 Figure 4.7: Mass flow of CO2 in a transcritical system for different mass flow measurement method ...............................................28 Figure 4.8:COP of a CO2 transcritical system for different mass flow measurement method............................................................30 Figure 5.1: Freezer unit in TR1 Supermarket....................................................................................................................................33 Figure 5.2: Schematic diagram of the TR1 system ...........................................................................................................................34 Figure 5.3: Booster unit in TR2 Supermarket....................................................................................................................................36 Figure 5.4: Schematic diagram of the TR2 system ...........................................................................................................................37 Figure 5.5: Two CO2 low temperature units in the CC1 supermarket ...............................................................................................39 Figure 5.6: Schematic diagram of the cooling system in the supermarket CC1.................................................................................40 Figure 6.1: Cooling capacity of one medium temperature unit (KA1) and one low temperature unit (FA1) during the years 2008 -

2009........................................................................................................................................................................................42 Figure 6.2: Compressors electrical power consumption for one medium temperature unit (KA1) and one low temperature unit (FA1)

during the years 2008 - 2009...................................................................................................................................................43 Figure 6.3: COP function of coolant temperature for medium temperature units and low temperature units, measures for TR1

supermarket during 2008.........................................................................................................................................................44 Figure 6.4: COP for each units during the whole testing period for the TR1 supermarket. ................................................................44 Figure 6.5: Different parameters plots for the KAFA1 unit during the whole period of study in the TR2 supermarket ........................45 Figure 6.6: Different parameters plots for the KAFA2 unit during the whole period of study in the TR2 supermarket ........................46 Figure 6.7: Different parameters plots for the KA3 unit during the whole period of study in the TR2 supermarket.............................46 Figure 6.8: COP for each units during the whole testing period for the TR2 supermarket. ................................................................47 Figure 6.9: Cooling capacity, compressor electrical power consumption, condensation and outside temperatures for medium

temperature units VKA1 and VKA2 during the whole testing period for the CC1 supermarket .................................................48 Figure 6.10: Cooling capacity, compressor electrical power consumption, condensation and outside temperatures for low

temperature units KS5 and KS6 during the whole testing period for the CC1 supermarket ......................................................49 Figure 6.11: COP for each units during the whole testing period for the CC1 supermarket. ..............................................................49 Figure 6.12: Load ratio for the three systems analysed during each period of analysis.....................................................................50 Figure 6.13: Condensation temperature for the three systems analysed during each period of analysis...........................................51 Figure 6.14: Condensation temperature for the three systems analysed during each period of analysis...........................................52 Figure 6.15: Load ratio correction for the three systems during the whole period of analysis ............................................................53 Figure 6.16: Total COP with a load ratio of 3 in function of the condensing temperature for the three systems analysed..................54 Figure 6.17 Total COP with a load ratio of 3 in function of the condensing temperature for the three systems analysed, TR1 system

with elimination of the borehole subcooling effect. ...................................................................................................................54 Figure 6.18: Condensation temperature versus outside temperature fort he three systems..............................................................55 Figure 6.19: Total COP with a load ratio of 3 in function of the outside temperature for the three systems analysed ........................56


David Freléchox

VI

Figure 7.1: Electrical power consumption of the compressors in KA2 during March 2009.................................................................57 Figure 7.2: Suction pressure and evaporation temperature of KA2 during March 2009.....................................................................58 Figure 7.3: Internal and external Superheat of KA2 during March 2009............................................................................................59 Figure 7.4: Suction pressure and evaporation temperature of FA2 during March 2009.....................................................................60 Figure 7.5: Daily average of the compressor electrical power and coolant temperature for KA2 during March 2009.........................61 Figure 7.6: Comparison the two KA units of TR1 after the installation of a frequency converter on KA2. ..........................................62 Figure 7.7: CO2 transcritical cycle with gas cooler exit temperature of 40°C at different discharge pr essure (left), (Sawalha, ( 2008) –

Danfoss ICM/ICAD Valve for condensation regulation (right) (Danfoss Refrigeration, 2006)....................................................63 Figure 7.8: IWMAC interface for compressor and gas cooler for KA2, 27 April 2009, 14h10.............................................................63 Figure 7.9: Opening of the discharge pressure regulation valve for KA1 during March 2009 ............................................................64 Figure 7.10: Subcooling of KA1 during March 2009..........................................................................................................................65 Figure 7.11: High pressure for KA1 before the ICAD valve (HP) and after the ICAD valve (HP before exp. Valve) during March 2009.

................................................................................................................................................................................................65 Figure 7.12: EES visualisation on the h-logP diagram for the two condensation regulation functions, KA1 during March 2009.........66 Figure 7.13: Effect of the internal and external superheat on the COP by using CO2 in standard refrigeration system.....................67 Figure 7.14: COP improvement due to the subcooling with the heat sink for KA3 medium temperature unit in TR2 supermarket . ...69 Figure 7.15: Isotherme shape in h-logP diagramm for CO2 near critical point ..................................................................................70 Figure 7.16: Effect on the COP of the subcooling at different condensation temperature for CO2 and R404A with an evaporation

temperature at -10°C and an internal superheat of 1 0 K. .........................................................................................................70 Figure 7.17: KA3 unit in TR2 system during one week at the end of June 2009................................................................................71 Figure 7.18: KA3 unit from TR2 system during two days at the end of June 2009 ............................................................................72 Figure 7.19: KAFA1 unit in TR2 system during four days at the end of June 2009............................................................................72 Figure 8.1: Function binding the percentage used of the maximal cooling capacity to the outdoor temperature................................76 Figure 8.2: Total efficiency of 3 CO2 compressors types in function of the pressure ratio.................................................................77 Figure 8.3: Day and night trend of the cooling capacity of the KA3 unit in the supermarket TR2 during February 2009. ...................80 Figure 8.4: Function binding the percentage used of the maximal cooling capacity to the outdoor temperature with variation range + /

- 25% for a typical medium temperature cabinet. .....................................................................................................................81 Figure 8.5: Comparison of the COP between the template calculation and the EES simulation........................................................82 Figure 9.1: Total COP for different condensation temperature..........................................................................................................83 Figure 9.2: Total COP for different outside temperatures..................................................................................................................84 Figure 9.3: Total COP for different outside temperatures using improvements possibilities for CO2 systems ...................................85 Figure 9.4: Number of hours per year for different outside temperature levels in Storön, Göteborg and Floda – the three locations

are in Sweden. ........................................................................................................................................................................87 Figure 9.5: Annual energy consumption for different supermarket systems in Storön, Göteborg and Floda – all three locations are in

Sweden. ..................................................................................................................................................................................88 Figure 9.6: Number of hours per year for different outside temperature levels in Stockholm / Sweden, Frankfurt / Germany and

Phoenix – Arizona / USA. ........................................................................................................................................................89 Figure 9.7: Annual energy consumption for different supermarket systems in Stockholm / Sweden, Frankfurt / Germany and Phoenix

– Arizona / USA.......................................................................................................................................................................90 Figure 9.8: Annual energy consumption with or without coolant loop for different supermarket systems in Stockholm / Sweden,

Frankfurt / Germany and Phoenix – Arizona / USA..................................................................................................................91 Figure 10.1: COP for medium and low temperature system at different evaporation temperatures on the TR1 system.....................93 Figure 10.2: Relative impact of the evaporation temperature on low and medium temperature systems with reference evaporation

temperatures at -10°C and -35°C. ................... ........................................................................................................................94 Figure 10.3: Enthalpy condenser – gas cooler outlet at different condensation temperature and for two transition temperatures 28

and 30°C. .......................................... ......................................................................................................................................95 Figure 10.4: Total COP and discharge pressure versus the condensing temperature for different transition temperatures with the

supermarket TR2 system (2Transcritical booster units and one chiller unit).............................................................................96 Figure 10.5: Total COP improvement in percentage with a transition temperature at 28°C instead of 31°C related to the Figure 9.2

as usual for a load ratio of 3. ...................................................................................................................................................97 Figure 10.6: Total COP for the supermarket CC1 with and without free desuperheat on the low temperature unit............................98 Figure 10.7: Effect of the ground heat sink when it is using to subcool the liquid in TR2 supermarket ..............................................99


David Freléchox

VII

List of tables

Table 7-1: Monthly average of the electrical power consumption for KA2 and FA2...........................................................................61 Table 7-2: List of performance data when one compressor is running for KA1 and FA1 for March 2009 before and after changing the

function of the regulation valve. ...............................................................................................................................................66 Table 8-1: Data input and assumptions for the simulations...............................................................................................................74 Table 9-1: Annual energy consumption excess in percentage in comparison with the better system for each supermarket systems in

Storön, Göteborg and Floda. ...................................................................................................................................................88 Table 9-2: Annual energy consumption excess in percentage in comparison with the better system for each supermarket systems in

Stockholm / Sweden, Frankfurt / Germany and Phoenix – Arizona / USA. ...............................................................................92


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VIII

Nomenclature

Roman

A Area [m2]

C Concentration [PPM]

CC Cascade refrigeration system

22 COorCO Carbone dioxide

COP Coefficient of performance [-]

pc Specific heat [kJ/kg*K]]

CR Circulation ratio [-]

d Pipe diameter [m]

dP Pressure drop [kPa]

dT Temperature drop [K]

DX Direct expansion

E& Electrical power [kW]

EES Engineering Equation Solver

f Friction factor [-]

FA Low temperature unit or cabinet

h Enthalpy [kJ/kg]

fgh Latent heat of vaporization [kJ/kg]

IDLH Immediately Dangerous to Life or Heath

IHE Internal heat exchanger

KA Medium temperature unit or cabinet

KAFA Booster system with low and medium temperature

L Pipe length [m]

LTMD Logarithmic Mean Temperature Difference [K]

LR Load ratio

corrLR Load ratio correction, fixed value

m& Mass flow [kg/s]

n Rotational speed [rpm]

33 NHorNH Ammonia


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Nu Nusselt Number [-]

P Pressure [bar absolute]

PPM Parts per Million

Pr Prandtl Number [-]

PR Pressure ratio [-]

cQ& Condensation capacity [kW]

oQ& Cooling capacity [kW]

vq Volumetric refrigeration effect [kJ/m3]

Re Reynolds Number [-]

SC Subcritical refrigeration system

SH Superheat [K]

T Temperature [°C]

TR Transcritical refrigeration system

V& Volume flow [m3/s]

w Velocity [m/s]

x Vapour quality [-]

*x Position along the heat exchanger [m]

Y Constant [kW2/m4]

Greek

α Heat transfer coefficient [W/m2*K]

∆ Difference [-]

ε Heat exchanger effectiveness [-]

λ Thermal conductivity [W/m*K]

ρ Density [kg/m3]

isη Isentropic efficiency [-]

vη Volumetric efficiency [-]

totη Total efficiency [-]

µ Dynamic viscosity [kg/s*m]

ν Specific volume [m3/kg]


David Freléchox

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Subscript

abs Absolute

air For air

amb Ambient

app Approach temperature difference

booster Booster system

cab Cabinet medium temperature

chiller Chiller

comp Compressor

cond Condenser

el Electric

evap Evaporation

in Inlet

is Isentropique

f Fluid

freezer Freezer

gc Gas cooler

losses Heat losses

map Map or design conditions

new New or running conditions

out Outlet

cooleroil Oil cooler losses

s Surface

sat Saturation


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Definitions

CFC: Chlorofluorocarbon is any of various halocarbon compounds consisting of

carbon, hydrogen, chlorine, and fluorine.

GWP: Global Warming Potential. The GWP is a measure of how much a given mass of

greenhouse gas is estimated to contribute to global warming. It is a relative scale

which compares the gas in question to that of the same mass of carbon dioxide

(whose GWP is by definition 1). A GWP is calculated over a specific time interval

and the value of this must be stated whenever a GWP is quoted or else the value

is meaningless.

HCFC: Hydro chlorofluorocarbons are halogenated compounds containing carbon,

hydrogen, chlorine and fluorine. They have shorter atmospheric lifetimes than

CFCs and deliver less reactive chlorine to the stratosphere where the “ozen

layer” is found.

HFC: Hydro fluorocarbons contain no chlorine. They are composed entirely of carbon,

hydrogen, and fluorine. They have no known effects at all on the ozone layer.

Only compounds containing chlorine and bromine are thought to harm the ozone

layer. Fluorine itself is not ozone-toxic. However, HFCs and perfluorocarbons do

have activity in the entirely different realm of greenhouse gases, which do not

destroy ozone, but do cause global warming. Two groups of haloalkanes, hydro

fluorocarbons (HFCs) and perfluorocarbons (PFCs), are targets of the Kyoto

Protocol.

ODP: Ozone Depletion Potential. The ODP of a chemical compound is the relative

amount of degradation to the ozone layer it can cause, with

trichlorofluoromethane (R-11) being fixed at an ODP of 1.

Source: Wikipedia.org and Wikipedia.fr, 10th march 2009

KTH Stockholm, Sweden Departement of Energy Technology

David Freléchox

1

1 Introduction

1.1 Background

Following the depletion problems of the ozone layer through the release of CFC refrigerant or

HCFCs in the atmosphere, the politicians decided through the Montreal Protocol in 1987 to

prohibit the use of these substances. CFCs are forbidden since 1996 and HCFCs about in

2010. The chemistry has proposed a new type of synthetic fluid called HFCs. These Hydro-

Fluoro-Carbons did not contain the chlorine molecule responsible for the depletion of the ozone

layer any more and showed promise. The ozone layer damaging problem is gently solving but

the problem related to the refrigerant has changed to the increase of the greenhouse effect on

the earth. HFCs have an ODP-Value (Ozone Depletion Potential) of zero given the absence of

chlorine in their structure, but they still have a high GWP value (Global Warming Potential). For

example R404A has a GWP value of 3800 that is much higher than the reference fluid, CO2,

which has a GWP value of 1. This has led to their integration into the Kyoto Protocol, which is in

application since 2005, and implicates various restrictive measures. HFCs emissions control is

imposed and reduction of global emissions are established. This is translated into reality by

limiting the amount of HFC fluids in the refrigeration systems and other periodic leak mandatory.

The ultimate objective is to prohibit the use of these substances.

The search for an adequate and satisfactory solution has resulted in the proposal of a returning

to natural refrigerants. NH3, most commonly ammonia and CO2 or carbon dioxide came back to

use after a widespread application in the early of the 20th century. In commercial refrigeration,

CO2 is favoured despite thermodynamic properties being not quite as good as NH3. It has a

main safety advantage over NH3 which can be dangerous already at relatively low

concentrations.

However, the use of a natural refrigerant will not solve the problems attached to the

environment if it consumes more energy. The solution has to take into account the direct and

indirect impact of the cooling solution on the environment. The new designed system using

natural refrigerant must match or surpass the energy efficiency of old solutions.


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2

1.2 Energy usage in Swedish supermarkets

Traditionally, supermarkets have always been major consumers of energy, particularly of

electrical energy for lighting and for the refrigeration systems. According to Orphans (1997), the

share of energy used from the U.S. and French supermarkets reach 4% of the national energy

consumption. Regulations and needs for the conservation of fresh and frozen products do not

offer many saving possibilities to decrease the cooling capacity and energy consumption of the

refrigeration system.

To reduce electrical consumption in supermarkets, research in more efficient systems is

needed. Our study focuses on the refrigeration systems which, as show on the Figure 1.1

covers 35 to 50% of the energy consumption in supermarkets. In any case, it is a very important

part of the energy and maintenance costs. The technology has future requirements in better

efficiency coupled with proven reliability.

Figure 1.1: A breakdown of energy usage in a supermarket in Sweden. (Arias, 2005)


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1.3 Refrigerant emissions

The impact of refrigerant leakage on the environment has affected the design of refrigeration

systems in supermarkets. Commercial refrigeration is the sector with the largest refrigerant

emissions totalling about 185,000 metric tonnes in 2002, which is equivalent to 37% of the

worldwide refrigerant emissions [PAL04].

Commercial refrigeration systems refrigerant emissions in the 1980’s were reported to be in the

range of 20-35% of refrigerant charge on an annual basis for developed countries. The high

emission rates were due to design, construction, installation, and service procedures followed

without awareness of potential environmental impact. Emissions have been decreasing due to

industry actions and governmental regulations for refrigerant containment, recovery, and usage

record keeping, increased personnel training, improved service procedures, and attention to

many details in system design [BIV04].

These enormous leakage rates have gradually decreased thanks to the use of more reliable

montage technique, reducing fluid amount through the use of indirect systems and the

introduction of a variety of regulations resulting in higher prices of refrigerants. The annual

leakage rate of this century is generally around 10%.

Bivens and Gage in their study from 2004 compare several European countries. Briefly and

generally, Germany announced an annual rate of leakage of 10% for a series of supermarket

with R404A. Denmark is also in this range but including all types of refrigerants. Norway

announced a rate of 14% following a study between 2002 and 2003. Sweden has improved its

annual rate from 14.0% in 1993 to a rate of 10.4% in 2001. In this case, there is a large disparity

between the fluids. Finally, the Netherlands is the best cited example. Since 1992, the country

has implemented strict regulation for the use of refrigerants with the creation of a special

structure (STEK) to reduce emissions of refrigerant. A Dutch report from 1999 is cited which

demonstrated the effectiveness of this process with an annual leakage rate of 3.2%.


David Freléchox

4

2 Objectives

The main objective of this project is to analyze and then compare the energy performance of

several supermarkets using CO2 as refrigerant. The supermarkets use different cooling system

configurations, and a simulation tool is essential to achieve comparisons. The in-situ

measurements allow validating these simulations which are then used to analyze the

performance of each system according to different variables or under different climates.

2.1 Background

This project is run by Sveriges Energi & Kylcentrum (SEK) which is a subsidiary company of

Installatörernas Utbildingscentrum in Katrineholm, in cooperation with KTH. Using CO2 as a

refrigerant in supermarkets is becoming increasingly popular in Sweden. It is a promising

technology which offers opportunities in energy savings and protecting the environment. CO2

has been used as refrigerant in different system solutions: transcritical, cascade and indirect.

The efficiency of each system solution depends on several parameters such as the system

capacity, heating requirements, climate conditions, etc. Proper evaluation of the currently

installed CO2 system solutions is needed to facilitate the application of the new technology.

This project is a continuation of the work that has been conducted by KTH in cooperation with

Sveriges Energi & Kylcentrum (SEK). The application of CO2 in supermarket refrigeration has

been theoretically and experimentally investigated. Computer simulation models have been built

for the theoretical analysis of the different CO2 system solutions.


David Freléchox

5

2.2 Project

In this project several supermarket installations with different CO2 systems and conventional

solutions will be evaluated. The main tasks in the project are to build the measuring equipment,

install it in the supermarkets and to create the templates for collecting the data and to run the

calculations. The task in the MSc thesis work includes analysis of data from the field

measurements, focus will be on system’s cooling performance and efficiency. Computer models

will be used to simulate the performance of the different system solutions. Compare the field

measurements with the results from the computer simulation models. Performance of the

different system solutions will be compared and suggestions for modifications will be made.

In summary, the following schedule has been followed:

� Documentation on CO2 refrigerating technology

� Collecting data on refrigeration systems

� Creating a database for each supermarket

� Data processing, calculation of needed thermodynamic parameters

� Modifications and adaptations of existing computer simulations for each system

� Calculation of the COP using the measurements and the computer models

� Model validations

� Proposition for improvements to optimize energy efficiency

2.3 Summary

The work in this thesis started by surveying three existing CO2 supermarket installations in

Sweden. Pressures, temperature and energy consumption were collected for different periods in

each supermarket. A template in order to calculate all the thermodynamic states of the systems

was created, cooling capacities and COP of each system were calculated.

The approach of this experimental project allows evaluating the performance of each

refrigeration system and compare them. Through the online data collection for three running

supermarkets, it is possible to assess the influence of external parameters, such as outside

temperature and/or evaporation temperature, and demonstrate their influence on the energy

consumption of various CO2 refrigeration systems.


David Freléchox

6

In order to perform theoretical evaluations of the performance of different CO2 system solutions

computer simulation models have been used to simulate CO2 transcritical parallel system, CO2

transcritical booster system and R404A/CO2 cascade system. Based on the measurement a

validation of the model has been achieved.

After studying the modifications made by the installer on different supermarkets and in

combination with the simulations based on the computer models, performance evaluation and

analysis of systems’ optimizations have been done. Suggestions for improvements and

recommendations for future research topics have subsequently been drafted.

This project will facilitate further development of refrigerating system using CO2 as refrigerant; it

will also provide answers on the efficiency of these systems and the possibility of using CO2 as

a long term solution in supermarket refrigeration.

2.4 Project partner

Organization: Participant/s:

Sveriges Energi- & Kylcentrum Institution Jörgen Rogstam

KTH – Energiteknik University Björn Palm / Samer Sawalha

ICA Supermarket Per-Erik Jansson

Green and Cool Supplier Micael Antonsson

WICA Supplier Peter Rylander

Ahlsell Installer Torbjörn Larsson

Huurre Installer Göran Sundin

AGA Christer Hens

Tranter Ulf Vestergren

Cupori David Sharp

Oppunda Svets Ken Johansson

Energimyndigheten Conny Ryytty


David Freléchox

7

3 CO2 Technology

3.1 Background

The use of CO2 as refrigerant was widespread at the end of the 19th and the beginning of the

20th Century, particularly in the applications of marine refrigeration. Thereafter, with the

appearance of new synthetic refrigerants like CFC, its use has gradually reduced until its

complete abandonment. [KIM03]

The reasons are relatively easily identifiable; the synthetic refrigerants generally work at a lower

pressure, which make the implementation easier during the assembly facilities. Component

manufacturers have also been able to maintain a competitive price for their components for this

reduced pressure. The relatively low temperature at the critical point for CO2 also caused

difficulties in providing the required capacity. The systems used did not work at optimal

transcritical operation which caused technical difficulties and efficiency losses.

Currently, we are witnessing a renaissance of this fluid for refrigeration applications. Its absence

of effect on the ozone layer and its minimal impact on the greenhouse effect compared to

synthetic refrigerants made it as a favourite alternative from environmental point of veiw. This

renaissance is largely attributed to the work of Professor Gustav Lorentzen who suggested the

use of CO2 in a transcritical cycle during the end of the 1980s [SAW08]. His solution for the

automotive air-conditioning has subsequently led to work on other applications, such as

commercial refrigeration and led about 15 years later on the construction of the first

supermarkets with CO2 transcritical application.

The CO2 has several interesting characteristics; it is non-flammable, non-explosive and

relatively non-toxic. It is present in the air at a concentration of 350-400 PPM. As stated its ODP

value is zero and its GWP value is very low, 1.


David Freléchox

8

The Figure 3.1 shows a comparison of the h-logP diagrams for CO2 and R134a using exactly

the same parameters. First difference, CO2 is transformed gradually into a solid form below 5.2

bars. Second major difference, the CO2 critical point is located almost at ambient temperature

31.1°C and a pressure of 73.8 bars, while that of R 134a is much more distant from the ambient

at 101.1°C and 40.1 bars. This implies for a standa rd condensing temperature of 40°C, a CO2

high pressure around 100 bars, and an R134a high pressure 10 times lower.

Figure 3.1: h-logP diagram for the CO2 (left) and the R134a (right)

3.2 CO2 as a refrigerant

3.2.1 Properties

The CO2 working pressures are a big challenge in the implementation of technical system,

particularly on the high pressure side. The pressures in a CO2 refrigeration system are 5 to 10

times higher than with traditional fluids, as can be observed in Figure 3.2. This leads to

development needs for components, particularly compressors. This challenge is taken up by

manufacturers. Installers must also follow a formation to be certified in the field of brazing /

welding and need to have experience with this new fluid.


David Freléchox

9

Figure 3.2: Saturation pressure versus temperature for selected refrigerants (Sawalha, 2008)

However, Figure 3.3 shows that R134a and CO2 have similar latent heat of vaporization at

common evaporation temperature. These two fluids have a latent heat of vaporization in the

range 200 to 400 kJ/kg. NH3 has significantly higher values.

Closely linked to its high saturation pressures, CO2 has a higher saturated vapour density than

other refrigerants, as can be observed in Figure 3.3. This is an advantage to obtain a low

volume flow for a given mass flow, and ensure small component size.

Figure 3.3: Latent heat of vaporization / condensation (left), saturated vapour density (right) for selected refrigerants (Sawalha, 2008)


David Freléchox

10

According Equation 3-1, the two properties of Figure 3.3 multiplied give the volumetric capacity

of refrigeration on Figure 3.4. This is an important parameter since it defines, for a given

capacity, the volume flow of the compressor and its size. CO2 with its high saturated vapour

density and its latent heat of vaporization equivalent to the other fluid has a volumetric capacity

of refrigeration about 5 times larger than other refrigerants. Thanks to this property, systems

running with CO2 have, for a given refrigeration capacity, smaller compressors. The tubing size

is also reduced since the pressure drop is inversely proportional to the volumetric refrigeration

capacity as shown in Equation 3-5.

..vapsatfgv hq ρ⋅= 3-1

Figure 3.4: Volumetric refrigeration capacity for selected refrigerants (Sawalha, 2008)

The CO2 pressure drop can be related to other refrigerants by applying the properties

presented above to the following single phase flow pressure drop equation [SAW08].

d

LwfP ⋅⋅⋅=∆

2

2

ρ 3-2

Assuming that the refrigerant enters the evaporator as saturated liquid and evaporates

completely the velocity can be expressed as

Ah

Q

A

mw

fg ⋅⋅=

⋅=

ρρ&&

3-3


David Freléchox

11

Then

d

L

Ah

QfP

fg

⋅⋅

⋅⋅⋅⋅=∆

21

2

ρρ

&

3-4

For a system with the same capacity, the same tube dimensions and same operating

conditions, the variables are the density ρ and the latent heat of vaporization fgh . And

rearranging the equation with 3-1.

Yhq

Pfgv

⋅⋅

=∆ 1 3-5

Using Equation 3-5 it is easy to prove that the pressure drop is based on the latent heat of

vaporization fgh and the volumetric capacity vq . Y is constant (kW2/m4) which collects the

fixed parameters related to the geometry and the operating conditions.

It is important that the saturation temperature drop that is attached to the pressure drop is low,

so that it will be less detrimental to the coefficient of performance (COP) of the system. The

Clapeyron equation [SAW08] is used to convert the pressure drop to an equivalent change in

saturation temperature

Ph

vvTT

fgabssat ∆⋅−⋅=∆ )( 12

3-6

Tabs is absolute temperature of the fluid (K), v2 is the specific volume for the saturated vapour

(m3/kg) which is much larger than the specific volume of the saturated liquid (v1), so v1 can be

ignored in Equation 3-5. Consequently and with Equation 3-1 the above relationship can be

expressed as follows

Pq

TTv

abssat ∆⋅⋅=∆ 1

3-7

For the same operating temperature, Tabs , CO2 will have a lower pressure drop, as discussed

above, and the corresponding temperature drop will also be lower due to the high volumetric

refrigerating effect of CO2.

Due to the high volumetric refrigerating effect, low pressure and low temperature drops it is

therefore possible to design smaller and more compact components with CO2.


David Freléchox

12

3.2.2 Heat exchange characteristics and high pressure com pression

The heat exchange properties of each refrigerant are important to ensure good performances. A

small reminder about the heat transfer allows identifying the key characteristics of CO2.

Equation 3-7 gives the capacity of a heat exchanger which is affected by the coefficient of

convection α defined through Equation 3-8. Nusselt Number from Equation 3-9 und Prandtl

Number from Equation 3-10 take place in the calculation of the coefficient of convection.

sfp TATcmQ ∆⋅⋅=∆⋅⋅= α&& 3-8

where

L

Nu λα ⋅= 3-9

and

Pr),Re,( *xxfNu = 3-10

where

λµ⋅

= pcPr 3-11

The specific heat determines the fluid’s ability to transfer heat at a given rate and a given

temperature differential. The Prandtl number shown in Equation 3-11 is characterized by the

fluid properties and influences the coefficient of convection and thus the heat transfer.

Figure 3.5 shows CO2 properties which change according to temperature during the heat

rejection process and result in considerable variations of pressure drop, temperature drop, and

heat transfer coefficient. It is important to take into account these changes in the calculation and

the dimensioning of heat exchangers.

Figure 3.5: Isobaric specific heat of CO2 (left), Isobaric Prandtl number of CO2 (right) (Kim et al., 2003)


David Freléchox

13

The low surface tension of CO2 and the low density ratio liquid/vapour which means good

distribution in the evaporator, allow CO2 to boil quickly and with low deltaT, which improves its

heat exchange abilities.

Figure 3.6: Liquid to vapour density (left), surface tension versus saturated temperature for selected refrigerants (Sawalha, 2008)

When evaporation occurs at a solid-liquid interface, it is termed boiling. The process occurs

when the temperature of the surface Ts, exceeds the saturation temperature Tsat corresponding

to the liquid pressure. The process is characterized by the formation of vapour bubbles, which

grow and subsequently detach from the surface. In fact through boiling or condensation, large

heat transfer rates may be archived with small temperature difference. In addition to the latent

heat hfg, two other parameters are important in characterizing the process, namely, the surface

tension σ between the liquid and vapour interface and the density between the two phases. This

difference induces a buoyancy force. Because of combined latent heat and buoyancy driven

flow effects, boiling and condensation heat transfer coefficients and rates are generally much

larger than those characteristic of convection heat transfer without phase change. [INC96]


David Freléchox

14

CO2 compressors generally operate at higher pressures and with a larger pressure differential

than traditional refrigerants; but its pressure ratio is lower. As can be seen in the Figure 3.7, the

piston displacement is 6.7 superior with R134a than with CO2 for the same cooling capacity.

Losses from the re-expansion of the fluid after the compression are also lower with CO2

compressors. Despite the high levels of pressure and the shape of its Pressure-Volume (PV)

diagram, the negative effects of the pressure drop through the valves tend to be lower for CO2

compressors and give them a better efficiency.

Figure 3.7: Compressor pressure diagrams for R134a and CO2 assuming equal cooling capacity (π: pressure ratio, pm: mean effective pressure) (Kim et al., 2003)

3.2.3 Efficiency of CO2 versus synthetic refrigerants

Assuming given evaporating temperature and given minimum heat rejection temperature, the

transcritical cycle suffers from larger thermodynamic losses than an ‘ordinary’ cycle with

condensation. Owing to the higher average temperature of heat rejection, and the larger

throttling loss, the theoretical cycle work for CO2 increases compared to a conventional

refrigerant as R-134a as indicated in the Figure 3.8.

Despite this, for a given heat exchanger and a given coolant temperature, the CO2 gas cooler

output temperature could be less than for a standard cycle. This results of the higher logarithmic

mean temperature difference between the refrigerant and the coolant. Moreover, given the

positive properties of CO2 for heat transfer, the evaporation temperature could generally be

higher with CO2. [KIM03]


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15

Figure 3.8: Comparison of thermodynamic cycles for R134a and CO2 in temperature-entropy diagrams, showing additional thermodynamic losses for the CO2 cycle when assuming equal evaporating temperature and equal minimum heat rejection temperature(left) (Kim et al., 2003); Comparison of thermodynamic cycles for R134a and CO2 in temperature-entropy diagrams, when assuming equal evaporating temperature and equal logarithmic mean temperature difference(right) (Sawalha, 2009)

According to an experimental study of Girotto et al. (2004) presented through Figure 3.9, the

performance of refrigeration systems using CO2 as refrigerant can matched or exceed the

performance of traditional systems using synthetic fluids. In order to compare CO2 to synthetic

fluids, we have to consider the favourable properties of CO2 in terms of heat exchange and

pressure drop. It is common to provide a suction temperature 2 K higher with CO2 than with

traditional systems. This obviously creates an improved COP. Note that in this study, the lower

limit for the floating condensation is set at 25°C in case of R404A and 10°C in case of CO2.

Floating condensation promotes CO2 as well. With carbon dioxide, it is possible to reduce the

condensing temperature and gas cooling temperature lower than with HFC. These differences

explain the good performance of CO2 systems during the winter months. Note that both system,

MT and LT, operate subcritical when the outside temperature is below 15 °C. [GIR04]

Figure 3.9: Average monthly COP of CO2 and R404A medium (left) and low (right) temperature unit in the climate of Treviso (Italy). (Girotto et al., 2004)


David Freléchox

16

In this experimental case, the annual energy’s consumption of the CO2 system is about 10%

higher than the R404A system, for the same cooling capacity. This difference is solely due to

the overconsumption of the MT unit. The LT unit energy consumption is equivalent to that with

R404A. Girotto et al. (2004) suggests that such CO2 systems could reach the annual

performance of traditional systems in cold climates, such as cities of central or north Europe like

Brussels or Stockholm.

Linde Kältetechnik GmbH is one of the leaders in the CO2 Technology. The Figure 3.10

illustrates the relationship between 1/COP and outdoor temperature for two different systems

using R404A and CO2. Operation with CO2 seems to possess distinct advantages in respect of

energy efficiency in the lower range of outdoor air temperatures, whereas the energy efficiency

of both refrigerants studied is equal in the mid-range up to an outdoor air temperature of 26 °C.

Above an outdoor air temperature of 28 °C (supercri tical operation), the reverse applies and

refrigeration systems working with R 404A are more efficient. At an outdoor air temperature of

35 °C, the difference in energy requirement is 13 % . In order to overcome this energy-related

drawback at high outdoor air temperatures, a CO2 gas cooler equipped with water spray system

has been installed. The water spray enables the gas cooler outlet temperature of the CO2

process to be lowered well below air temperature, greatly reducing energy consumption in the

summer and preventing an increase in electric power requirement at temperatures over 29 °C.

[HAA05]

Figure 3.10: Comparison of R404A and CO2 for energy efficiency, medium temperature refrigeration, single stage compressor, direct expansion, no heat recovery. (Haaf, 2005)


David Freléchox

17

3.3 CO2 solutions in supermarket refrigeration

In general, two temperature levels are required in supermarkets for chilled and frozen products.

Product temperatures of around +3°C and –18°C are c ommonly maintained. In these

applications, there are mainly three design options: indirect system, cascade DX system or

transcritical DX system. It is also possible to advantage of different system and built mixed

system. The following section describes CO2-based solutions that fulfil the refrigeration

requirements of supermarkets.

3.3.1 Indirect systems

The main refrigeration circuit, of the conventional type with HFC or NH3, conveys heat from the

main evaporator to the secondary fluid, which is pumped, obviously in liquid state, into the

evaporators positioned inside the units to be refrigerated. The secondary fluid CO2 evaporates

and removes heat from the units. The circulation ratio of the CO2 in the secondary circuit is

generally between 1.5 and 3, and its highest operating temperature is at -10°C and lowest at -

40°C. The following Figure 3.11 shows a schematic o f the system and the cycle on the h-logP

diagram.

Figure 3.11: Secondary fluid systems with phase change. (Girotto, 2005)

Compared to traditional indirect systems, with propylene or ethylene glycol, the system in

question requires lower flow rates and consequently smaller pipes and less pumping power and

of course does not feature any change in temperature in the evaporators. The solution with

forced circulation of CO2 offers major advantages in very extensive systems, with hundreds of

metres between the units to be refrigerated and the central refrigeration unit, main advantages

are: [GIR05]


David Freléchox

18

� non-toxic fluid in circulation

� no problem as regards the return of the oil regarding DX systems

� low energy consumption for pumping regarding indirect brine systems

3.3.2 Cascade DX systems

At low temperatures (evaporation at temperature below -30°C) the cascade system is

preferable. As can be seen on the Figure 3.12, two refrigeration units, each optimised for its

own operating range, are thermally linked in series by means of an intermediate exchanger,

which for one of the units represents the evaporator and for the other, the condenser.

For commercial refrigeration, for cost reasons, R404A or R507 are used in the high-temperature

circuit, while NH3 is used for industrial refrigeration, this is especially suitable for this application

because each NH3 and CO2 operates in its optimum temperature range. The risk related to the

use of NH3 in premises where there could be people is thus eliminated, and this represents a

big advantage.

For evaporation temperatures around -30°C, in appli cations where well water can be used as

heat source, instead of using a cascade system with NH3, it could be possible to operate in

single stage with the CO2 system, in the event of the size of the compressors available today

for a supply pressure of up to 70 bar being big enough. In cold climates, air from outside could

even be used for most of the year. [GIR05]

Figure 3.12: Direct expansion system in cascade. (Girotto, 2005)


David Freléchox

19

3.3.3 Transcritical DX systems

In this system the CO2 at high pressure rejects heat directly into the air heat exchanger (a

secondary water circuit can naturally be used, but the cost is higher). As can be seen on the

Figure 3.13, the cycle should be supercritical when the condensing temperature is above the

critical point at 31°C. For lower temperatures, the cycle is subcritical.

The fundamental difference between operations in the two conditions lies in the way the high

pressure is controlled. While in subcritical conditions, the high pressure is indirectly set by the

temperature of the cooling fluid. In the case of operation with cooling above critical pressure, a

special control is required, and this can be done through the valve positioned between the

exchanger and the receiver. The control method used must optimise capacity and efficiency.

The efficiency of the standard cycle in supercritical conditions is much below that of the same

cycle with HFC/ HC or NH3, conditions being the same. In cold climates, the system can

operate in subcritical conditions for most of the time. [GIR05]

Figure 3.13: Direct expansion system and transfer of heat directly into the environment. (Girotto, 2005)


David Freléchox

20

3.4 Safety issues

3.4.1 Concentration levels and safety limits

The high pressure of the CO2 generates technical difficulties but also a safety issue. However,

CO2 is non-flammable and non-explosive so the overpressure risks in CO2 systems are

relatively easily solved by installing discharge valves.

CO2 is odourless and deadly at high concentration; this obviously makes it dangerous for use in

places such as supermarkets. Being naturally present in the air, it is not dangerous in low

concentrations. Figure 3.14 presents the different levels of dangerousness of this fluid. A brief

comparison with synthetic fluids highlights the fact that these substances are dangerous for

humans at high concentrations, as well. Their use in the last century has not caused any

particular problem on this side. However, CO2 is more difficult to identify as it is absolutely

odourless, its density is higher than the air’s so it tends to accumulate towards the ground. This

makes it particularly dangerous for young children. To mitigate this risk, a CO2 refrigeration

system requires CO2 concentration detectors in the lower part of each enclosed spaces,

especially the engine room, insulated rooms, and several places of business premises.

Figure 3.14: CO2 concentration limit for many safety levels. (Sawalha, 2009)


David Freléchox

21

3.4.2 Case study

An experience of Samer Sawalha [SAW08] as part of his doctoral thesis allows establishing the

dangerousness of this fluid. The case study is an average class size supermarket of 40 x 30 x 5

m with 75 kW capacities for medium temperature refrigeration system and 30 kW for the low

temperature. CO2 is used as secondary refrigerant – CO2 pumped circulation – for the low

temperature and the total charge is about 100 kg. It is admitted that this supermarket, for the

shopping area, has 0.5 air changes per hour (ACH).

Figure 3.15 shows the levels of the concentration after the escape of all of the fluid in a given

time. Generally the concentration peak is reached relatively quickly after the leakage. The

ventilation system with its air change rate of 0.5 per hour progressively dilutes the CO2. In this

case, the threshold concentration of 5000 PPM is exceeded in all cases and for about 2 hours.

However, the concentrations peak of 9000 PPM when 100 kg of CO2 escape in 15 minutes is

not a life danger for the customers and employees of the supermarket. But an alarm system is

required to treat the problem quickly.

Figure 3.15: CO2 concentration against time in a shopping area for leakage durations of 15 minutes, 30 minutes, 1 hour and 2 hours. (Sawalha ,2008)

For machine rooms or insulated cells, the situation is different and critical thresholds can be

achieved. A detection system equipped with warning light and sound must be installed to

mitigate those risks. Nevertheless, the probability that all the refrigerant escape in such short

time is very low and rare in practice. Concentrations due to leakage will generally evolve at

lower levels.


David Freléchox

22

4 Measurements and Evaluation Methods

The important parameters for the evaluation of cooling systems are mainly the cooling

capacities and the COPs. For these capacities, the temperatures and pressures are needed to

determine the enthalpies and then the mass flow rate is needed in order to determine the

cooling capacities and different losses. Mass flow rate is not measured directly and it is evaluted

from the compressor side. Then COPs are calculated using the cooling capacity and the

measured or calculated electrical consumption.

4.1 Pressure and temperature measurements

The input data for the calculation of the refrigerant thermodynamics states are the measures of

pressure and temperature. The temperature sensor types are generally PT100 or PT1000 and

widely used in the refrigeration regulation. Pressure sensors give an absolute or relative

pressure depending on their initial settings. The sensors used are generally from the

manufacturer Danfoss and types are AKS or HSK according to their pressure range.

The sensors were not installed especially for our study but are primarily used to operate the

systems and are essential regulation elements. On the Figure 4.1, the measurement points are

shown. These points allow tracing the refrigeration cycle in the h-logP diagram and calculating

the cooling capacity as well as various parameters which could influence this capacity, such as

the internal and external superheat, the subcooling and the pressure ratio.

We used two different systems for the data acquisition:

� -IWMAC with an interval of 5 minutes for TR1 and TR2 supermarket [IWM09]

� -RDM with an interval of 15 minutes for CC1 supermarket [RDM09]


David Freléchox

23

Figure 4.1: Schematic of a CO2 Transcritical Supermarket with the pressure and temperature measurement points.

4.2 Electrical power consumption; measurement or calcul ation

Measuring electrical energy or electrical power consumptions is not necessarily complicated,

but it is usually expensive and is not needed for the regulation. It is merely informative and

important for our project, but not essential for the refrigeration system. Therefore, it was often

difficult to obtain these measurement points. In the case of impossibility to get these measures,

we adopted a method of calculating value based on the pressure ratio. We used the calculation

methods on the CC1 supermarket since we do not have any energy measurement tools

installed on this supermarket.

Finally, we used two different methods to define the electrical power consumption:

� Power consumption measurements for the TR1 and TR2 supermarket

� Power consumption calculations for the CC1 supermarket

The first method is easy. A measuring device collects consumption values with the same

definition that the measures of pressure and temperature, so 5 or 15 minutes. The Figure 4.2

the electrical consumption for one day of July for the freezer (FA) unit and the chiller (KA) unit.

The collecting interval is 5 minutes for this device, thus about 300 measurement points per

parameters each days.


David Freléchox

24

Figure 4.2: Compressor electrical power measured for one day in July 2008 (01.07.08) in TR1 Supermarket.

The second method uses a mathematic formula to calculate the power in function of the

pressure ratio. These two formulas are shown on the Figure 4.3. The determination of this

formula has been done with compressor manufacturer data [BIT09]. Obviously it is different for

each type of compressor. The function is slightly different for each evaporation pressure, but

this one is rather stable on our systems, so we decided to use the function for a given

evaporation pressure. This gave satisfactory results.

y = -0.2807x2 + 4.0975x + 3.3957

y = 0.65x - 0.060.00

2.00

4.00

6.00

8.00

10.00

12.00

14.00

16.00

18.00

20.00

0.0 1.0 2.0 3.0 4.0 5.0 6.0

Pressure ratio [-]

Ele

ctric

al p

ower

con

sum

ptio

n [k

W]

Bitzer 4H-15.2YP_evap = 4 bar

Bitzer 2KC-3.2K-40SP_evap = 10 bar

Figure 4.3: Compressor’s electrical power consumption as a function of the pressure ratio for Bitzer compressors in CC1 supermarket.


David Freléchox

25

To check the accuracy of the method based on a calculation, we ran a comparison with a

refrigeration system for which we had electrical power measures. Figure 4.4 shows conclusive

result. The difference between the measured and the calculated value is at most 5%. The origin

of this divergence may be various, such as uncertainty in the definition of the number of

compressor running, or of course the change in operating conditions of the system because the

calculation method uses manufacturer data to create the function. However, the variations are

very reasonable and the use of this method is therefore a good alternative when we do not have

any measurement points for the electrical energy or power.

0.00

5.00

10.00

15.00

20.00

25.00

30.00

Jan_

08

Feb_0

8

Mar

ch_0

8

April_

08

May

_08

June

_08

July_

08

Aug_0

8

Sept_

08

Oct_08

Nov_0

8

Dec_0

8

Ele

ctric

al p

ower

con

sum

ptio

n [k

W] -

Low

pre

ssur

e [b

ar]

1.00

1.50

2.00

2.50

3.00

3.50

Pre

ssur

e ra

tio [-

]

Compressor Power Measured Compressor Power Calculated Low pressure Pressure ratio

Figure 4.4: Electrical power consumption, comparison with the two methods for a single stage CO2 system during the whole year 2008, KA1 unit in the TR1 Supermarket.

Note that in all our calculations and simulations, we use the energy consumption of the

compressors and for indirect systems we add also the energy consumption of the brine pumps.

The power of the pumps was evaluated using the nominal power of the pumps, as we do not

have any energy measurements. Defrost heater, fans, lighting of the cabinets are not included.


David Freléchox

26

4.3 Mass flow evaluation

The mass flow measurement is always a difficult process and is generally a key factor to obtain

good results. None of the studied supermarkets had any mass flow measurement point. So we

used a method based on the pressure and temperature measures at the compressor inlet to get

the specific volume. We used compressor data [DOR09] and [BIT09] to obtain the swept

volume, which is given as a fixed value in m3/h when the compressor is running under 50 Hz

and the volumetric efficiency in function of the pressure ratio shown on Figure 4.5. The swept

volume multiplied by the volumetric efficiency could be se as the volumetric flow through the

compressor. In order to calculate the mass flow with Equation 4-1, the state (pressure and

temperature) of the fluid at the compressor inlet were used to define the specific volume.

incomp

SVCO v

Vm

_2

&

&⋅

=η

4-1

);(]/[

]/[

][

___3

_

3

incompincompabsstateincomp

S

V

TPfkgmvolumespecificv

datacompressoronbasedsmvolumesweptV

fitteddatacompressoronbasedefficiencyvolumetric

==

=

−=&

η

TCS373-D = -0.4079x2 - 6.5843x + 102.42

TCDH372= 0.0251x2 - 1.1706x + 93.424

SCS 362 SC = -0,1139x2 - 4,1854x + 95,12

0

10

20

30

40

50

60

70

80

90

100

0.00 2.00 4.00 6.00 8.00 10.00 12.00

Pressure ratio [-]

Vol

umet

ric e

ffici

ency

[%]

TCS373-D

TCDH372 B-D

SCS 362 SC

Figure 4.5: Volumetric efficiency based on compressor data for three CO2 compressors


David Freléchox

27

The following Figure 4.6 shows the variation of the CO2 mass flow during one day in July 2009

in the freezer system of the TR1 supermarket using the method based on the volumetric

efficiency.

0

0.05

0.1

0.15

0.2

0.25

30.06.200819:12

01.07.200800:00

01.07.200804:48

01.07.200809:36

01.07.200814:24

01.07.200819:12

02.07.200800:00

02.07.200804:48

CO

2 m

ass

flow

[kg/

s]

-25

-20

-15

-10

-5

0

5

10

15

20

Tem

pera

ture

[°C

] - P

ress

ure

[bar

]

CO2 mass flow Compressor inlet temperature Compressor inlet pressure

Figure 4.6: Mass flow of CO2 in the freezer system FA1 during one day of July 2008 in the TR1 supermarket

As can be seen on the figure, the compressor inlet conditions are unstable mainly depending on

the cooling capacity used in the cabinets and also the control of the internal superheat by the

expansion valve, thus the compressor inlet temperature could vary quite a lot. The volume flow

is quit constant because the pressure ratio is stable and the only things which affected it is the

number of compressor working. But the compressor inlet conditions of the fluid vary and affect

the stability of the mass flow. Thus when only one compressor is working the mass flow of

refrigerant could vary between 0.06 and 0.1 kg/s. In this case one or two compressors could be

working. When the mass flow is above 0.1 kg/s then the second compressor has started

working.


David Freléchox

28

We chose this method based on the volumetric efficiency after several tests in order to apply

the method that we consider the most reliable. We have carried out several researches to find

comparable methods in the literature. To assess the reliability of our method, we made various

comparisons of which we present in the Figure 4.7 below.

0.00

0.05

0.10

0.15

0.20

0.25

0.30

0.35

0.40

0.45

Jan_

08

Feb_0

8

Mar

_08

Apr_0

8

May

_08

Jun_

08

Jul_0

8

Aug_0

8

Sep_0

8

Oct_08

Nov_0

8

Dec_0

8

Jan_

09

Feb_0

9

Mar

_09

Mas

s flo

w C

O2

[kg/

s]

0.0

0.5

1.0

1.5

2.0

2.5

Pre

ssur

e ra

tio [

-], E

ta_t

ot [-

]

mCO2 ηvol mCO2 Dabiri mCO2 15%Oil cooler Pressure Ratio Eta tot

Figure 4.7: Mass flow of CO2 in a transcritical system for different mass flow measurement method

The first comparative method is the Dabiri’s method based on an article proposed by Dabiri and

Rice [DAB82]. Here, it is briefly summarized, firstly through Equation 4-2 which makes a ratio

between design (map) conditions and actual (new) conditions:

−⋅+= 11

map

new

map

new Fm

m

ρρ

&

& 4-2

Where F is a chosen percentage of the theoretical mass flow rate increase and where the

densities are evaluated based on suction port conditions. F = 0.75 is usually used.


David Freléchox

29

This method is difficult to apply because of the proposed correction factor is the result of

experience with R22 and the experience is from 1982. Nonetheless, it has recently been used in

laboratory test and gave satisfaction.

The second comparative method is based on the energy balance around the compressor

according Equation 4-3. The compressor can be seen as a black box and the method is to do a

simple energy balance.

cooleroillossescompel QQhmE &&&& ++∆⋅= 4-3

The electrical consumption is measured and the enthalpy before or after the compressor is

given from pressures and temperatures at the compressor inlet and outlet. Based on general

experience and manufacturer information the heat losses are about 7% of electrical input and

the oil cooler losses are about 15%. This last value does not seem to be a fix value as the oil

cooler losses are affected from many parameters as the air or water inlet temperature and the

pressure ratio of the compressor.

The Figure 4.7 shows differences between the three proposed methods. We chose to use the

first method based on compressor data because it seems the most reliable method. It is less

dependent on external parameters than the others. The method of Dabiri is difficult to apply

because of the use of a correction factor which is unreliable, particularly when we do not know

the bases of this correction. Moreover, it seems to be very responsive to the pressure ratio and

suffered large fluctuations. The evaluation of mass flow by the energy balance around the

compressor uses fixed percentages of losses although the dissipated energy by the oil cooler

fluctuates. Eta tot is the total efficiency of the compressors including heat losses, oil coolers

losses, isentropic losses, volumetric losses. Its value is around 0.6. The method we chose

allows to calculate the heat dissipation in the oil cooler and to improve the technical knowledge

of this item.


David Freléchox

30

The Figure 4.8 below gives an overview of the effects of these various methods on our final

objective, the COP calculation. Again, our method based on the volumetric efficiency, COPηvol,

gives satisfactory results. It correlates very well with the COP resulting from the use of losses of

15% through the oil cooler, as well. In contrast, the method proposed by Dabiri and Rice seems

doubtful. Indeed, we do not notice a real correlation with the pressure ratio while we know its

importance on the efficiency of a system. The decrease of the pressure ratio in November 2008

does not really increase the COP which is unlikely.

0.0

1.0

2.0

3.0

4.0

5.0

6.0

Jan_

08

Feb_0

8

Mar

_08

Apr_0

8

May

_08

Jun_

08

Jul_0

8

Aug_0

8

Sep_0

8

Oct_08

Nov_0

8

Dec_0

8

Jan_

09

Feb_0

9

Mar

_09

CO

P [-

], P

R [-

]

COP ηvol COP Dabiri COP 15%oil cooler Pressure Ratio

Figure 4.8:COP of a CO2 transcritical system for different mass flow measurement method


David Freléchox

31

4.4 COP calculation

Eventually, the value that we are interested in is the coefficient of performance of the system or

COP. This value gives information about the efficiency of each system, thus provided,

comparing them at identical operating conditions. The COP of a refrigeration system is

calculated using the following Equation 4-4:

nconsumptiopowerElectrical

capacityCooling

E

QCOP

comp

oinst ==

&

&

. 4-4

We want to have a single value for the whole cooling system and thus the Equation 4-5:

)( ___

__

brinepumpchillercompfreezercomp

chillerofreezerotot EEE

QQCOP

&&&

&&

+++

= 4-5

A COP for the booster system must also be calculated. Since the high stage compressors and

the booster compressors are located in different places in the system it is possible to calculate

two mass flows. One mass flow is the total mass flow going through the high stage compressors

and one mass flow is the mass flow maintaining the freezers. A mass balance can be applied to

calculate the mass flow going trough the medium temperature cabinets, see equation 4-6.

freezertotalchiller mmm &&& −=

4-6

This mass flow and the pressure and temperature measurements allow calculating the power of

each part of the system. Thus, the total COP of the booster system could be calculated in

Equation 4-7. Only the cooling capacity from the freezer side and the capacity from the medium

temperature side which goes to the medium temperature cabinets is taken into account. The

medium temperature power used for the condensation on the freezer side is eliminated.

chillercompfreezercomp

freezercchillerofreezeroboostertot EE

QQQCOP

__

____ &&

&&&

+−+

= 4-7

For a cascade system, with the mass flow and the temperatures and pressure it is possible to

calculate the cooling capacity of the R404A- and CO2-units.

oo hmQ ∆⋅= &&

4-8

Where ∆ho is the enthalpy difference over the evaporator.


David Freléchox

32

The condenser load of the CO2-unit can be calculated with Equation 4-9.

shaftfreezercompfreezerofreezerc EQQ ____&&& +=

4-9

To decide the load of the medium temperature side cabinets Equation 4-10 are used.

freezercchillerocabo QQQ ___&&& −=

4-10

The electrical energy from the chiller which goes to the freezer can be calculated by Equation 4-11.

chillercompchillero

freezercfreezerforchiller E

Q

QE _

_

_ &&

&& ⋅=−− 4-11

The COP for the freezers can be calculated by Equation 4-12.

freezerpumpsfreezerforchillerfreezercomp

freezerofreezer EEE

QCOP

__

_

&&&

&

++=

−−

4-12

The COP for the chillers can be calculated by Equation 4-13.

cabpumpsfreezerforchillerchillercomp

cabochiller EEE

QCOP

__

_

&&&

&

+−=

−−

4-13

Where opumps QE && ⋅= %4 [GRA07]

To compare the concepts between them, the load ratio has to be identical, i.e. the ratio of the

cooling capacity between the chiller and the freezer is the same for each installation. An

approximate value for European supermarket is 3, so 3 times more cooling capacity for medium

temperature cabinets than for low temperature cabinets. In order to correct our COP according

to a fix load ratio (LRcorr), we developed the Equation 4-14. It is possible to see the

demonstration in the Appendix 1. The abbreviation of load ratio is LR, thus COPtot_LR is the

total COP of system with a defined load ratio LRcorr.

chillercompfreezero

chillerofreezercomp

corr

corr

corrchillero

LRtot

EQ

QE

LR

LR

LRQ

COP

__

__

_

_1

1

&&

&&

&

⋅+⋅

+⋅= 4-14

Note that the COP is the instantaneous efficiency of the installation. We calculated it for each

measurement interval (5 or 15 minutes). Then we did an average to get a monthly value. It may

slightly differ from the monthly COP which is a ratio of energy rather than power.


David Freléchox

33

5 Field installations

In this thesis two transcritical systems and one cascade system will be evaluated. The

supermarkets will be named transcritical system 1 and 2 (TR1 and TR2) and cascade system 1

(CC1). They are located in different places in Sweden meaning that the outside air temperature

will be different for each supermarket. TR1 is the most northern and TR2 is the most southern

supermarket, CC1 is rather centrally located.

5.1 Supermarket with transcritical system TR1

The TR1 supermarket has been open since autumn 2007. The maximal cabinet design cooling

load is 230 kW for cold products and 60 kW for frozen products. There are four separated

transcritical units, two for the medium temperature cabinets and two for the low temperature,

with an indirect water-glycol system for the heat rejection. The nearest weather station to the

supermarket is Storön.

Figure 5.1 represents a chiller unit installed in the TR1 supermarket; three compressors are

visible at the bottom of this unit. They produce the cooling capacity for the medium temperature.

The 4th compressor is barely visible behind the electrical panel. On each compressor the oil

cooler can be distinguished, oil heat is transferred to the coolant.

Figure 5.1: Freezer unit in TR1 Supermarket


David Freléchox

34

Figure 5.2 shows:

� Two Coolers

o Transcritical CO2, single-stage / Compressor four Dorin TCS 373-D

o Oil cooler

o Heat recovery

o Coolant

� Two Freezers

o Transcritical CO2, two-stages with intercooler / Compressor:

two Dorin TCDH 372 B-D

o Oil cooler

o Heat recovery

o Coolant

Figure 5.2: Schematic diagram of the TR1 system

The system is a parallel solution where there are two separate carbon dioxide circuits, one for

the medium temperature side (KA1/KA2) and one for the cold temperature side (FA1/FA2). A

benefit from using a parallel solution is that if one of the cycles fail, the other cycle can

unaffectedly continue to work (Sawalha, 2008).


David Freléchox

35

The cold temperature side, seen to the right in Figure 5.2, has a two-stage compression with an

intercooler in between. This is arranged to achieve cold temperatures but still keep low pressure

ratios in the compressors. This will lower the inlet temperature to the second compressor,

decrease the discharge pressure after the second stage and decrease the losses, thereby

increase the efficiency of the system. The carbon dioxide is condensed in the condenser and

expanded in the expansion valve before entering the evaporator (freezers). The expansion

valves are placed out in the supermarkets close to the evaporators. The reason is to minimize

the losses in the system by transporting the refrigerant with high pressure. After the freezer the

refrigerant return to the machinery room and enters a liquid separator (LS) before the

compressors. This is done to make sure that no liquid is going in to the compressors. There are

two units for the cold temperature side (FA1 and FA2). Each unit has two two-stage

compressors.

The medium temperature side has a one-stage compressor since it doesn’t need to operate

with as high pressure ratio as the cold temperature side to maintain the chillers. After the

condenser the refrigerant is expanded in the expansion valve where the pressure is reduced,

before entering the evaporator (chillers). For the same reasons as in the FA-units, the

expansion valves are placed in the Supermarket area close to the cabinets. There are two units

for the medium temperature side (KA1 and KA2). There are four one-stage compressors in

every unit.

The refrigerant in both cycles is gas cooled by a brine circulating between the main condenser

and the two CO2-cycles. The cold brine is used for the oil coolers and the condensers/gas

coolers in both circuits and for the intercooler in the cold temperature side, see Figure 5.2. The

brine condenser is placed on the roof and is using the outside air temperature to cool down the

brine. There is an additional heat exchanger in the brine circuit, placed before the condenser,

for maintaining a heat pump that is supplying the supermarket with air conditioning and heating.

[JOH09]


David Freléchox

36

5.2 Supermarket with transcritical system TR2

The supermarket TR2 has been open since august 2008. The maximal cabinet design cooling

load is 200 kW for cold products and 50 kW for frozen products. There are three separated

transcritical units, two booster types for the medium temperature cabinets and the low

temperature cabinets with a load ratio of about 2, and one standard-type for the rest of the

medium temperature cabinets, with a direct system for the heat rejection. The nearest weather

station to the supermarket is Goeteborg.

Figure 5.3 represents a booster unit installed in the TR2 supermarket. Three compressors are

visible at the bottom of this unit. They produce the cooling capacity for the medium temperature.

The two compressors for the low temperature are behind the electrical panel. On each

compressor an air cooled oil coolers can be distinguished. On top of the large tanks, there are

three valves to avoid overpressure in the system. The 3 tanks are used as receiver and oil

separator.

Figure 5.3: Booster unit in TR2 Supermarket


David Freléchox

37

Figure 5.4 shows :

� Two Boosters

o Transcritical CO2, two-stage intercooling booster

o Compressor: two Dorin SCS 362 (LS), three Dorin TCS 373 (HS)

o Oil cooler

o Heat recovery

o Subcooling from ground heat sink

o Gas cooler on the roof

� Single Standard

o Transcritical CO2, single-stage / four Dorin TCS 373

o Oil cooler

o Heat recovery

o Subcooling from ground heat sink

o Gas cooler on the roof

Figure 5.4: Schematic diagram of the TR2 system


David Freléchox

38

In this supermarket there are one circuit for the medium temperature side (KA3) and one circuit

for a combined medium and cold temperature side (KAFA1/KAFA2). There are two units for the

combined side (KAFA1 and KAFA2) and one unit for only medium temperature cabinets (KA3).

The KA3 cycle can be seen to the right in figure 5.4 and is similar to the KA-unit in trans-critical

system 1. After the evaporator (chillers) the refrigerant enters the one-stage compressor. An

extra heat exchanger is placed after the compressor to recover heat to floor and space heating

of the supermarket. The refrigerant is after that gas cooled/condensed in the gas cooler. The

gas cooler is placed on the roof and uses the outside air temperature to cool down the

refrigerant. Before the refrigerant reaches the expansion valve an extra heat exchanger is

placed to further cool down the refrigerant and gain some additional heat recovery. This heat

exchanger uses a ground heat source for heat exchange with the carbon dioxide.

The combined circuit (KAFA1/KAFA2) side can be seen to the left in figure 5.4 and it serves

both medium temperature cabinets and freezers. The gas cooler is placed on the roof and uses

the outside air to cool down the refrigerant. After the gas cooler/condenser the CO2 runs

through an additional heat exchanger for heat recovery, which also uses the same ground heat

source as in KA3. The ground heat source is used for heating the supermarket. The mass flow

of the refrigerant is separated before it reaches the expansion valves and cabinets/freezers.

After the freezers two compressors called “booster compressors” are located. They increase the

pressure of CO2 to the same pressure as the CO2 has during the evaporation in the medium

temperature cabinets. The mass flows from the medium temperature cabinets and from the

freezers are mixed in the liquid separator before the high stage compressors. The high stage

compressors raise the pressure of the CO2 to condensing pressure. The refrigerant runs

through an additional heat exchanger, for floor and space heating as in the case of KA3-unit,

before it is gas cooled/condensed in the gas cooler. [JOH09]


David Freléchox

39

5.3 Supermarket with cascade system CC1

The supermarket CC1 has been open since 2006, but the measured data are only available

since December 2008. The cooling load is 220 kW for cold products and 60 kW for frozen

products. It is a cascade R404A / CO2 system, R404A for the first stage, brine for the medium

temperature cabinets and CO2 DX for the low temperature cabinets, with an indirect water-

glycol system for the heat rejection. The nearest weather station to the supermarket is Floda.

Figure 5.5 presents two CO2 low temperature units in the CC1 supermarket. Both units are

composed of four compressors. The condensation capacity is transmitted to the brine circuit

through plate heat exchangers. If the installation should be stopped, a small refrigeration unit

(on top of each unit) maintains the CO2 at proper temperature and pressure so safety valves

are not activated.

Figure 5.5: Two CO2 low temperature units in the CC1 supermarket


David Freléchox

40

Figure 5.6 shows:

� Two R404A DX units (stage 1)

o Three compressor: Bitzer 4H-15.2Y-40P

o Internal heat exchanger (IHE)

o Heat recovery

o Coolant

� Single Brine loop (intermediate stage)

o Brine 34% glycol

o Pumped

� Two CO2 DX units (stage 2)

o Four Compressor Bitzer 2KC-3.2K-40S

o CO2 subcritical

o Internal heat exchanger (IHE)

Figure 5.6: Schematic diagram of the cooling system in the supermarket CC1


David Freléchox

41

The system solution is a cascade solution with R404A in the high stage and CO2 in the low

stage. There is direct expansion with CO2 for the freezers and indirect for the chillers. Figure 5.6

shows a schematic picture of the system. There are two units of the low stage (KS5 and KS6)

and two units for the high stage (VKA1 and VKA2). There is only one brine circuit.

R404A in the high stage are condensed by a coolant that is heat exchanging with the outside

air. Before the condenser a desuperheater are located for reuse some of the heat after the

compressor. A subcooler, using the coolant, is placed after the condenser for subcooling the

fluid. There is an internal heat exchanger (IHE) in the system where the refrigerant is further

subcooled by transfer heat to the refrigerant after the evaporator. After the evaporator and IHE

the refrigerant enters the compressors before returning to the desuper heater. There are two

units of the high stage R404A and three compressors in every unit.

The brine evaporating the R404A is cooling the medium temperature cabinets and is circulated

by pumps. The brine is condensing the CO2, used as a refrigerant, in the low stage. After the

condenser the CO2 is heat exchanging in an IHE to be subcooled before the expansion valve

and the freezers. After the freezers the refrigerant enters the IHE before it enters the

compressors and then back to the condenser. There are two units of the low stage CO2 and four

compressors in every unit. [JOH09]


David Freléchox

42

6 General system analysis

This chapter includes an analysis of each supermarket on the basis of collected data and also

various figures showing the behaviour of each system during the observation period. This

mainly shows the cooling capacity, the power consumption and the COP for each system.

6.1 Supermarket TR1

The Figure 6.1 shows the cooling capacity for the medium and low temperature units KA1 and

FA1. A peak of consumption appears during the summer. This increase is particularly visible on

the medium temperature unit. Freezers most of which are fitted with glass doors, are less

responsive to ambient conditions.

0.00

10.00

20.00

30.00

40.00

50.00

60.00

70.00

Jan Feb Mar Apr May Jun Jul Aug Sep Oct Nov Dec

Coo

ling

capa

city

[kW

]

-20.00

-15.00

-10.00

-5.00

0.00

5.00

10.00

15.00

20.00

25.00

30.00

Tem

pera

ture

[°C

]

KA1 cooling capacity 2008 KA1 cooling capacity 2009 FA1 cooling capacity 2008FA1 cooling capacity 2009 Outdoor temperature 2008 Outdoor temperature 2009

Figure 6.1: Cooling capacity of one medium temperature unit (KA1) and one low temperature unit (FA1) during the years 2008 - 2009

The plot on Figure 6.1 is divided in curves for 2008 and 2009 because the system seems to

have different control schemes during these periods. Since the end of 2008 the limit of the

floating condensation was lowered. The elevated consumption during January and February

2008 on KA1 is linked to the commissioning of the cooling system. The installation was still in a

settings stage. From 2008 to 2009 the load falls, while external conditions are almost identical

and that the layout of the store has not changed. To our knowledge, no changes have been

made on cabinets, the load should not vary. However, several external parameters may explain

this decrease as decrease in a customer’s numbers or an adjustment of the regulation on the

HVAC system.


David Freléchox

43

Some improvements on the system after the summer 2008 can also play a role in this

development. The setting on the condensers’ coolant temperature was lowered which led to a

COP improvement and thus reduced the compressors’ power consumption. This trend is clearly

visible on Figure 6.2 below. The evaporation temperature has been rather constant all the way

around -10°C for the medium temperature units and - 35°C for the low temperature units.

0.00

5.00

10.00

15.00

20.00

25.00

Jan Feb Mar Apr May Jun Jul Aug Sep Oct Nov Dec

Ele

ctric

al p

ower

[kW

]

0.00

5.00

10.00

15.00

20.00

25.00

30.00

35.00

40.00

Tem

pera

ture

[°C

]

KA1 comp electrical consumption 2008 KA1 comp electrical consumption 2009FA1 comp electrical consumption 2008 FA1 comp electrical consumption 2009Coolant temperature 2008 Coolant temperature 2009

Figure 6.2: Compressors electrical power consumption for one medium temperature unit (KA1) and one low temperature unit (FA1) during the years 2008 - 2009

The modification of the coolant temperature is particularly important. Its effect is clear on the

consumption curve of FA1. Just after the change during August 2008, the power consumption

decreases. To highlight the impact of coolant temperature on the COP, we present the Figure

6.3.

The data based on the field measurements show a clear correlation between the coolant

temperature at the entrance of the condenser / gas cooler and the performance of the system.

The impact on the COP of decreasing the coolant temperature is more important on medium

temperature unit. This is evidently because of its lower pressure ratio.


David Freléchox

44

y = 0.0058x2 - 0.2932x + 6.3315

y = 0.0024x2 - 0.0957x + 2.2924

0.00

0.50

1.00

1.50

2.00

2.50

3.00

3.50

4.00

4.50

5.00

0.00 2.00 4.00 6.00 8.00 10.00 12.00 14.00 16.00 18.00 20.00

Coolant temperature [°C]

CO

PCOP_chiller_KA1 COP_chiller_KA2COP_freezer_FA1 COP_freezer_FA2Polynomial (COP_chiller_KA1) Polynomial (COP_freezer_FA2)

Figure 6.3: COP function of coolant temperature for medium temperature units and low temperature units, measures for TR1 supermarket during 2008.

Finally the Figure 6.4 shows the COP of each unit for the whole test period. The low

temperature units FA 1 and 2 show small changes in function of the ambient conditions and

also following the modification of the coolant temperature. In contrast, the medium temperature

COP of the KA 1 and 2 units can vary from 2.8 to 4.5. This is the result of the use of the floating

condensation which considerably increases the COP during the winter. From winter 2008 to

winter 2009, the COP was improved of about 25 % following the lowering of the coolant

temperature which was reduced from 12 to 7 K.

0.00

0.50

1.00

1.50

2.00

2.50

3.00

3.50

4.00

4.50

5.00

Jan_

08

Feb_0

8

Mar

_08

Apr_0

8

May

_08

Jun_

08

Jul_0

8

Aug_0

8

Sep_0

8

Oct_08

Nov_0

8

Dec_0

8

Jan_

09

Feb_0

9

Mar

_09

Apr_0

9

May

_09

Jun_

09

CO

P

COP KA1 COP KA2 COP FA1 COP FA2

Figure 6.4: COP for each units during the whole testing period for the TR1 supermarket.


David Freléchox

45

6.2 Supermarket TR2

A figure per unit has been achieved, KAFA 1 and 2 are booster units and KA3 is a medium

temperature unit. These figures show the evolution of the cooling power and the related power

consumption, the curves of condensing and outdoor temperature, as well as the effect of

subcooling produced by the borehole. This effect is expressed by its ∆T in Kelvin. Note that the

borehole is connected to the heat pump as well and used as heat source to heat the building

during the winter.

0.00

10.00

20.00

30.00

40.00

50.00

60.00

70.00

Sept_

08

Oct_08

Nov_0

8

Dec_0

8

Jan_

09

Feb_

09

Mar

ch_0

9

April_

09

May

_09

June

_09

Coo

ling

capa

city

[kW

] - E

lect

rical

pow

er [k

W]

0.00

5.00

10.00

15.00

20.00

25.00

30.00

Tem

pera

ture

[°C

] - d

T [K

]

Cooling capacity Compressor electrical consumptionCondensation temperature dT subcooling boreholeOutside temperature

Tevap = -10°C / -35°C

Figure 6.5: Different parameters plots for the KAFA1 unit during the whole period of study in the TR2 supermarket

Figure 6.5 shows, as expected a cooling capacity drop during the winter. In contrast the power

consumption does not follow the same trend, although logically it should take advantage of low

winter temperatures. The cause is simply forcing the condensing temperature at about 25°C in

order to increase the capacity for the heat recovery system. The refrigeration system and heat

pump are connected via the borehole but also on the "warm" side through a plate heat

exchanger disposed on the high pressure circuit at the compressor exit.

To compensate this rise of the condensation temperature and in order to maintain the COP at a

high level, the borehole is used to subcool the fluid. Note that the higher the condensing

temperature is kept, the greater is the ∆T subcooling. The fact that the heat pump for heating

the store is also connected to the borehole can justify this principle of operation as the rejected

heat by the subcooling can then be used by the heat pump. Similar parameters plots as in the

previous figure have been developed for KAFA2 in Figure 6.6 and for KA3 in Figure 6.7.


David Freléchox

46

0.00

10.00

20.00

30.00

40.00

50.00

60.00

70.00

Sept_

08

Oct_08

Nov_0

8

Dec_0

8

Jan_

09

Feb_0

9

Mar

ch_0

9

April_

09

May_0

9

June

_09

Coo

ling

capa

city

[kW

] - E

lect

rical

pow

er [k

W]

0.00

5.00

10.00

15.00

20.00

25.00

30.00

Tem

pera

ture

[°C

] - d

T [K

]


Tevap = -10°C / -35°C

Figure 6.6: Different parameters plots for the KAFA2 unit during the whole period of study in the TR2 supermarket

The observations on KAFA2 are similar to that for the unit KAFA1. The cooling capacity

produced is lower, although the units are identical. After the starting period (Sept - Oct) and also

through the significant use of the subcooling, the electricity consumption could be reduced.

0.00

10.00

20.00

30.00

40.00

50.00

60.00

70.00

Sept_

08

Oct_08

Nov_0

8

Dec_0

8

Jan_

09

Feb_0

9

Mar

ch_0

9

April_

09

May

_09

June

_09

Coo

ling

capa

city

[kW

] - E

lect

rical

pow

er [k

W]

0.00

5.00

10.00

15.00

20.00

25.00T

empe

ratu

re [°

C] -

dT

[K]


Tevap = -10°C

Figure 6.7: Different parameters plots for the KA3 unit during the whole period of study in the TR2 supermarket


David Freléchox

47

The study of the KA3 unit confirms our previous observations. The Figure 6.8 shows the trend

of the COPs and their slight decrease during the winter period related to the high condensing

temperature. The use of subcooling does not seem to compensate completely for these losses.

The differences on the 2 booster units are linked to the missing of separate energy

measurement for the medium temperature and low temperature compressors on the unit

KAFA2. Some assumptions have been made for KAFA-units to be able to perform these

calculations. Before January there was only data available of the total energy that goes o the

KAFA-unit and no separate measurement of the energy that goes to the booster compressors

was done. From January the measurement of the energy to the high stage and booster

compressors are separated for KAFA1. Based on that information an average of the energy that

goes to the booster compressors of the total power consumption of the compressors was

estimated. This was used to perform calculations of COP and cooling capacity for the months

prior to the separate energy measurements on KAFA1 and for all the month for KAFA2.

0.00

0.50

1.00

1.50

2.00

2.50

3.00

3.50

4.00

4.50

5.00

Sept_

08

Oct_08

Nov_0

8

Dec_0

8

Jan_

09

Feb_0

9

Mar

ch_0

9

April_

09

May

_09

June

_09

CO

P

COP KAFA1 COP KAFA2 COP KA3

Figure 6.8: COP for each units during the whole testing period for the TR2 supermarket.

As can be seen on the figure, the COP of the booster systems are around 2.5 and the COP of a

medium temperature unit is around 4.


David Freléchox

48

6.3 Supermarket CC1

Figures of the main parameters of the two VKA medium temperature units with R404A and the

two KS low temperature units with CO2 were developed, Figure 6.9 and Figure 6.10

respectively. The figures show the evolution of the cooling capacity and the related power

consumption and also the curves of condensation and outside temperature.

0.00

20.00

40.00

60.00

80.00

100.00

120.00

Dec_08 Jan_09 Feb_09 Mar_09 Apr_09 May_09 Jun_09

Coo

ling

capa

city

[kW

] - E

lect

rical

pow

er [k

W]

-10.00

0.00

10.00

20.00

30.00

40.00

50.00

60.00

70.00

80.00

Tem

pera

ture

[°C

]

Compressor electrical power VKA1 Compressor electrical power VKA2Cooling capacity VKA1 Cooling capacity VKA2Condensation temperature VKA1 Condensation temperature VKA2Outside temperature

Tevap = -11°C

Figure 6.9: Cooling capacity, compressor electrical power consumption, condensation and outside temperatures for medium temperature units VKA1 and VKA2 during the whole testing period for the CC1 supermarket

Globally the Figures 6.9 and 6.10 demonstrate an essential fact of CC1 system, the stability of

its operating parameters. The condensing temperature is permanently kept at a high level. The

lower limit of floating condensing is set at 30°C s o the monthly average temperature does not

fall below this value. Even the increase of the outside temperature does not really affect the

condensation level. The cooling capacity was slightly lowered during the winter months like

January and February.

The analysis of this supermarket does not raise any significant changes or developments. The

only question mark is the justification for maintaining the condensing temperature as high, even

if the coolant circuit is connected to HVAC system and allows heat recovery in winter.


David Freléchox

49

0.00

5.00

10.00

15.00

20.00

25.00

30.00


Coo

ling

capa

city

[kW

] - E

lect

rical

pow

er [k

W]

-10.00

-5.00

0.00

5.00

10.00

15.00

Tem

pera

ture

[°C

]

Cooling capacity KS5 Cooling capacity KS6Compressor electrical power KS5 Compressor electrical power KS6Condensation temperature KS5 Condensation temperature KS6Outside temperature

Tevap = -36°C

Figure 6.10: Cooling capacity, compressor electrical power consumption, condensation and outside temperatures for low temperature units KS5 and KS6 during the whole testing period for the CC1 supermarket

The comparison of units’ COP on Figure 6.11 does not really give much of variations. The

medium temperature unit’s COP is slightly decreased approaching the summer period. The

COPs of the low temperature units are constant due to the rather constant condensing and

evaporating temperatures.

0.00

0.50

1.00

1.50

2.00

2.50

3.00

3.50


CO

P

COP VKA1 COP VKA2 COP KS5 COP KS6

Figure 6.11: COP for each units during the whole testing period for the CC1 supermarket.


David Freléchox

50

6.4 Comparison of the three systems

In this chapter the three supermarkets previously studied will be compared. In order to achieve

a fair comparison, it should also show the evolution of parameters such as the load ratio or the

condensing temperature during the period of analysis.

The load ratio defined the ratio of cooling capacity between the medium temperature and the

low temperature cabinets. This is different for each supermarket and may also change during

the year. The medium temperature cabinets are generally more sensitive to the outside

temperature and humidity conditions during the summer. That is the explanation why the load

ratio increases during this period.

The Figure 6.12 shows the evolution of the load ratio for each supermarket. CC1 is rather

constant. TR1 is slightly affected by the summer period and TR2 is very high during the starting

phase of the system before stabilizing just above a value of 3. We can see that the value of the

load ratio of each supermarket moves relatively close to the value of 3, which is what would be

expected in a Swedish supermarket.

0.00

0.50

1.00

1.50

2.00

2.50

3.00

3.50

4.00

4.50

5.00

Jan_

08

Feb_0

8

Mar

_08

Apr_0

8

May

_08

Jun_

08

Jul_0

8

Aug_0

8

Sep_0

8

Oct_08

Nov_0

8

Dec_0

8

Jan_

09

Feb_0

9

Mar

_09

Apr_0

9

May

_09

Jun_

09

Load

rat

io

LR TR1 LR TR2 LR CC1

Figure 6.12: Load ratio for the three systems analysed during each period of analysis


David Freléchox

51

The condensing temperature is also an essential factor to take into consideration for the

analysis of the different COPs. Average condensing temperatures for the three systems

analyzed are plotted in Figure 6.13.

0.00

5.00

10.00

15.00

20.00

25.00

30.00

35.00

40.00

Jan_

08

Feb_0

8

Mar

_08

Apr_0

8

May_0

8

Jun_

08

Jul_0

8

Aug_0

8

Sep_0

8

Oct_08

Nov_0

8

Dec_0

8

Jan_

09

Feb_0

9

Mar

_09

Apr_0

9

May_0

9

Jun_

09

Con

dens

atio

n T

empe

ratu

re [°

C]

Condensation temperature TR1 Condensation temperature TR2 Condensation temperature CC1

Figure 6.13: Condensation temperature for the three systems analysed during each period of analysis

As can be seen in the figure, main differences exist among the three systems. While the system

TR1 seems to run properly at floating condensation since September 2008, the supermarket

TR2 also functioning with CO2 maintains its condensing temperature at high level in order to

recover heat during the winter. The lower condensing temperature of TR2 system than TR1

during the month of June 2009, even though TR2 is much further south, is may be due to water

spraying on the gas coolers.

The temperature of condensation of CC1 system can be up to 20°C higher than the two other

systems, so the CC1 installation is greatly disadvantaged compared to the other two in terms of

energy efficiency. This is mainly related to the system control, not necessarily the design or

solution. The measurements made on the system do not indicate that heat is being recovered,

however, the system has been built in order to recover heat but the heat recovery system had

some technical problems and the system control has been kept to run as if heat is being

recovered.

Figure 6.14 shows the different levels of COP for the medium temperature - chiller units- and

the low temperature – freezer units.


David Freléchox

52

0.00

0.50

1.00

1.50

2.00

2.50

3.00

3.50

4.00

4.50

5.00

Jan_

08

Feb_0

8

Mar

_08

Apr_0

8

May

_08

Jun_

08

Jul_0

8

Aug_0

8

Sep_0

8

Oct_08

Nov_0

8

Dec_0

8

Jan_

09

Feb_0

9

Mar

_09

Apr_0

9

May

_09

Jun_

09

CO

P

COP chiller TR1 COP chiller TR2 COP chiller CC1 COP freezer TR1COP freezer TR2 COP freezer CC1

Figure 6.14: Condensation temperature for the three systems analysed during each period of analysis

As can be seen in the figure, significant variations were mostly observed on medium

temperature units, but the percentage of changes may also reach 20% on low temperature

systems.

The impact of low condensing temperature is particularly visible on the curve chiller COP TR1

during winter 2009. The proper use of floating condensing in this supermarket can achieve a

COP greater than 4.5 on the chiller and greater than 1.6 on the freezer.

The objective of recovering a maximum of heat during the winter in the TR2 system is rather

negative for its COP. On the other hand the missing of a coolant loop on the condenser/gas

cooler allows it to get the best COP for both chiller and freezer during warm periods. Evidently,

the COP of the chillers and freezers in supermarket CC1 are the lowest mainly due to the high

condensing temperature.


David Freléchox

53

6.5 Comparison of the three systems with a load ratio o f 3

The use of Equation 4-14 allows correcting the COP in function of the load ratio and thus to

obtain the COP equivalent to a load ratio of 3 for each system. We have made this correction on

the total COP for each supermarket including medium and low temperature. The Figure 6.15

shows the total COP for each system with and without the load ratio correction.

2.00

2.20

2.40

2.60

2.80

3.00

3.20

3.40

Jan_

08

Feb_0

8

Mar

_08

Apr_0

8

May

_08

Jun_

08

Jul_0

8

Aug_0

8

Sep_0

8

Oct_08

Nov_0

8

Dec_0

8

Jan_

09

Feb_0

9

Mar

_09

Apr_0

9

May

_09

Jun_

09

CO

P to

tal

COPtot TR1 jan08-aug08 COPtot TR2 COPtot CC1 COPtot TR1 LR3 COPtot TR2 LR3 COPtot CC1 LR3

Figure 6.15: Load ratio correction for the three systems during the whole period of analysis

As can be seen in the figure, the correction of the load ratio slightly reduces the gap between

the 2 transcritical supermarkets and the supermarket using the cascade. The three cases had a

load ratio close to 3 for the period from January 2009 and on, so the impact of this correction is

less significant.

Figure 6.16 compares each COP according to their respective condensing temperature. A main

reason why the COP obtained with the TR2 system could be higher than TR1 is due to the use

of the borehole to subcool the refrigerant. The stability of CC1 operation results in a small cloud

point above 30°C. Observating the trend of TR1 allo ws establishing that at such high

temperature of condensation cascade system will be more efficient than the TR1 system.


David Freléchox

54

1.00

1.50

2.00

2.50

3.00

3.50

10.00 15.00 20.00 25.00 30.00 35.00

Condensation temperature [°C]

CO

Pto

tal

COPtot TR1 jan08-aug08 COPtot TR1 sep08-jun09COPtot TR2 with HR COPtot CC1 COPtot TR2 without HR Linéaire (COPtot CC1 )Polynomial (COPtot TR2 with HR) Polynomial (COPtot TR1 jan08-aug08)Polynomial (COPtot TR1 sep08-jun09) Linéaire (COPtot TR2 without HR)

Figure 6.16: Total COP with a load ratio of 3 in function of the condensing temperature for the three systems analysed

The positive impact of the borehole could be eliminated using the analysis done in the Chapter

7.4. The following Figure 6.17 present the COP of TR2 system without the subcooling effect of

the borehole. Thus, it allows better comparison to the value of the other two systems.

1.00

1.50

2.00

2.50

3.00

3.50

10.00 15.00 20.00 25.00 30.00 35.00


CO

Pto

tal

COPtot TR1 jan08-aug08 COPtot TR1 sep08-jun09

COPtot TR2 no subcooling COPtot CC1

Figure 6.17 Total COP with a load ratio of 3 in function of the condensing temperature for the three systems analysed, TR1 system with elimination of the borehole subcooling effect.


David Freléchox

55

The elimination of the borehole effect was calculated as a percentage for each unit KAFA1,

KAFA2 and KA3. Then using the average of these three values, it was possible to create this

plot. These assumptions could explain the spread point on the TR2 curve on the figure above.

In comparison with the simulation and as can be seen on Figure 9.1 the trends of the curve

correlates well. The value does not match du to the characteristics of each supermarket which

are then unified through the data input of the models.

The negative impact of the coolant circuit on the systems TR1 and CC1 is clearly visible on the

Figure 6.19 below, where the condensing temperature of the three systems is plotted against

the ambient temperature.

0.00

5.00

10.00

15.00

20.00

25.00

30.00

35.00

40.00

-15.00 -10.00 -5.00 0.00 5.00 10.00 15.00 20.00

Outside temperature [°C]

Con

dens

atio

n te

mpe

ratu

re [°

C]

COPtot TR1 jan08-aug08 COPtot TR1 sep08-jun09 COPtot TR2 with HR

COPtot CC1 COPtot TR2 without HR

Figure 6.18: Condensation temperature versus outside temperature fort he three systems

As can be seen in the figure, the cascade system CC1 works independently of the outside

temperature, so the condensation temperature is mainly constant. The 2 curves for TR1 show

the settings changes on the coolant loop after August 2008. It also shows clearly the rise of the

condensing temperature at low outside temperature for the TR2 system. When the heat

recovery system is not used in TR2 system, the supermarket without coolant loop has the

lowest condensation temperature for a specific outside temperature.


David Freléchox

56

The following Figure 6.18 shows the relation between the total COP and the outside

temperature.

1.00

1.50

2.00

2.50

3.00

3.50

-15.00 -10.00 -5.00 0.00 5.00 10.00 15.00 20.00

Outside temperature

CO

Pto

tal

COPtot TR1 jan08-aug08 COPtot TR1 sep08-jun09COPtot TR2 with HR COPtot CC1 COPtot TR2 without HR Linéaire (COPtot CC1 )Polynomial (COPtot TR2 with HR) Polynomial (COPtot TR1 jan08-aug08)Polynomial (COPtot TR1 sep08-jun09) Linéaire (COPtot TR2 without HR)

Figure 6.19: Total COP with a load ratio of 3 in function of the outside temperature for the three systems analysed

At an average outside temperature above 10°C, the u se of heat recovery system on the TR2 is

no longer necessary and the supermarket has the best COP mainly due to the absence of the

coolant loop. TR1 curve after the change of the settings of its coolant/condensing temperature

(curve COPtot TR1 Sep08-Jun09) is interesting. COP sink with increasing outside temperature

and below 0°C, stabilization appears caused by the limitation of the minimum condensing

temperature at around 10°C. CC1 system has the lowe st COP mainly due to its high

condensing temperature.


David Freléchox

57

7 Specific system analysis

After studying the systems in general and comparing them, specific analysis will be performed.

After some modifications, especially in regulation, their positive or negative effects on the

cooling system can be observed.

7.1 Effects of the installation of a frequency converte r on the

compressor

After more than a year in operation, the installer of the TR1 supermarket decided to do some

improvements in the refrigeration system. The supermarket has two transcritical chiller units

KA1 and KA2, and two transcritical freezer units with two-stage compressor FA1 and FA2. In

March 2009, a frequency regulator was installed on one compressor of the rack KA2 and FA2.

Below, we discuss the effects of this change. The purpose of this improvement is to limit the

start and stop of compressors in order to limit the electrical energy consumption and adapt the

compressors’ capacity to the load. Indeed, each start of a compressor generates a consumption

peak. Figure 7.1 shows the consumed energy of the compressors.

0

5

10

15

20

25

30

26.02.200900:00

03.03.200900:00

08.03.200900:00

13.03.200900:00

18.03.200900:00

23.03.200900:00

28.03.200900:00

02.04.200900:00

Ele

ctric

al p

ower

[kW

]

Electrical power consumption of the compressor KA2

Figure 7.1: Electrical power consumption of the compressors in KA2 during March 2009


David Freléchox

58

When one compressor is running at nominal speed which corresponds to 50 Hz, it consumes

about 11 kW. When two are running, that means about 22 kW. Since the 12th of March 2009

and following the installation of a frequency converter, the consumed power evolves on a wider

band. The compressor speed is regulated according to the needs using the frequency levels.

The frequency varies from 30 to 60 Hz, this last value explain why the electrical power could

reach 13 kW when only one compressor is running. Thus, the second compressors briefly starts

and not as often as without frequency converter.

The frequency converter regulates the operation of the compressor as a function of the suction

pressure. As shown on Figure 7.2, this maintains it more constantly and provides a better

stability of the evaporating temperature. The average evaporation temperature increases from -

10°C to -7°C which is related to the regulation tec hnique which allows better control of the

evaporation temperature settings when a frequency converter is used. This highest evaporation

pressure should improve the COP.

0

5

10

15

20

25

30

35

40

26.02.200900:00

03.03.200900:00

08.03.200900:00

13.03.200900:00

18.03.200900:00

23.03.200900:00

28.03.200900:00

02.04.200900:00

Pre

ssur

e [b

ar]

-20.00

-15.00

-10.00

-5.00

0.00

5.00

10.00

Tem

pera

ture

[°C

]

Suction pressure Evaporation temperature Mob. Avg. T_evap on 20 per.

Figure 7.2: Suction pressure and evaporation temperature of KA2 during March 2009


David Freléchox

59

The suction pressure stability is an important factor for the proper operation of refrigeration

system. Electronic expansion valves are used in these supermarkets to control the internal

superheat, which is a temperature difference between the outlet evaporator temperature and

the temperature of evaporation. To get the outlet evaporator temperature we did an average

with all cabinets. If the evaporation temperature is constant, the expansion valve suffers fewer

disturbances and therefore uses the evaporation surface more efficiently. A decrease of the

superheat and therefore an increase of the efficient heat exchange surface enable to reduce the

temperature difference between refrigerant and the external medium. The evaporating

temperature could be increased, and offers good prospects for improving the COP.

The Figure 7.3 demonstrates the better control of the superheat following the implementation of

a frequency converter. The spread of the superheat value is smaller. Note that the superheat

settings for CO2 systems – about 8 to 12 K - still exceeds settings for traditional systems as

R404A which is 4 to 7 K. These observations were made on a R404A supermarket at the

IWMAC interface [IWM09].

-5.00

0.00

5.00

10.00

15.00

20.00

25.00

26.02.200900:00

03.03.200900:00

08.03.200900:00

13.03.200900:00

18.03.200900:00

23.03.200900:00

28.03.200900:00

02.04.200900:00

Tem

pera

ture

diff

eren

ce [K

]

Internal SH External SH Mob. Avg. Internal SH Mob. Avg. External SH

Figure 7.3: Internal and external Superheat of KA2 during March 2009


David Freléchox

60

The external superheat produced by heat transfer with the ambience along the suction pipe is

less influenced by this change and could be observed in the Figure 7.3. On one side the

refrigerant mass flow is often lower and thus its temperature should increase, but on the other

side this could be offset by a generally higher evaporating temperature and lower heat transfer

coefficient due to the lower speed. These two phenomenons in opposition could explain the

small changes in the external superheat behaviour before and after the implementation of

frequency converter.

The better stability of the evaporating pressure, the external and the internal superheat are also

visible on the freezer system FA2 after putting on the frequency converter. However, as can be

seen in the Figure 7.4 the evaporation pressure does not increase. The average is apparently

little lower, probably due to a reference settings for the frequency converter at -35°C.

0

2

4

6

8

10

12

14

16

18

20

26.02.200900:00

03.03.200900:00

08.03.200900:00

13.03.200900:00

18.03.200900:00

23.03.200900:00

28.03.200900:00

02.04.200900:00

Pre

ssur

e [b

ar]

-45.00

-40.00

-35.00

-30.00

-25.00

-20.00

-15.00

-10.00

-5.00

0.00

Tem

pera

ture

[°C

]

Suction pressure Evaporation temperature Mob. Avg. T_evap

Figure 7.4: Suction pressure and evaporation temperature of FA2 during March 2009


David Freléchox

61

Obviously the interest of this modification is to reduce the power consumption of the system.

The objective is reached on the medium temperature cooling system, mainly due to keeping an

average evaporation temperature 3 K higher than before. The Figure 7.5 shows the compressor

electrical power and the coolant temperature in daily average for the chillers system KA2. After

the 12th of March 2009 - the day when the frequency converters were putting in - the electrical

power consumed by the compressors significantly fell and even though the coolant temperature

remained constant. It is an important improvement and the energy saving reach about 10 %.

0.00

2.00

4.00

6.00

8.00

10.00

12.00

14.00

16.00

28.02.2009 04.03.2009 08.03.2009 12.03.2009 16.03.2009 20.03.2009 24.03.2009 28.03.2009 01.04.2009

Ele

ctric

al p

ower

[kW

] - T

empe

ratu

re [°

C]

Compressor electrical power Coolant temperature Mobil avg. on 3 periods

Figure 7.5: Daily average of the compressor electrical power and coolant temperature for KA2 during March 2009

This modification does not affect the electrical power consumption for the freezer system FA2.

The evaporation temperature is more stable but its average is still on the same level as before

the change. There is therefore no influence on the power consumption. The Table 7-1 shows

the improvements for the chiller KA2 and the stability for the freezer FA2.

KA2 FA2 KA2 - FA2 Eel_comp_avg Eel_comp_avg Coolant_T_in

[kW] Var. [%] [kW] Var. [%] [°C]

Jan_09 12.56 9.14 7.1 Feb_09 12.49 -0.6% 9.01 -1.4% 7.0 Mar_09 11.63 -6.9% 9.10 1.0% 7.0 Avr_09 11.31 -2.8% 9.15 0.5% 7.4

Table 7-1: Monthly average of the electrical power consumption for KA2 and FA2


David Freléchox

62

On the Figure 7.6, it is easy to see the energy savings associated to the frequency converter

installation on the medium temperature unit KA2. From February to April 2009, the electricity

consumption has dropped by about 10% while at the same time on the parallel system KA1,

without frequency converter, the consumption rose by about 4%. Note that for this period the

heat sink temperature was quite stable at about 7 °C.

9.00

9.50

10.00

10.50

11.00

11.50

12.00

12.50

13.00

13.50

14.00

Dec_08 Jan_09 Feb_09 Mar_09 Apr_09

Ele

ctric

al p

ower

[kW

]

5.00

6.00

7.00

8.00

9.00

10.00

11.00

12.00

Tem

pera

ture

[°C

]

KA1 Avg. electrical consumption KA2 Avg. electrical consumption Heat sink temperature

Hz Converter installationon the 12th of March 2009

Figure 7.6: Comparison the two KA units of TR1 after the installation of a frequency converter on KA2.

7.2 Discharge pressure valve regulation

Transcritical systems are usually equipped with a regulation on the high pressure side. This

valve is designed to regulate the discharge pressure in order to reach the optimal temperature /

pressure relation in transcritical operation. In subcritical regime, it can also be used to increase

the fluid temperature in the condenser / gas cooler. This could be necessary when the system is

equipped with a heat recovery system.

A defined function transmits a signal to the actuator and handles the gradual opening of the

valve to get the better operating point. This function normally binds the discharge pressure to

the gas cooler outlet temperature. Figure 7.7 shows cycles with different discharge pressure at

the same gas cooler outlet temperature. It also shows a Danfoss discharge pressure valve.


David Freléchox

63

Figure 7.7: CO2 transcritical cycle with gas cooler exit temperature of 40°C at different discharge pressure (left), (Sawalha, ( 2008) – Danfoss ICM/ICAD Valve for condensation regulation (right) (Danfoss Refrigeration, 2006).

ICM are direct operated motorized valves driven by actuator type ICAD (Industrial Control

Actuator with Display). ICM valves are designed to regulate an expansion process in liquid lines

with or without phase change or control pressure or temperature in dry and wet suction lines

and hot gas lines. The ICM motorized valve and ICAD actuator assembly offers a very compact

unit with small dimensions. The ICAD is controlled via a modulating analogue signal (e.g. 4-20

mA/2-10 V) or a digital ON/OFF [DAN06].

The IWMAC-view, Figure 7.8, shows the compressors and gas cooler for a transcritical system

and shows the position of the ICAD valve. It is placed directly after the gas cooler. A liquid

accumulator is installed after the valve to balance the mass flow demand from the expansion

valves.

Figure 7.8: IWMAC interface for compressor and gas cooler for KA2, 27 April 2009, 14h10.


David Freléchox

64

The control function of this valve has been changed on all systems in the TR1 Supermarket

during March 2009. The change was made at the same time as the frequency converters on

KA2 and FA2 (Chapter 7.1). The influence of these changes was studied during March 2009 on

KA1 and FA1. These units were not equipped with frequency regulator, so the changes and

disturbances can only be attributed to the modification of the settings of the regulation valve.

The Figure 7.9 shows the operation changes of the regulation valve after the modification of the

function.

0

20

40

60

80

100

120

26.02.200900:00

03.03.200900:00

08.03.200900:00

13.03.200900:00

18.03.200900:00

23.03.200900:00

28.03.200900:00

02.04.200900:00

Val

ve o

peni

ng [%

]

Condensation valve KA1 Mobile average on 20 periods

Figure 7.9: Opening of the discharge pressure regulation valve for KA1 during March 2009

As can be observed in the figure, the average opening percentage moved from about 60% to

about 30%. This modification "drowns" the gas cooler and allows a better exchange. AS can be

seen on the Figure 7.10, the fluid is totally condensed and the subcooling is improved.


David Freléchox

65

-2.00

-1.00

0.00

1.00

2.00

3.00

4.00

5.00

6.00

7.00

8.00

26.02.200900:00

03.03.200900:00

08.03.200900:00

13.03.200900:00

18.03.200900:00

23.03.200900:00

28.03.200900:00

02.04.200900:00

Tem

pera

ture

diff

eren

ce [K

]Subcooling KA1

Figure 7.10: Subcooling of KA1 during March 2009

The closure of this valve causes a pressure drop. The Figure 7.11 shows the pressure drop

after the regulation valve of about 6 bars, whereas before the change, the difference was less

than 2 bars. This pressure drop causes partial evaporation of liquid and causes the formation of

a liquid - gas mix in the liquid tank. Note that the partial closure of the valve causes a big

pressure drop but a minimal increase in the high pressure of about 1 bar. Thus, the compressor

should consume slightly more energy due to this additional ∆P.

30

35

40

45

50

55

60

26.02.200900:00

03.03.200900:00

08.03.200900:00

13.03.200900:00

18.03.200900:00

23.03.200900:00

28.03.200900:00

02.04.200900:00

Pre

ssur

e [b

ar]

HP HP before exp. valve

Figure 7.11: High pressure for KA1 before the ICAD valve (HP) and after the ICAD valve (HP before exp. Valve) during March 2009.


David Freléchox

66

The Figure 7.12 shows an EES graphical visualization of the different thermodynamic changes

associated with the use of this new function. The increase of the pressure drop caused by the

regulation valve is evident, increasing the subcooling also. However, this large pressure drop

probably causes a partially evaporation of the fluid, so the output state of the liquid tank is still a

question mark. If none IHE is installed before the expansion valves (as it is for this transcritical

system), there is a risk of flash-gas amount the liquid line. This could create bad operating of

the expansion valve.

-320 -310 -300 -290 -280 -270 -260 -25020

60

h [kJ/kg]

P [b

ar]

15.5°C

1.23°C

-11.8°C

0.2 0.6 0.8

CarbonDioxide

Liquid saturation line Old f unction

New f unction

Cond.out

Exp.v alv e.in

Figure 7.12: EES visualisation on the h-logP diagram for the two condensation regulation functions, KA1 during March 2009

The Table 7-2 shows the increase of a few percent of the cooling capacity generated through

the improved subcooling. As assumed, the more important closure of the regulation valve

creates a slight increase of the electric power consumption due to the higher pressure drop.

Thus, we could not observe an improvement of the coefficient of performance. The reason for

this modification may also be a need for additional power on the heat recovery. But again, the

small increase in condensation pressure can not allow recovering additional heat.

March 2009 Qel [kW] [%] Q cool [kW] [%] Q cond [kW] [%] Sub. [K] COP [%] Before 10.33 48.18 47.38 2.05 4.68 KA1 After 10.54 2.0% 49.20 2.1% 48.34 2.0% 4.05 4.68 0.0%

Before 11.12 17.75 21.84 0.87 1.60

FA1 After 11.24 1.1% 18.37 3.5% 22.52 3.1% 3.53 1.63 1.9%

Table 7-2: List of performance data when one compressor is running for KA1 and FA1 for March 2009 before and after changing the function of the regulation valve.


David Freléchox

67

7.3 Influence of the internal and external superheat on the DX CO2

refrigeration systems

This part is based on data and reviews of the KTH book “Refrigerating Engineering” [GRA05],

section 3.34 to 3.43.

Most supermarkets use direct expansion systems with electronic or thermostatic expansion

valves. They cause overheating in an order of 7 to 12 K to ensure 100% vapour state at the

outlet of the evaporator. CO2 usually requires the use of a higher superheat of about 5 K than

refrigeration systems using traditional fluids. This is mainly due to the use of evaporator with

standard design and not designed according to CO2’s properties.

Internal superheat can also be created using an internal heat exchanger (IHE). It subcools the

liquid and overheats the vapour. Increasing the internal superheat, it acts on the whole system.

It ensures 100% liquid state at the entrance of the expansion valve and 100% steam state at the

compressor inlet. The Figure 7.13 shows the effects of the internal and external superheat on

the COP at different condensation temperature.

-1

-0.8

-0.6

-0.4

-0.2

0

0.2

0 5 10 15 20 25 30 35 40 45


Effe

ct o

n th

e C

OP

[%/°C

]

To -30°C _ SH int To -30°C _ SH ext To -10°C _ SH int To -10°C _ SH ext

Figure 7.13: Effect of the internal and external superheat on the COP by using CO2 in standard refrigeration system.


David Freléchox

68

Depending on the condensation temperature the effect of the internal superheat on the COP

can be either negative or positive. In relatively cold climates where the use of floating

condensing is prominent, the internal superheat has a negative effect, while in warmer climates;

it could have a positive effect. So there are opportunities to improve the COP by using an IHE

under special conditions.

The positive or negative effects are usually small and the use of an IHE can seldom be justified

on this basis. However, the use of an IHE is a way to avoid the heavily detrimental influence of

external vapour superheat and to provide subcooling to avoid flash gas in the long supply line

before the expansion valve. This type of heat exchanger is often used with long suction line.

Owing to the raised temperature of the suction vapour, the heat transfer to the piping from the

surroundings will be diminished or fully eliminated.

External superheat is caused by energy exchange with the atmosphere along the suction pipes.

The influence of external superheat is more important at high evaporation temperatures

because of the lower pressure ratio and hence the lower compressor energy consumption.

However, in absolute terms, the external superheat of the low temperature system is higher

than the one of the medium temperature systems. Thus, the external superheat has an

important negative effect on the COP of both low and medium temperature systems.

In refrigeration system, external vapour superheat is a main negative burden. It is detrimental

not only to the coefficient of performance but also to the volumetric refrigeration effect, leading

to a demand of a compressor with larger swept volume flow.

7.4 Subcooling with ground heat sink

The TR2 supermarket is equipped with a heat exchanger to subcool the liquid using ground

heat sink. The potential of this free subcooling is obvious. The COP improvement mainly

depends on the temperature of the cold source. This source can be water, air, or the ground.

Reaching a big subcooling is most important and efficient in the summer when condensation

temperatures are relatively high. In winter, the use of floating condensation reduces the

potential and need of subcooling. The Figure 7.14 presents positive effects of subcooling for the

TR2 supermarket. Note that improvements of this magnitude are also visible on the medium

temperature and low temperature of the booster system.


David Freléchox

69

0.00%

5.00%

10.00%

15.00%

20.00%

25.00%

30.00%

35.00%

Sep_08 Oct_08 Nov_08 Dec_08 Jan_09 Feb_09 Mar_09 Apr_09 May_09 Jun_09

CO

P im

prov

emen

ts [%

]

0.00

5.00

10.00

15.00

20.00

25.00

30.00

Tem

pera

ture

[°C

] - d

T s

ubco

olin

g [K

]

COP improvement due to subcooling dT subcooling ground heat sinkCondensation temperature

T evap = -10°C

SH int = 10 KSubcool. Ref = 0 K

Figure 7.14: COP improvement due to the subcooling with the heat sink for KA3 medium temperature unit in TR2 supermarket .

In contrast to what we have just cited, the improvement is important in winter as well in this

case. This is due to the high pressure increase in order to recover more energy to heat the

store. If the gas coolers are normally working, so they do not maintain the condensation

pressure too high, as it is the case in May and June 2009, then a COP improvement of about

15% by using the ground heat sink to subcool the liquid is possible. As the heat sink

temperature is almost constant the potential of the subcooling is higher when the condensing

temperature is raised.

As can be seen in Figure 7.14 the impact of subcooling on COP varies with the temperature of

condensation. The explanation lies in the isotherms shape near the critical point. Figure 7.15

shows the big impact of subcooling when the condensation temperature is between 20 and

30°C. Just above the critical point the influence o f subcooling would be even greater for the

same raison.

The influence of subcooling is presented in more detail through a simulation in Chapter 10.4.


David Freléchox

70

-300 -275 -250 -225 -200 -175 -150 -1253x101

102

102

h [kJ/kg]

P [b

ar]

30°C

25°C

20°C

15°C

10°C

CarbonDioxide

Figure 7.15: Isotherme shape in h-logP diagramm for CO2 near critical point

The impact of subcooling on systems using CO2 as fluid may also be a decisive advantage

compared to traditional fluids such as R404A. On Figure 7.16, it is easy to see that 1 K

subcooling give about 1% of COP improvement in the case of R404A. But it is not valid for CO2.

In this case, the influence of subcooling is more important and varies greatly depending on the

condensation temperature.

0.0%

10.0%

20.0%

30.0%

40.0%

50.0%

60.0%

0.00 2.00 4.00 6.00 8.00 10.00 12.00 14.00 16.00 18.00 20.00

Subcooling [K]

CO

P im

prov

emen

ts [%

]

CO2 - Tcond=20°C CO2 - Tcond=25°C CO2 - Tcond=30°C

R404A - Tcond=20°C R404A - Tcond=25°C R404A - Tcond=30°C

Tevap=-10°C

SH int = 10KSubcool. Ref. = 0 K

Figure 7.16: Effect on the COP of the subcooling at different condensation temperature for CO2 and R404A with an evaporation temperature at -10°C and an internal superheat of 10 K.


David Freléchox

71

7.5 Analyse of TR2 system under transcritical regime

As presented in Chapter 5, two of our systems can operate under transcritical regime if the

outside temperature rises. TR1 is located in the north and therefore does not really suffer the

influence of outdoor climate. While, TR2 operated under transcritical regime in late June 2009.

The Figure 7.17 shows the trend of the COP when the system operates in transcritical regime

0

10

20

30

40

50

60

70

80

90

20.06.200900:00

21.06.200900:00

22.06.200900:00

23.06.200900:00

24.06.200900:00

25.06.200900:00

26.06.200900:00

27.06.200900:00

28.06.200900:00

Tem

pera

ture

[°C

] - P

ress

ure

[bar

]

0

1

2

3

4

5

6

7

8

9

10

CO

P [-

] - T

empe

ratu

re d

iffer

ence

[K]

High pressure Outside temperatureMob. avg. dT gc out-outside temp on 200 per Mob. avg. COP_evap on 200 per.

Figure 7.17: KA3 unit in TR2 system during one week at the end of June 2009

Logically, the COP follows a downward trend when the pressure increases. However, there is

no big step in COP between subcritical and transcritical regime. Another important effect is the

approach temperature difference between the gas cooler outlet temperature and the outside

temperature, it falls during transcritical operation.

Figure 7.18 below provides a more detailed display. The COP and the approach temperature

difference are both falling with high pressure increase. The approach temperature difference

even falls below 0 K. In theory, this is not possible unless spraying water on the condenser is

used. It creates an adiabatic cooling process on the air at the inlet of the gas cooler.


David Freléchox

72

0

10

20

30

40

50

60

70

80

90

24.06.2009 00:00 24.06.2009 12:00 25.06.2009 00:00 25.06.2009 12:00 26.06.2009 00:00

Tem

pera

ture

[°C

] - P

ress

ure

[bar

]

-2

0

2

4

6

8

10

CO

P [-

] - T

empe

ratu

re d

iffer

ence

[K]

High pressure Outside temperaturedT gc out-outside temp COP_evapMob. avg. dT gc out-outside temp on 200 per Mob. avg. COP_evap on 200 per.

Figure 7.18: KA3 unit from TR2 system during two days at the end of June 2009

This fall of the approach temperature difference is even more visible on the booster system

KAFA1 on Figure 7.19. The temperature difference clearly drops below 0 K and therefore

indicates that the water spray is activated. This helps to eliminate as far as possible transcritical

operations and limit the COP fall during the summer periods.

0

10

20

30

40

50

60

70

80

90

22.06.200900:00

22.06.200912:00

23.06.200900:00

23.06.200912:00

24.06.200900:00

24.06.200912:00

25.06.200900:00

25.06.200912:00

26.06.200900:00

Tem

pera

ture

[°C

] - P

ress

ure

[bar

]

-4

-2

0

2

4

6

8

10

12

14

16

18T

empe

ratu

re d

iffer

ence

[K]

Outside temperature High pressure dT gc out - outside temperature

Figure 7.19: KAFA1 unit in TR2 system during four days at the end of June 2009


David Freléchox

73

8 Simulation model

Modelling is an important part of this study. It helps to unify the parameters for each

supermarket and offers to make comparisons fair and independent. The base of each

simulation is a model created using the software EES. Each model uses input data own or

unified which allow calculating all thermodynamic points of the system. The main functions used

are listed and explained below.

The models are defined by the functions and assumptions below. Thermodynamic equations

are used to simulate the thermodynamic state of the fluid after a heat exchanger or a

compressor, for example.

Models are written using EES software. Its basic function is to provide the numerical solution to

a set of algebraic equations. It has many built-in mathematical and thermo-physical property

functions for refrigerants. EES uses an equation of state approach rather than internal tabular

data to calculate the properties of fluids. Details about EES and the method of properties

calculation can be found in [KLE06].

8.1 Data input and assomptions

These data are used to compare each system through simulation. We tried as much as possible

to have identical operating conditions and to represent the reality for each model. Below is a list

of input parameters and assumptions for the simulations.


David Freléchox

74

TR1 TR2 CC1 Cooling capacity freezer [kW] 50 Cooling capacity chiller [kW] 150 Load ratio [-] 3 Evaporation temperature freezer [°C] -35 -35 -35 Evaporation temperature chiller [°C] -10 -10 -12 Internal superheat CO2 [K] 10 10 10 Internal superheat R404A [K] NA NA 7 External superheat freezer[K] 15 15 15 External superheat chiller [K] 10 10 NA1 Subcooling [K] 0.5 0.5 0.5 Oil cooler losses related to total compressor power [%] 15 15 NA Heat losses from compressor related to total compressor power [%] 7 7 7 Pump brine power related to the cooling capacity [%] NA NA 4 Heat gains in the brine loop [kW] NA NA 10 ∆T coolant in / out [K] 5 NA 5 Brine temperature in [°C] NA NA -8 Brine temperature out [°C] NA NA -4 LMTD cascade condenser [K] NA NA 6 Condensers and condensers/gas coolers approach temperature [K] 5 5 5 IHE effectivness freezer [%] NA NA 20 IHE effectivness chiller [%] NA NA 50

Table 8-1: Data input and assumptions for the simulations

1 Instead of external superheat, we used brine losses


David Freléchox

75

A load ratio of 3 was chosen as a representative value for Swedish supermarkets [SAW08]. The

refrigeration capacity is based on a mid-size supermarket with 50 kW for the low temperature

unit and 150 kW for the medium temperature unit [SAW08]. Evaporating temperature values are

recommended by cabinet manufacturers and used in most studies. Nevertheless, it appears

according to Girotto et al. (2005) that the use of higher temperatures evaporation of

approximately 2 K is possible with CO2. However, in order to produce a neutral study, we

decided to keep the standard values at -10°C and -3 5°C. They also correlate quite well with our

measurements in-situ.

Concerning the internal superheat, we remained in magnitudes and choose fixed values closer

to our measurements. Especially since these values depend on the configuration of each

system. The superheat of 7 K and the absence of external superheat for the R404A unit result

from the fact that this unit is a compact system.

The heat losses and the oil cooler losses are based on manufacturer’s data, as well as some

measures. They are a percentage of the compressors power consumption. Note that R404A

and CO2 compressors for the CC1 supermarket are not equipped with external oil cooler. The

power of the brine pumps is a percentage of the cooling capacity of the R404A unit. The values

are based on a study from E. Granryd (2007) and correlate with our measurements.

The effectiveness of heat exchangers and the logarithmic mean temperature difference are

based on our measurements and seem to be reasonable values. The temperature difference

between ambient air and the condenser/gas cooler outlet temperature is defined as the

approach temperature (in the case of a direct heat exchange air/CO2). Samer Sawalha

[SAW08] used 5 K, Girotto et al. [GIR04] 5 K as well, but they suggest the possibility of

improving the gas cooler and reach an approach temperature up to 2 K. A recent study by Ge et

al. [GEY09] uses 3 K. Our measurements on existing system give values between 4 and 5 K.

Finally, the value of 5 K was chosen as the most independent and realistic.

Ambient temperatures were logged for all the systems under investigation, A problem occurred

on the temperature sensors in CC1 supermarket and in order to have outside temperature for a

given period for the three systems, we use temperature data from Sveriges meteorologiska och

hydrologiska institute [SVE09] for the simulation. A weather station near each supermarket was

chosen as reference.


David Freléchox

76

8.2 Function to simulate the dependence of the cooling capacity to the

outdoor temperature.

The cooling capacity needed for the cabinets mainly depends on the ambience of the store. The

store temperature and humidity are influenced by external climatic conditions. Based on data in

the doctoral thesis of Jaime Arias [ARI05] and on our measurements for the TR1 supermarket

during summer 2008, we developed the Equation 8-1 linking the cooling capacity to the outside

temperature.

%100_)3.002.0( ooutsideo QTQ && ⋅+⋅= 8-1

As can be seen in the Figure 8.1, a limit was set at 10°C below which the outdoor temperature

has no more influence. Under these conditions, the supermarket is heated and maintained at

constant conditions. The humidity is low during the winter period and thus has little influence on

our system. This is obviously an estimation based on information collected through several

sources.

0

20

40

60

80

100

120

-20 -10 0 10 20 30 40

Outdoor temperature [°C]

Coo

ling

capa

city

[%]

Qo=(0.02*Toutside+0.3)*Qo_100%

Figure 8.1: Function binding the percentage used of the maximal cooling capacity to the outdoor temperature


David Freléchox

77

8.3 Function to simulate the fluid compression

In order to calculate the compressor outlet conditions of the fluid, the Equation 8-2 based on the

total efficiency of the compressor was used.

incompoutcomp

incompisoutcomptot hh

hh

__

___

−−

=η 8-2

Where the total efficiency was defined by a function developed for each compressor type and

based on the compressor data [BIT09] and [DOR09]. The functions are presented on the

following Figure 8.2.

Eta tot = -0.1086x2 + 1.0494x + 59.877

Eta tot= -2.7854x2 + 19.477x + 33.013

Eta tot = -0.2105x2 + 1.9784x + 50.212

0

10

20

30

40

50

60

70

80

90

100

0 2 4 6 8 10 12

Pressure ratio [-]

Tot

al e

ffici

ency

[%]

TCDH372B-D

TCS373-D

SCS 362 SC

Figure 8.2: Total efficiency of 3 CO2 compressors types in function of the pressure ratio

As can be seen on the figure, the functions are based on spreader points at low pressure ratio.

Generally the medium temperature compressor TCS373-D has a higher efficiency than the low

temperature compressor SCS362-SC. The TCDH372B-D compressor is a two stages

compressor which explains its large pressure ratio range.


David Freléchox

78

8.4 Limit of the condensing temperature

The TR1 supermarket uses the floating condensation to limit the compressors power

consumption. The idea is to use the condenser capacity at the maximum to maintain the

condensing pressure as low as possible. The pressure ratio is lowered and compressors

consume less energy.

However, this method leads to technical constraints. Lowering the condensing temperature will

generally increase the cooling capacity and reduce the electrical energy consumed. Thus, it is

preferable to install a frequency converter on the compressor to continuously adapt its power,

and adjust the produced cooling capacity on the demand. If the pressure drop across the

expansion valve becomes too low, then its capacity goes down and there is a risk of under-

supply of refrigerant in the cabinets.

These constraints mean that there should be a minimum condensation temperature for each

system. The CO2 appears to have an advantage at this level regarding its high pressure

difference between low and high pressure (about 40 bars at -10°C +25°C). It should be possible

to lower the condensing temperature more than in the case of using traditional HFC, which are

working at lower pressure differences. For example, the R404A has a pressure difference of 8

bars between evaporation at -10°C and condensation at +25°C.

To fix a limit in our simulation model, we get contact with companies having experience in the

floating condensation field, for both CO2 and R404A. Regarding CO2, Micael Antonsson from

Green and Cool AB [ANT09] mentioned that it would be possible to work down to -5°C

condensation with evaporation at -10°C. The pressur e difference is still about 5 bars and this is

sufficient for proper functioning of the expansion valve. Nevertheless, there is a lack of

guarantee from compressors’ manufacturer at this low pressure difference. Long-term tests are

underway. Our measurements of both TR1 and TR2 supermarket indicate that condensation

temperatures are limited to 10 - 12°C.


David Freléchox

79

Regarding R404A, Lasse Viktorsson from Partor AB indicates [VIK09] that it is possible to

achieve a condensation temperature of 5°C for mediu m temperature system and -5°C for low

temperature systems. An observation on the system in use from Partor AB indicates limit

condensation temperatures between 5 and 10 ° C.

The two contacts raised a warning that the lowest condensing temperature is not necessarily

the ideal energy wise. Indeed, there is an optimum for each system depending of its design and

its location between the energy saved and the additional electrical energy consumption by the

condenser fans. According to the measures and the diverse experiences of these companies,

floating condensing temperature limit lies at approximately 10°C for most installations in

Sweden.

With all this information and our measures, we finally decided to limit the temperature of

condensation of our model at 10°C.

8.5 Day and night influence on the cooling capacity

The differences between day and night are mainly du to the activity in the store, which suppose

an increase in humidity and temperature, the cabinet lights, the night curtains on medium

temperature cabinets and also the fill in of new products. All these parameters lead to an

increase in the cooling capacity demanded on the cabinets during the day.

This capacity variation can be more or less important depending on the store, the quality of its

ventilation and air conditioning system, number of its customers, its assortment of cabinets and

its management.

However, the difference between these two regimes is generally much larger on the medium

temperature system. The freezers are usually covered with lids or glass doors; it eliminates the

influence of the ambience on the low temperature system for a large degree.

The Figure 8.3 shows the day and night trends of the cooling capacity for medium temperature

unit in the TR1 supermarket during 10 days at the beginning of February 2009.


David Freléchox

80

0.00

10.00

20.00

30.00

40.00

50.00

60.00

70.00

80.00

90.00

100.00

01.02.200900:00

03.02.200900:00

05.02.200900:00

07.02.200900:00

09.02.200900:00

11.02.200900:00

Coo

ling

capa

city

[kW

]

-10.00

-8.00

-6.00

-4.00

-2.00

0.00

2.00

4.00

6.00

8.00

10.00

Tem

pera

ture

[°C

]

Cooling capacity TR2 - KA3 Outdoor temperature Mob. avg. 20 per (Qo)

T evap = ~ -10°C

Figure 8.3: Day and night trend of the cooling capacity of the KA3 unit in the supermarket TR2 during February 2009.

The figure above shows a decrease in cooling capacity up to 50% on the medium temperature

system overnight. Low temperature systems rather suffer a decrease of about 20%. Note that

this is always subject to the influence of several parameters, including climate. One of our

supermarkets, TR1, is situated in the northern part of Sweden and shows lower variations,

respectively in an order of -30% and -15%.

From the point of view of the consumption, because that is what we are interested in for the

simulation, the results of Figure 8.1 include the day and night variations. As shows on the

Figure 8.4, the fluctuations may be imagined as an over or under-consumption in relation to our

basic function. During the opening hours of the store, the cooling capacity is increased and

conversely decreased during the night.

On balance the consumption’s increase or decrease during day and night depends only on the

opening hours. If the store is open 12 hours and closed 12 hours (8-20H) as it is often the case,

the effect of the two regimes is cancelled. As we will compare several supermarkets

independently of their opening hours and that the introduction of a day-night function needs the

integration of a new time variable, we will not consider this parameter. Its influence is quite low

and already partly taken into account in our function linking the cooling power to the outside

temperature.


David Freléchox

81

0

20

40

60

80

100

120

140

-20 -10 0 10 20 30 40

Outdoor temperature [°C]

Coo

ling

capa

city

[%]

Qo=(0.02*Toutside+0.3)*Qo_100%

Figure 8.4: Function binding the percentage used of the maximal cooling capacity to the outdoor temperature with variation range + / - 25% for a typical medium temperature cabinet.

8.6 Validation of the models

Our model must be validated by comparing these results obtained using the EES simulation

with those obtained using the templates created in Excel which have enabled us to create the

COP curves shown previously.

Each model has different input parameters. For this comparison and in order to use the EES

model, we had to unify the models. While only the outside temperature is used as data input for

the simulation, for this validation the number of data input was increased to define the operating

points on the basis of our measurements.

The Excel templates calculate the COP of the system for each measurement point. Then the

monthly average COP is defined by the average of the instantaneous COP’s. EES simulation

uses the monthly average of each point of the circuit to calculate the monthly COP. Thus, there

is a slight difference in the two calculations, one is done for each measuring point, and which is

then used to obtain the monthly average, the other uses the monthly average of the

measurement points to calculate one COP.


David Freléchox

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The Figure 8.5 shows the trends for the 2 methods of calculation, calculation with Excel and

simulation using the EES models.

1

1.5

2

2.5

3

3.5

4

Jan_

08

Feb_0

8

Mar

_08

Apr_0

8

May

_08

Jun_

08

Jul_0

8

Aug_0

8

Sep_0

8

Oct_08

Nov_0

8

Dec_0

8

Jan_

09

Feb_0

9

Mar

_09

Apr_0

9

May

_09

Jun_

09

CO

P to

tal

COPtot TR1 COPtot TR2 COPtot CC1 COPtot EES TR1 COPtot EES TR2 COPtot EES CC1

Figure 8.5: Comparison of the COP between the template calculation and the EES simulation.

The figure shows interesting results for the three systems. The variation between the templates

and the model is quite low for the supermarket TR1. The gap between the templates and

simulation results is slightly higher for the last three months measured; this is mainly due to the

change in the function of regulation valve (see Chapter 7.2). This modification disturbs the

definition of the entry point into the expansion valve and thus causes small differences.

TR2 supermarket gives good matching results throughout the whole period analyzed. However,

the results between model and template are more deviated in the case of supermarket CC1

using the cascade R404A/CO2. The difference between the two calculation modes is constant

and approximately 5%. Several reasons contributed to this difference, first the outlet

temperature of compression is measured after each compressor and not on the common tube.

It is difficult to set the exact temperature to be used as data input in the simulation. Secondly

there is no energy measurement at the supermarket CC1, the compressors’ consumption is

based on the pressure ratio. This energy is an output in the templates but used as an input in

the simulation. This can of course also influence the results.

Nevertheless and in general the matching in the results between the Excel template calculation

and the EES simulation can be described as fairly good.


David Freléchox

83

9 Systems simulation

9.1 Variation of the condensing temperature

The use of floating condensation is an important improvement factor for the COP. Current

technologies such as electronic expansion valves allows also an easier control of this system.

Moreover, the TR1 supermarket to which data was collected use floating condensation.

A simulation of each system gives the COP according to the condensation temperature; the

results can be seen in Figure 9.1. It is evident that the coefficient of performance of refrigeration

system is strongly dependant on the condensing and evaporating temperature levels, in this

case the evaporation temperature is kept constant for all systems. For the systems using CO2

TR1 and TR2, in the subcritical operating range, a small difference in condensation temperature

can lead to a relatively bigger change in the COP. The use of a cascade system R404A/CO2

seems profitable for condensation temperatures higher than 26°C. In practice, this transition

depends on the design of each supermarket, its location, its climate and conditions of use.

However, the R404A actually requires the use - in Sweden – of a coolant circuit for heat

rejection on the condenser; this is due to the limitations on the amount of HFC’s used.

0

0.5

1

1.5

2

2.5

3

3.5

4

10 15 20 25 30 35 40 45 50


CO

P_t

otal

COPtot_TR1 COPtot_TR2 COPtot_CC1

Figure 9.1: Total COP for different condensation temperature

In comparison with the field measurement and as can be seen on Figure 6.17 the trends of the

curve correlates well. The value does not match du to the characteristics of each supermarket

which are then unified through the data input of the models.


David Freléchox

84

The higher COP of TR1 compared to TR2, can be observed in Figure 9.1, mainly results from

the use of free desuperheating between the two compression stages on the low temperature

system in TR1. This free desuperheat is related to the coolant temperature, so to the outside

temperature. This is useful at low outdoor temperatures, once a condensing temperature of

35°C is exceeded the benefit of the free desuperhea t is diminishing.

The Figure 9.2 also represents the total COP of the three simulated supermarkets, but this time

as a function of the outside temperature. Here, we incorporate in our analysis the effect of the

coolant circuits on CC1 and TR1, while TR2 uses a condenser/gas cooler on the roof. The

coolant loop was incorporated in the model of TR1 and CC1 in order to have similar systems

than the field measurements. Thus, TR1 and CC1 have 10 K temperature difference between

condensing and outside temperature because of the additional heat exchanger in the coolant

loop with an assumed approach temperature difference from 5K. The TR2 system includes only

the gas cooler with 5 K approach temperature difference. The condensation temperature limit of

10°C creates a COP limitation at low outside temper ature. The COP difference between TR1

and TR2 at low outside temperature is mainly due to the use of free desuperheating between

the two compression stages on TR1 low temperature unit.

0

0.5

1

1.5

2

2.5

3

3.5

4

-10 -5 0 5 10 15 20 25 30 35 40


CO

P_t

otal


Figure 9.2: Total COP for different outside temperatures

The specificity of TR2 supermarket allows a good COP. From the systems we have simulated,

the supermarket using CO2 with transcritical regime and direct condensation has the most

favourable coefficient of performance up to an outside temperature of about 23°C.


David Freléchox

85

9.2 Simulation using improvement possibilities for CO2 systems

The return interest in the use of CO2 is partly due to its low impact on the greenhouse effect,

but also to its favourable thermodynamic properties presented in Chapter 3.2 (CO2 as

refrigerant). Throughout the work in this thesis, a special care was given to making the

comparisons between CO2 and HFCs fair.

The simulation tool used allows evaluating the potential of CO2 by using its positive properties.

Thus, we will use for this new simulation an evaporation temperature of -7°C for medium

temperature applications and – 33°C for low tempera ture applications. These values are based

on our measurements in the supermarket TR1 and also on various publications, including

particularly Girotto et al. in 2004.

A 2K lower approach temperature difference applied to the condenser / gas cooler seems,

based on our measurements and the article cited previously is realistic. Obviously, in the case

of a real application that implies design improvements of condenser / gas coolers and

evaporators.

Figure 9.3 shows the evolution of the COP as a function of the external temperature before and

after improvement of the parameters on CO2 systems. In the case of the cascade system the

only change is the modification of the evaporation temperature on the low temperature system

from -35 to -33°C, which results in a small improve ment.

0

0.5

1

1.5

2

2.5

3

3.5

4

4.5

5

-10 -5 0 5 10 15 20 25 30 35 40


CO

P_t

otal


COPtot_TR1_improv COPtot_TR2_improv COPtot_CC1_improv

Tevap FA : -35 >> -33 °CTevap KA : -10 >> -7 °C∆tapproach: 5 >> 3 K

Figure 9.3: Total COP for different outside temperatures using improvements possibilities for CO2 systems


David Freléchox

86

In contrast, the two CO2 transcritical systems benefit greatly from these changes. Specifically,

the 2 K lower approach temperature, which was not applied on the R404A system, has a

significant positive effect. This also helps to raise the maximum COP when the limitation of the

condensation temperature at 10°C is reached.

The improvement for the cascade system does not exceed 2%, while supermarkets using CO2

transcritical systems can have a COP improvement up to 20%, especially at low temperature

just before the condensation limit at 10°C. When th e temperature increases, the improvement

possibility decreases, but even at 35°C the benefic ial effect is greater than 10%.

Improving the design of condenser/gas coolers and evaporators based on the advantageous

properties of CO2 is still a major challenge.

9.3 Annual simulation – comparison of the three systems in different

climates in Sweden

The three supermarkets analysed are located in Sweden. However, depending on their location,

the climate can considerably vary and therefore affect their energy consumption. To have a

better overview, the dynamic behaviour of each system was simulated for each location during

a year.

The simulation uses the outside temperature as data input and generates cooling capacity and

COP in hourly intervals. Thus, the power consumption of the refrigeration system can be easily

calculated. To use reliable and independent data, we used hourly outside temperatures of three

weather station located close to our supermarkets. The locations are Storön, Göteborg and

Floda linked to the three supermarkets TR1, TR2 and CC1 respectively.

The following Figure 9.4 presents the temperature range for three different weather stations in

Sweden. As can be seen, the Swedish temperatures are quite low and most of the time below

20°C. The analysis of the figure informs about the regime of the CO2 systems. It is clear that

they will operate in subcritical regime 98% of the time. Götenborg has the hottest climate,

followed by Floda. The temperature in Storön which is located in the north rarely exceeds 20°C.


David Freléchox

87

6520

35 9 0

4623

1957

8

5541

341133 5

626

1570

163384

1625

1075

1665

0

1000

2000

3000

4000

5000

6000

7000

<10 10-15 15-20 20-25 25-30 >30


Num

bers

of h

ours

[-]

Storon Goteborg Floda

Figure 9.4: Number of hours per year for different outside temperature levels in Storön, Göteborg and Floda – the three locations are in Sweden.

Using the temperature values in the above figure and on the basis of our simulations (see

Figure 9.2), systems using CO2 should be the most energy efficient. Figure 9.5 shows the

corresponding results. The system used by the supermarket TR2 consumes less energy in the

year. It is followed by the second transcritical TR2 system and the highest consumption is by

the R404A/CO2 cascade. The main advantage of TR2 compared to the other two systems is

the absence of a coolant loop. It is clear that the absence of the secondary circuit would be an

advantage for each system. However, different parameters influence this choice: first, if the

designers wish to use the heat rejected by the refrigeration system to heat the store (heat

recovery system), the use of a coolant system simplifies the system. Secondly and in relation to

the cascade system using R404A, Sweden has many rules legislating on synthetic refrigerants;

one of them strictly limits the amount of fluid in each installation. Thus, the use of a coolant

limits the amount of fluid, but unfortunately at the expense of effectiveness.


David Freléchox

88

342

373

419

317

384

424

370

296

447

0

50

100

150

200

250

300

350

400

450

500

Storon Goteborg Floda

Ann

ual e

nerg

y co

nsum

ptio

n [M

Wh]

TR1 coolant TR2 gas cooler CC1 coolant

Figure 9.5: Annual energy consumption for different supermarket systems in Storön, Göteborg and Floda – all three locations are in Sweden.

Table 9-1 shows the annual energy consumption excess in percentage for each location. As

expected the TR2 is the system with the lowest annual energy consumption regardless of the

location. Its advantage over the other two systems TR1 and CC1 is particularly important in a

relatively warm climate like in Göteborg. In a colder climate, as in Storön, limiting the

temperature of condensation at 10°C whatever the ou tside temperature decreases the

difference in energy consumption. Thus, in a very cold climate, the use of a coolant circuit in

order to recover heat is reasonably justified.

The use of a cascade system R404A/CO2 equipped with a coolant loop generates an

overconsumption of at least 20% compared to the best transcritical system using a

condenser/gas cooler. A transcritical system with a coolant loop as TR2 is also more efficient

than a R404A/CO2 cascade system.

Storon Goteborg Floda TR1 107% 115% 112% TR2 100% 100% 100% CC1 126% 121% 123%

Table 9-1: Annual energy consumption excess in percentage in comparison with the better system for each supermarket systems in Storön, Göteborg and Floda.


David Freléchox

89

9.4 Annual simulation – comparison of the three systems in different

climates in the World

Since performance differences between the systems under investigation vary over the ambient

temperature range, then differences in annual energy consumption will depend on the climate in

which the system will operate. Three different climatic examples were selected. The first is a

typical European climate represented by the climate of Frankfurt/Germany. A hot climate in the

USA is represented by Phoenix-Arizona/USA. Stockholm/Sweden is the third example, selected

to represent a cold climate.

Ambient temperatures and their variation over the year were generated using Meteonorm

[REM01] for every hour of the year. As can be seen in Figure 9.6 the temperature in Stockholm

and Frankfurt is under 10°C for more than 50% of th e time. In Phoenix the temperature levels

tend to be higher; the temperature is above 30°C fo r the greatest number of hours compared to

the other temperature ranges.

5479

1227

483

41 0178

7

1397 1367

2334

1530

4574

773

1436

1792

13981377

887

0

1000

2000

3000

4000

5000

6000

<10 10-15 15-20 20-25 25-30 >30


Num

bers

of h

ours

[-]

Stockholm Frankfurt Phoenix

Figure 9.6: Number of hours per year for different outside temperature levels in Stockholm / Sweden, Frankfurt / Germany and Phoenix – Arizona / USA.

A CO2 transcritical system in Stockholm or Frankfurt will operate subcritically for most of the

time while in Phoenix the system will operate in the transcritical region for about 40% of the

time. Using ambient temperatures from Meteonorm [REM01] for the selected cities as input

variables in the simulation model the annual energy consumption for each system in Figure 9.2

is calculated.


David Freléchox

90

Figure 9.7 represents annual energy consumption for the different systems under investigation

for the three selected cities. As can be seen in the figure, the cascade R404A/CO2 system CC1

has the highest energy consumption in cold climates, but the lowest in hot climates.

Despite the use of a coolant circuit, the CC1 system consumes less energy than the CO2

transcritical TR2 system using a condenser/gas cooler. The warm climate of Phoenix with over

40% of the time a temperature over 25°C explains th e advantage of the cascade.

900

419

816

384

1062

440 385345

456

0

200

400

600

800

1000

1200


Ann

ual e

nerg

y co

nsum

ptio

n [M

Wh]

TR1 coolant TR2 gas cooler CC1 coolant

Figure 9.7: Annual energy consumption for different supermarket systems in Stockholm / Sweden, Frankfurt / Germany and Phoenix – Arizona / USA.

Under temperate climates such as Frankfurt and Stockholm, TR2 transcritical system is the

most efficient system and is therefore perfectly suited for this range of ambient conditions. To

make a comparison less dependent on the use of a coolant circuit on the condenser, the

simulations were adapted and the benefits of direct condensation via a condenser/gas cooler

are visible in Figure 9.8. In all cases, the deletion of the coolant circuit is naturally advantageous

in terms of energy. However, its influence is less marked in the cascade which could be

explained by the flat COP lines with condensing temperature for CC1.


David Freléchox

91

383 403439 460

1060 1072

814

364 385

875 899

737

456418

344324

416385

0

200

400

600

800

1000

1200

StockholmTR1

StockholmTR2

StockholmCC1

FrankfurtTR1

FrankfurtTR2

FrankfurtCC1

PhoenixTR1

PhoenixTR2

PhoenixCC1

Ann

ual e

nerg

y co

nsum

ptio

n [M

Wh]

Coolant Gas cooler

Figure 9.8: Annual energy consumption with or without coolant loop for different supermarket systems in Stockholm / Sweden, Frankfurt / Germany and Phoenix – Arizona / USA.

Thus, for temperate climates, the TR1 with its advantage of using a free desuperheat process

between two compression stages on the low temperature unit, is the most efficient. It consumes

in all cases lower energy than the second transcritical TR2 system. The principle of the booster

used in TR2 therefore may be more convenient from installation point of view in order to reduce

the piping and therefore the costs than a real improvement in terms of energy. In very hot

climates like Phoenix, the use of a cascade is clearly beneficial with or without coolant loop.

With all the systems having a water/brine loop on the condenser/gas cooler side an additional

5K temperature difference is assumed between the heat transfer fluid and the ambient.

Accordingly, the temperature difference between the condenser/gas cooler exit and the ambient

becomes 10 K. The temperature limit for floating condensing is kept the same as in the

calculations above; 10°C of condensing or at the ga s cooler exit.

The values presented in Figure 9.8 can be related to each other in percentages. This is detailed

in Table 9-2 where the energy consumptions excess are related to the better system as a

percentage for each location. In the case of Stockholm weather conditions the transcritical TR1

condenser/gas cooler system solution has the lowest energy consumption over the year, 19%

less than R404A and 6% less than transcritical TR2 condenser/gas cooler system. This is also

the result in the case of Frankfurt, which can be observed in Figure 9.8. The difference between

the transcritical TR1 condenser/gas cooler system and the cascade CC1 condenser system is

little lower.


David Freléchox

92


condenser/ gas cooler coolant



TR1 100% 118% 100% 121% 119% 144% TR2 106% 124% 106% 126% 122% 145% CC1 119% 129% 114% 125% 100% 110%

Table 9-2: Annual energy consumption excess in percentage in comparison with the better system for each supermarket systems in Stockholm / Sweden, Frankfurt / Germany and Phoenix – Arizona / USA.

In the case of a hot climate, as Phoenix, the R404A/CO2 cascade system has the lowest annual

energy consumption; about 20% lower than for the two transcritical system using as well a

condenser/gas cooler. The two transcritical systems have almost the same energy

consumption. The difference between the systems when using a coolant loop is similar in case

of using a gas cooler.

As can be seen in Table 9-2, it is clear that differences in performances between the systems

depend on the ambient temperature. The choice of using direct condensation or a coolant loop

has also a big impact on the energy consumption.


David Freléchox

93

10 Specific system simulation

10.1 Impact of the evaporation temperature on COP

The impact on COP of the evaporation temperature is not insignificant. Its level must guarantee

the transfer of heat, but it is also important to maintain as high as possible to guarantee a high

COP. Currently, most manufacturers advise evaporation temperature of -10°C for the medium

temperature cabinets and -35°C for the low temperat ure cabinets. Nevertheless, the trend is

clearly to improve the cabinets and to increase these levels of few degrees.

A recent document of the firm EPTA GmbH [EPT09] presents a revolutionary technology using

an evaporation temperature of 0°C in order to keep fresh produce between 2 and 4°C. EPTA

evaluates the energy saving potential of this innovation of about 20%. Note that this article does

not refer to a specific refrigerant.

The Figure 10.1 shows the impact on COP from evaporation temperature for each system on

the TR1 supermarket. The pressure ratio is lower for the medium temperature system, so the

relative improvement in its COP is more prominent.

0.0

1.0

2.0

3.0

4.0

5.0

6.0

7.0

8.0

10 15 20 25 30 35 40 45 50


CO

P

COPm - To = -10°C COPm - To = -5°C COPm - To = 0°C

COPf - To = -35°C COPf - To = -30°C COPf - To = -25°C

Figure 10.1: COP for medium and low temperature system at different evaporation temperatures on the TR1 system


David Freléchox

94

The relative impact of evaporation temperature on the COP presented in the Figure 10.2

correlates well with the information from EPTA GmbH. The transition of the evaporation from

-10 to 0°C with a traditional condensing temperatur e between 35 and 40°C, give a COP

improvement of approximately 27%. This value corresponds reasonably well to the 20%

previously mentioned. Moreover, EPTA certainly consider all energy consuming parameters

associated to the cabinets, such as fans and lighting for example. Our simulation includes only

the power consumption of compressors and where applicable, brine pumps.

0.0%

10.0%

20.0%

30.0%

40.0%

50.0%

60.0%

70.0%

80.0%

90.0%

100.0%

10 15 20 25 30 35 40 45 50


CO

P im

prov

emen

ts [%

]

COPm improv. To = -5°C COPm improv. To = 0°C

COPf improv. To = -30°C COPm improv. To = -25°C

Reference:

Tevap_m = -10°C

Tevap_f = -35°C

Figure 10.2: Relative impact of the evaporation temperature on low and medium temperature systems with reference evaporation temperatures at -10°C and -35°C.

The development or use of evaporator and cabinets allowing the use of lower approach

temperature is needed to increase the performance of refrigeration systems. In addition, the use

of low condensing temperature which is now advocated makes the impact of evaporation

temperature level even higher.


David Freléchox

95

10.2 Optimal condensation temperature for the subcritica l / transcritical

operation transition

The discharge pressure regulation valve (Chapter 7.2) is primarily used to control an optimal

discharge pressure when the system is operating transcritically. From a thermodynamic point of

view, the transition between the sub- and trans-critical regimes is clear. Above the critical point,

31.1°C - 73.8 bars the system is working on a trans critical process. However, in the refrigeration

system, this transition can be applied before the critical point is reached in order to improve

system performances.

An increase in discharge pressure implies increase in compressor’s power consumption. It

usually involves a reduction of the cooling capacity. However during the transition subcritical /

transcritical and mainly owing to the special shape of the isotherms at this point, a small

increase in pressure does not lead to a decrease in cooling capacity actually it increases. The

additional cooling capacity is greater than the increase in compressor’s power consumption and

the coefficient of performance of the system is improved. The Figure 10.3 shows this effect on

an h-logP diagram. The transition at 28°C instead o f 31°C has a significant enthalpy decrease

at the condenser / gas cooler exit.

-250 -225 -200 -175 -15037

100

149

h [kJ/kg]

P [b

ar]

10°C

20°C

25°C

31°C

CarbonDioxide

Tc 27°C - Ttrans 28°CTc 27°C - Ttrans 28°C

Tc 30°C - Ttrans 31°CTc 30°C - Ttrans 31°CTc 30°C - Ttrans 28°CTc 30°C - Ttrans 28°C

Figure 10.3: Enthalpy condenser – gas cooler outlet at different condensation temperature and for two transition temperatures 28 and 30°C.

The function used to define the optimum discharge pressure is based on the doctoral thesis of

Samer Sawalha [SAW08] and shows as Equation 10-1.

1.6)(7.2 , −∆+⋅= appgcambopt TTP 10-1


David Freléchox

96

Thus, once the temperature exceeds the transition, the minimum discharge pressure is set up to

the critical point at 74 bars and this function is used to define the optimum discharge pressure

according to changes in ambient temperature. Note that from the addition of ambient

temperature and gas cooler approach temperature which is 5 K results the gas cooler outlet

temperature. This last value is preferable as set points in order to control the regulation valve.

The following Figure 10.4 presents the COP curve for different transition temperature. The

curves diverge only for a specific range. The curve which has the smoothest line should be the

better transition with the smallest losses. It also shows the changes in high pressure due to the

transition.

1

1.1

1.2

1.3

1.4

1.5

1.6

1.7

1.8

1.9

2

25 26 27 28 29 30 31 32 33 34 35Condensation temperature - Temperature gas cooler out [°C]

CO

P_t

otal

60

65

70

75

80

85

90

Pre

ssur

e [b

ar]

COP_tot_27 COP_tot_28 COP_tot_29 COP_tot_30 COP_tot_31Pgc_27 Pgc_28 Pgc_29 Pgc_30 Pgc_31

Load ratio = 3

Subcooling = 0 [K]

SH int = 10 [K]

SH ext = 10 [K]

Figure 10.4: Total COP and discharge pressure versus the condensing temperature for different transition temperatures with the supermarket TR2 system (2Transcritical booster units and one chiller unit)

On the figure above, the smooth shape of the COP curve using a transition temperature of 28°C

can be seen. COP losses resulting from a lower or higher transition temperature can be

observed in the plot. These losses are quite small and for a specific condensation temperature

range from 28 to 31°C, however, they are not neglig ible. They occur at high condensation

temperature when the refrigeration system must generate its full potential. The improvement of

the total coefficient of performance can achieve 7% with 28°C as transition temperature instead

of 31°C; this can be clearly seen in the plot in Fi gure 10.5. Conclusion, in reality, the optimum

transition temperature from subcritical to transcritical is 28°C and not 31°C. This corresponds to

the condensation temperature or gas cooler exit temperature. This improvement avoids

efficiency discontinuities at regime change.


David Freléchox

97

-1.00%

0.00%

1.00%

2.00%

3.00%

4.00%

5.00%

6.00%

7.00%

8.00%

25 26 27 28 29 30 31 32 33 34 35Condensation temperature - Temperature gas cooler out [°C]

CO

P im

prov

emen

t [%

]

COP improvement with transition at 28°C

Figure 10.5: Total COP improvement in percentage with a transition temperature at 28°C instead of 31°C related to the Figure 9.2 as usual for a load ratio of 3.

It is also imperative to maintain a condensing temperature or gas cooler outlet temperature as

low as possible. The use of particularly efficient condenser / gas cooler allows important COP

improvements. The temperature difference between outside temperature and gas cooler outlet

temperature is defined as the approach temperature (in the case of a direct interchange

air/CO2). Measurements on existing system TR2 give values between 4 and 5 K. But in order to

be independent of this influence, we decided to present the results according to the

condensation temperature.

10.3 Potential of desuperheating for low stage

The temperature of the compressed fluid after the compressor is generally very high, about 80 -

120°C. It depends on the properties of each fluid, on the discharge pressures and on the

temperatures and pressures at the compressor inlet. Before the condensation process the fluid

must be desuperheated.

When the refrigeration system uses a two-stage compression with a booster system, a cascade

system, or a two-stage compressor, it is possible to desuperheat the fluid between the two

stages or at the medium temperature level. This will improve the COP and therefore reduce

operating costs. This procedure could be called free desuperheat. In regard to the systems we

have studied, TR1 uses a process of free desuperheat between the first and second

compression stage on the low temperature system. TR2 could use the free desuperheat on the

booster, and CC1 could use the free desuperheat on the CO2 subcritical system, however, this

is not applied in the systems analyzed.


David Freléchox

98

A simulation has allowed assessing the potential for improving the total COP, when the free

desuperheating is used on the low temperature system in the CC1 supermarket. This system

has a coolant circuit. We have identified the two approach temperatures at 5 K, bringing the

difference between the outside air and CO2 desuperheater outlet temperature to 10 K. Those

approach temperature differences are at the gas cooler and the heat exchanger between the

two compression stages where desuperheating takes place. Depending on the external

temperature, the temperature of coolant entering in the heat exchanger vary and therefore the

COP too.

Figure 10.6 demonstrates the significant increase of the refrigeration system COP. In this case,

the use of free desuperheating allows a COP gain of up to 7%. Of course, this gain is influenced

by the outside temperature, which is used as heat sink. Note that the simulated system uses the

floating condensation on the R404A medium temperature unit up to a limit of 10°C. The CO2

low temperature system keeps the condensation at about -5°C.

1

1.5

2

2.5

3

3.5

-10 -5 0 5 10 15 20 25 30 35 40


CO

P to

tal

COP tot_no_free_desuperheat COP tot_free_desuperheat

∆T app_air = 5 K∆T app_coolant = 5 KT cond_limit = 10 °C

Figure 10.6: Total COP for the supermarket CC1 with and without free desuperheat on the low temperature unit

There are therefore relatively simple opportunities to improve the COP of booster or cascade

systems when using the free desuperheating of hot gas on "low temperature" compressors. This

should be applied more regularly; especially when the refrigeration system is equipped with a

coolant loop.


David Freléchox

99

10.4 Potential of subcooling with ground heat sink in TR 2

The TR2 supermarket is equipped with a heat exchanger that allows subcooling the fluid at the

exit of the liquid tank through the use of groundwater. The energy transmitted to the water can

then be used by the store’s heat pump. This simulation is based on a constant ground heat sink

temperature of 8°C. In reality, the water temperatu re varies with the seasons, the rejected heat

by the refrigeration system and the energy absorbed by the heat pump. The Figure 10.7 shows

the potential of subcooling with ground heat sink and its effect on the total COP.

-10

0

10

20

30

40

50

60

70

80

-10 -5 0 5 10 15 20 25 30 35 40

Ambient temperature [°C]

Tem

pera

ture

[°C

] - S

ub c

apac

ity [k

W]

0

0.5

1

1.5

2

2.5

3

3.5

4

CO

P to

tal

Condensation temperature Q subcooling booster Q subcooling mediumCOP total COP tot no ground heat sink

Temperature ground heat sink = 8 °C

Cooling capacity booster = 150 kWCooling capacity medium = 50 kW

Figure 10.7: Effect of the ground heat sink when it is using to subcool the liquid in TR2 supermarket

As can be seen on the figure, the use of groundwater in order to subcool the liquid is effective

for ambient temperatures condition higher than 10°C . Below which, the outside temperature is

colder than the heat sink source, so it does not make sense to try to achieve a heat exchange.

Between 15 and 35°C outside temperature, the use of the cold source is justified and can

improve the COP up to 50%. Therefore, a maximum use of the subcooling capacity during

summer is justified, but naturally it depends of the heat sink’s capacity. As soon as the ground

heat sink temperature increases, the system's efficiency will drop and this could cause problems

related to the environment. Note that if the high pressure side is forced high in order to recover

a maximum of energy from the refrigeration system to heat the supermarket, as it is the case

during winter in the supermarket TR2 then the use of the heat sink for an outside temperature

below 10°C could be justified. That partly offset t he losses on the COP due to the higher

pressure ratio. At the same time, the energy transmitted to the ground heat sink is immediately

used by the heat pump that extracts this energy to heat the supermarket.


David Freléchox

100

11 Discussion

This study reveals some important points and provides opportunities to continuously improve

the efficiency of refrigeration systems. The focus was set on the natural fluid CO2 which limits

drastically the direct contribution of refrigeration to the greenhouse effect. Its GWP is also the

base unit of measure and equal to 1. It reduces the impact on environment by a factor of

several thousand comparing with other traditional fluids.

Mainly due to its high pressure, CO2 presented a new challenge to the market players in

refrigeration and air conditioning. Most of the components must be improved or completely re-

developed. This process is being carried out by the main manufacturers. Compressors and

most of the valves were well adjusted to fit the use of CO2; however, improvements on cabinet

evaporators and gas coolers are still needed. One of the main problems of the CO2 technology

currently applied in supermarkets is the use of evaporators poorly adapted to this new fluid.

Indeed, tubes diameters and the layout of the evaporator is the same as those used with

traditional fluids. Thus, in order to create enough pressure drops in the evaporator, the

manufacturers voluntarily extend the circuit of the evaporator with a serial connection. Certainly

if smaller and more appropriate tubes were used, a traditional configuration using series /

parallel combinations would provide a better utilization of the exchange surface, or even raise

the temperature of evaporation, which has a direct effect on the energy efficiency. On the other

hand, refined design evaporators would certainly match the internal superheat used with

traditional fluids. Thus, the observed values on CO2 systems around 12 K could be reduced to

7 K. 7 K is a standard value for R404A systems and has been observed in at least one of the

R404A installation at the IWMAC interface.

Field measurements on three supermarkets have been carried out. Two supermarket use CO2

in a transcritical mode and one with a cascade R404A/CO2. The comparison show the highest

COP of about 4.5 on the medium temperature unit and 1.6 on the low temperature unit for the

supermarket using CO2 transcritical TR1. This is mainly du to the use of floating condensation

on this system which is an important advantage to the others. Therefore, the use of the floating

condensation is almost mandatory in the implementation of the fluid in order to achieve high

efficiency. The current limit for condensation temperature is set at 10°C, which should be further

lowered to improve the benefit of CO2 compared to traditional fluids in cold regions. CO2 take

advantage of its high working pressure, the pressure drop across the expansion valve is still

higher than with R404A as example and it allow reducing the condensation temperature

furthermore.


David Freléchox

101

However, several installations combine cooling systems with the heating system of the store.

This creates conflicts between the optimization of refrigeration COP and the maximization of

heat recovery. At least two of the facilities investigated in this thesis increase the condensing

temperature to recover more heat during winter. TR2 system raises its high pressure and uses

a borehole to subcool the liquid and avoid the COP losses. Note that the borehole was also

connected to the heat pump of the store. The second system CC1 keep its high pressure at a

high level during the whole year. We have not the possibility to calculate the capacity which is

recovered, so it is difficult to comment on this. As shown in this study, the negative impact of

condensation temperature rise is clearly more pronounced for CO2 systems.

Free heat sink sources as a borehole are particularly favourable in case of carbon dioxide.

Indeed, the COP of a CO2 installation is quite sensitive to subcooling. One Kelvin subcooling

could easily result in 2 times higher effect on COP in the case of CO2 than with R404A. This is

primarily applicable to condensing temperature near the critical point. Thus, the subcooling

potential of a borehole should mainly be used in summer.

Simulations were used in order to unify the working conditions for each supermarket and thus

allow fair comparison between the three systems. Firstly, the simulations were validating with

field measurement. Several comparisons have been carried out. It particularly shows the

influence of the coolant loop on the efficiency. On the function COP versus condensation

temperature, TR1 and TR2 systems have the better COP until the crossover at 25°C

condensation temperature when the cascade system CC1 has the highest COP. When the

influence of the coolant loop is inserted on the plot COP versus outside temperature, then the

TR1 system without coolant loop has the highest COP until 23°C outside temperature. This

transition of the highest COP between transcritical system and cascade system result from a

simulation and in reality the good heat transfer capacities of CO2 could favour it furthermore.


David Freléchox

102

Using this simulation and the outside temperature of three different climates which are

Stockholm, Frankfurt and Phoenix an annual simulation has been done. In most countries of

Western Europe the technology of transcritical carbon dioxide is more appropriate than cascade

system R404A/CO2 and reduces the energy consumption of refrigeration systems of about 15 –

20 %. In warmer climate, with temperatures averaging around 20°C, the carbon dioxide in

transcritical mode is less suitable. For very hot climates such as Phoenix/USA, the

overconsumption of a CO2 transcritical installation compared to a cascade system is about

20%. Thus, the elimination of synthetic fluids for this type of climate needs still to be analysed.

The TEWI analysis should be done then to investigate which system has less effect at the

environment. In this analysis the direct and indirect impact are included. Otherwise, an efficient

environmentally friendly replacement for the R404A could be ammonia in a cascade system

with CO2. Elimination may not be the best way; however, limitations on the amount of R404A

used would favour certain environmentally friendly solutions.

Our study also demonstrated some improvements possibilities on CO2 transcritical systems.

Particularly, the importance of the transition temperature applied in the regulation of the

condensation level. A transition temperature at 28°C instead of 31°C will help avoiding increase

in the enthalpy at the evaporator inlet and thus may lead under certain conditions to a 7%

increase of the COP. Another possibility is to desuperheat the fluid at the outlet of the first stage

on a two stage compressor. This leads to a gain up to 7% of the COP. The use of bigger and

high efficient evaporator leads also to a COP improvement of 12 % due to an increase of 5 K

evaporation temperature from -10 to -5°C.

At a practical level, a modification has come to our attention in a positive way. Installing a

frequency converter on compressors, which allow a better control of the system and give

easiest way to fix the evaporation temperature, has lead to a reduction of about 10 % of the

electrical power consumption on medium temperature systems. This is due to the settings

changes and better stability of the evaporating temperature which has led to a few kelvins

reduction of the approach temperature difference. Stable operating conditions give the

possibility to improve most of the settings and adjusting the speed of the compressors at the

request load fits with the optimization process.


David Freléchox

103

12 Conclusions and suggestions for future Works

This thesis evaluated different field installations with different refrigeration system solutions.

Computer simulation models for each of the systems under investigation have been run. The

systems under investigation have been compared based on the results from the field

measurements and the computer simulation models. This work compares three different field

installations and shows the difficulties to make comparison on the base of field measurement.

Various different parameters such as the load ratio, the use of a borehole, the heat recovery

requirements or the heat recovery system influence the COP of the system and thus make the

comparison difficult.

This study demonstrated using our simulation models and according to the field measurements

that at low condensing temperatures the CO2 transcritical system reach the highest COP and at

high condensing temperatures the cascade R404A/CO2 system is more efficient. The crossover

on the COP between these two solutions is floating between 15 and 27°C depending on the

absence of a coolant loop, or the use of free desuperheating between two compression stages,

or is influenced by the heat exchange proprieties of the fluid.

In addition, the floating condensation and lowering the condensing temperature limit allows a

significant increase in the COP. On a transcritical system, the use of floating condensation up to

10°C allow reaching a total COP of 3.7 with a load ratio of 3.

In the near future and within this project, other installations will be analyzed and simulated

following the same procedure as adapted in this thesis. Specifically, one more cascade system

with variable speed pumps and a supermarkets using CO2 pump circulation technology. Also

analysis of supermarkets using traditional technologies with only synthetic fluids, especially

R404A is underway. This will allow comparison of all systems, new and old technologies,

natural and synthetic refrigerants.

The experimental and theoretical studies reported in this thesis prove that CO2 based system

solutions investigated can be efficient solutions for supermarket refrigeration; however,

comparison with traditional systems is needed and will be presented in following publications in

the ongoing project.

Freléchox David Stockholm

13 August 2009


David Freléchox

104

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Appendix 1: Formula COPtot with load ratio correcti on

Formula:

chcompfrcomp

chofrotot EE

QQCOP

__

__

&&

&&

++

=

LR

QQ

Q

QLR cho

frofro

cho __

_

_&

&&

&

=→=

chcomp

chom E

QCOP

_

_

&

&

=

fr

cho

fr

frofrcomp

frcomp

frofr COPLR

Q

COP

QE

E

QCOP

⋅==→= __

__

_&&

&&

&

Calculation:

chfrch

cho

fr

cho

chocho

chcompfrcomp

chofrotot

COPCOPLR

LR

COP

Q

COPLR

Q

QLR

Q

EE

QQCOP

11

11

__

__

__

__

+⋅

+=

+⋅

+=

++

=&&

&&

&&

&&

Demonstration

correctedLRLR

realLRLR

corr ==

chcompfro

chofrcomp

corr

corr

corrch

chcompcorrfrcomp

corrch

frchcompcorrfrfrcomp

corrchfr

frchcompcorrchfrcomp

corrchfr

chfrcorr

frchcompchfrcomp

corr

corr

ch

chcomp

frcorr

frcomp

corrLRtot

EQ

QE

LR

LR

LRQ

ELRLRE

LRQ

QELRLRQE

LRQQ

QELRQE

LRQQ

QQLR

QELRQELR

LR

Q

E

QLR

ELR

COP

__

__

__

____

_____

1

1

)1(

)1()1(

11

1

&&

&&

&

&&

&

&&&&

&&

&&&&

&&

&&

&&&&

&

&

&

&

⋅+⋅

+⋅=

⋅+⋅+⋅=

⋅⋅+⋅⋅+⋅⋅

=⋅⋅+⋅

+⋅⋅=

⋅⋅⋅⋅+⋅

+

=+

⋅

+=

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