Evaluation of CO2

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Evaluation of CO

2

Ice rink heat recovery

system performance

SOTIRIOS THANASOULAS

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Master of Science Thesis EGI 2018: TRITA-ITM-EX 2018:618

Evaluation of CO2 Ice rink heat recovery

system performance Thanasoulas Sotirios Approved Date Examiner Samer Sawalha Supervisor Jörgen Rogstam Commissioner Contact person

Master student: Sotirios Thanasoulas

Studentbacken 25 Läg 1310 11557 Stockholm

Registration Number: 910817-3239

Department Energy Technology

Degree program Sustainable Energy Engineering

Examiner at EGI: Samer Sawalha

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ABSTRACT

Ice rinks are the largest energy consumers in terms of public buildings due to their simultaneous need of cooling, heating, ventilation, and lighting for different parts of the building which means that these facilities also have a lot of potential for energy saving. Due to the size of the cooling unit in an ice rink the refrigerant charge can become quite high, which potentially has a big impact on the environment. CO2 refrigeration units could cover all these challenges that are linked to ice rink

operation. CO2 as a refrigerant has a very low impact on the environment and at the

same time it could provide enough energy to cover the heating demands of an ice rink. CO2-based systems should operate in trans-critical mode which affects the

performance of the refrigeration system, but by using the released heat that otherwise would be rejected to the ambience the total energy consumption becomes lower. The process of heat recovery is therefore vital for an efficient system. The refrigeration unit can produce enough energy to cover all the heating demands of an ice rink, but only when the heat recovery is controlled properly.

The energy recovery method is very important, but it should also be tailored in order to cover all demands. This is because all the subsystems, i.e. demands, have different temperature and load requirements. The energy could be recovered in one or two stages from the refrigeration system. However, hardware is not enough in order to achieve proper operation, the system should also operate in the best conditions (discharge pressure and subcooling) in order to be efficient. The more proper operation, the less energy consumption.

This energy recovery method could also be used as subcooling in climates where the ambient temperature is very high, making CO2 a very efficient solution. Regular

refrigerants are still often used in warm countries despite their high environmental impact. A refrigeration system using natural refrigerants and more specific CO2 does

not have constraints, however. The only limitation is the wrong operation.

Key Words: Ice rinks, CO2 as refrigerant, Heat recovery, Efficient systems,

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SAMMANFATTNING

Isrinkar är de största energikonsumenterna när det gäller offentliga byggnader på grund av deras ständiga behov av nedkylning, uppvärmning, ventilation och belysning. Detta innebär också att anläggningarna har en stor potential att effektivisera sin energibesparing. Isrinkar konsumerar stora mängder kylmedel på grund av deras storlekar, vilket potentiellt har en stor negativ inverkan på miljön.

CO2 kylenheter skulle kunna klara av alla dessa utmaningar som är kopplade till isrinkens drift. Att använda CO2 som en kylarvätska har en ytterst liten inverkan på miljön och kan dessutom bidra med tillräckligt mycket energi för att täcka uppvärmningsbehovet för en isrink.

CO2 baserade system bör köras i ett transkritiskt läge vilket påverkar kylsystemets prestanda, men genom att återanvända den utsläppta värmen som annars skulle gå förlorad till omgivningen så blir den totala energiförbrukningen lägre. Värmeåtervinningsprocessen är därför avgörande för ett effektivt energisystem. Kylaggregatet kan producera tillräckligt med energi för att täcka alla värmebehov för en isrink, men endast när värmeåtervinningen behärskas ordentligt.

Energiåtervinningsmetoden är också väldigt viktig, men den bör skräddarsys för att täcka alla krav. Detta beror på att alla delsystem, dvs krav, har olika temperatur- och belastningskrav. Energin kan återvinnas i ett eller två stadier från kylsystemet. Tyvärr så räcker dock inte hårdvaran till för att uppnå en önskad drift, men systemet bör även fungera under de bästa förutsättningarna (utloppstryck och underkylning) för att vara effektiv. Ju bättre drift, desto mindre är energiförbrukningen.

Denna energiåtervinningsmetod kan också användas som underkylning i varma klimat vilket gör CO2 till en mycket effektiv lösning. Vanliga typer av kylmedel används fortfarande ofta i varma länder trots att deras negativa miljöpåverkan. Ett kylsystem med ett naturligt kylmedel som till exempel koldioxid har emellertid inga begränsningar. Den enda begränsningen är den felaktiga hanteringen av driften.

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Acknowledgment

I would like to express my special gratitude to those who had given me their contribution, in some way, to complete this master project report. Particularly, I would like to express my deep appreciation and respect to Jörgen Rogstam, for his kind support, valuable lessons and for never getting tired of my endless questions.

I would like to thank Simon Bolteau and Cajus Grönqvist, who helped me on everything I asked them. I would like to thank my master thesis supervisor in KTH, Samer Sawalha, for the discussions and the value comments that helped me achieve this project.

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Figure Table

Figure 1 Regular ice rink demands ... 11

Figure 2 Indoor ice rink at Dallas’ Galleria (MAKHNATCH, 2010) ... 13

Figure 3 Europe’s largest outdoor ice skating rink in City Park (MAKHNATCH, 2010) ... 14

Figure 4 Portable ice rink for exhibition games in Tokyo, Japan (MAKHNATCH, 2010) ... 14

Figure 5 Type of Arenas (MAKHNATCH, 2010)... 15

Figure 6 Energy demand per sector (Rogstam, J., 2010) ... 16

Figure 7 Indirect and direct systems (Zhongyuan Zhang 2012) ... 16

Figure 8 Ice rink refrigeration scheme with heat recovery in condensing loop ... 17

Figure 9 Refrigerants’ saturation pressure vs saturation temperature (Sawalha2008) 18 Figure 10 Refrigerants’ volumetric refrigeration effect (Sawalha2008) ... 19

Figure 11 CO2 refrigerating cycle in p-h diagram for different after gas cooler temperatures (Sawalha, 2008) ... 20

Figure 12 Refrigerating cycle layout-1 stage heat recovery ... 21

Figure 13 Heat recovery-water loop in Gimo (Bolteau, Rogstam and Tazi, 2016) ... 22

Figure 14 Refrigerating cycle layout-2 stage heat recovery ... 23

Figure 15 2 stages heat recovery-water loops ... 23

Figure 16 CO2 unrecovered portion vs de-superheater outlet Temperature ... 25

Figure 17 CO2 p-h diagram ... 25

Figure 18 CO2 unrecovered portion vs de-superheater outlet Temperature ... 26

Figure 19 Water’s recovered portion vs de-superheater outlet Temperature ... 26

Figure 20 Water’s recovered portion vs de-superheater outlet Temperature ... 27

Figure 21 Unrealistic CO2 and water Temperature profile ... 27

Figure 22 CO2 and water Temperature profile ... 28

Figure 23 de-superheater temperature profiles (Industrialheatpumps.nl, 2018) ... 29

Figure 24 Approach Temperature ... 29

Figure 25 CO2 and water Temperature profile, 40% recovery in the high temperature De-superheater ... 30

Figure 26 CO2 and water Temperature profile, 30% recovery in the high temperature De-superheater ... 31

Figure 27 CO2 profile for different Head Pressure ... 31

Figure 28 CO2 Heat Capacity for different Pressure ... 32

Figure 29 CO2 Density for different Pressure ... 32

Figure 30 CO2 and water profile for different Head Pressure ... 33

Figure 31 CO2 and water profile Temperature ... 33

Figure 32 CO2 profile Temperature for different discharge Temperature, constant head Pressure (90 bar) ... 34

Figure 33 CO2 and water profile Temperature,1 stage HR with and without IHE, 90 bar Head Pressure... 35

Figure 34 CO2 and water profile Temperature,1 stage HR with and without IHE, 90 bar Head Pressure... 36

Figure 35 Subcooling in CO2 cycle (Sawalha, 2013)... 37

Figure 36 Influence of subcooling in the discharge pressure (Sawalha, 2013) ... 37

Figure 37 CO2 system optimum performance ... 38

Figure 38 Influence of subcooling in the COP ... 38

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Figure 40 Gimo operation drawing ... 40

Figure 41 Sensors’ placement using climaCheck metod ... 41

Figure 42 Temperature after De-superheater, season 2017-2018 in Gimo ... 42

Figure 43 Difference of enthalpy for different Temperature after De-superheater ... 43

Figure 44 Recovered Energy, season 2017-2018 in Gimo ... 44

Figure 45 Temperature before expansion device, season 2017-2018 in Gimo... 44

Figure 46 Power consumption, season 2017-2018 in Gimo ... 45

Figure 47 mass flow, season 2017-2018 in Gimo... 45

Figure 48 COP_HR, season 2017-2018 in Gimo... 46

Figure 49 Ventilation capaciry, season 2017-2018 in Gimo ... 46

Figure 50 Cooling load, season 2017-2018 in Gimo ... 47

Figure 51 Indoor temperature, season 2017-2018 in Gimo ... 48

Figure 52 volumetric flow in water loop, season 2017-2018 in Gimo ... 49

Figure 53 Supply and return Temperatures in water loop, season 2017-2018 in Gimo ... 49

Figure 54 Dorin CD 350H ... 52

Figure 55 de-superheater temperature profiles principle ... 53

Figure 56 Optimum operation ... 54

Figure 57 Discharge and after de-superheater temparatures ... 54

Figure 58 Optimum operation ... 55

Figure 59 Discharge and after de-superheater temperatures ... 56

Figure 60 de-superheater temperature profiles,discharge pressure lower than 80 bar 56 Figure 61 de-superheater temperature profiles,discharge pressure 94 bar ... 57

Figure 62 Rejected energy in Gas cooler ... 58

Figure 63 Optimun operation ... 59

Figure 64 Discharge and after de-superheater temperatures ... 59

Figure 65 1st de-superheater 77 bar discharge pressure ... 60

Figure 66 2nd de-superheater 77 bar discharge pressure ... 61

Figure 67 1st de-superheater 81 bar discharge pressure ... 62

Figure 68 2nd_{de-superheater 81 bar discharge pressure ... 62}

Figure 69 Discharge pressure comparison between 1 and 2 stages heat recovery ... 63

Figure 70 COPc comparison between 1 and 2 stages heat recovery ... 63

Figure 71 COP_HR comparison between 1 and 2 stages heat recovery... 64

Figure 72 Global COP comparison ... 64

Figure 73 Control of operation according Heat recovery demand ... 65

Figure 74 2 stage optimum control ... 66

Figure 75 Optimum operation with and without IHE, 1 stage (20-60) ... 66

Figure 76 Global COP with and without IHE, 1 stage (20-60) ... 67

Figure 77 Optimum operation with and without IHE, 1 stage (20-70) ... 67

Figure 78 Global COP with and without IHE, 1 stage (20-70) ... 68

Figure 79 Optimum operation with and without IHE, 2 stages (20-40,40-70) ... 68

Figure 80 Global COP with and without IHE, 2 stages (20-40,40-70)... 68

Figure 81 de-superheater conditions ... 69

Figure 82 Temperature profiles from CAS ... 70

Figure 83 de-superheater options ... 70

Figure 84 de-superheater conditions ... 71

Figure 85 Temperature profiles from CAS ... 71

Figure 86 de-superheater options ... 71

Figure 87 1st Desuperheaater conditions ... 72

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Figure 89 1st de-superheater options ... 73

Figure 90 2nd de-superheater conditions ... 73

Figure 91 2nd_{de-superheater temperature profiles ... 74}

Figure 92 2nd de-superheater options ... 74

Figure 93 Warm climate operation ... 75

Figure 94 Refrigerating cycle layout-subcooler ... 75

Figure 95 Subcooler-water loops ... 76

Figure 96 Subcooling effect in warm climates (TGC,out=40ºC) ... 76

Figure 97 2stages heat recovery for different Pinch Point ... 78

Figure 98 High Temperature de-superheater (3 ºC pinch point) ... 79

Figure 99 High Temperature de-superheater options (3 ºC pinch point) ... 79

Figure 100 Low Temperature de-superheater (3 ºC pinch point) ... 80

Figure 101 High Temperature de-superheater options (3 ºC pinch point) ... 80

Figure 102 Gimo Energy Data, season 2016-2017 ... 81

Figure 103 Gimo Energy Data before and after renovation ... 81

Figure 104 Hallevi ice arena ... 83

Figure 105 Connected water loops ... 84

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1. Introduction

Nowadays, Sustainable Energy systems are very important around the world. The climate is changing year to year into very cold winters and very warm summers even in countries like Sweden. Therefore, the need for more efficient systems drives to combined solutions. Refrigerating systems produce heating and cooling load the same time in many applications like ice rinks or supermarkets.

Environmental impact, as well as, investment and operation cost of a refrigeration unit are very important factors which should be taken under consideration when a renovation is decided. These applications are addressed to customers who take care of their profit, so systems which produce more heating and cooling load by consuming the same energy, are more preferable. What it is needed are high efficient systems.

Ice rink systems consume a lot of energy and it is a rapidly increasing application. The ice rink refrigeration technology must be optimized technically and economically in such a way that it may become more sustainable.

1.1. Background

Refrigerating systems, as well as, heat recovery solutions in supermarkets and ice rinks have been evaluated. A comprehensive knowledge base is available together with a number of real installations. These installations give real data which can be compared with previous theoretical studies.

Ice rinks operate for a long term which means that they operate in wide range of different ambient temperatures. According to the ambient temperature, ice rinks have different energy demand per sector. The five big sectors in an ice rink are refrigeration, heating, ventilation, dehumidification and lighting systems.

Figure 1 Regular ice rink demands (ROGSTAM, ABDI and SAWALHA, 2014)

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1.2. Objectives

The objective of this research is to evaluate the energy recovery potential of existing ice rinks, optimize the heat recovery systems and analyse the economic aspect of the ice rink operation in several locations with different climate (Temperature, Humidity etc.).

To reach this aim, several sub-objectives should be accomplished.  Literature study.

 Study the existing refrigerating and heat recovery system solutions to understand the challenges.

 Analyse and evaluate current calculation and evaluation methods  Compile relevant data from field measurements

 Study selected ice rinks with respect to the performance

 Propose modification in the hard ware design or in the system control  Provide economical study (including LCC)

1.3. Methodology

The methodology has been divided into several steps according to the objectives. The thesis work will start with the literature review of ice rink design technologies and existing saving technologies used in this area.

All the acquired information will be analysed using qualitative methods. The analysis results will be supplemented with the experimental studies in order to define most promising energy saving methods. Finally, both the experimental data and analysis results will be used to identify the energy saving actions and potential.

1.4. Scope and limitations

In this study, ice rink is evaluated from the energy aspect and the main interest lies in the refrigeration system. More specific, the heat recovery system is discussed. Scope of this research is the evaluation of existing ice rinks, which use heat recovery in order to cover heating demand. A comparison between these systems and ideal cycle will answer the question “How efficient is the heat recovery in these systems?”. The second part of this research is the modification of the heat recovery system as to become more efficient and/or a modification of the refrigerating cycle’s operation for a more efficient heat recovery.

Scope of this research is the answer of the question “How important is the heat recovery in warm climates and how much efficient is an ice rink with CO2 refrigerating system?”

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2. Ice rink technology

A heat transfer fluid is circulated through a network of pipes to provide required cool load in the ice rink surface. These pipes are located inside a concrete layer under the ice sheet. The heat transfer liquid can be the refrigerant in case of direct system or a brine like calcium chloride or glycol in case of indirect systems.

2.1. Types of Ice rinks Indoor Ice rinks

Ice rinks can be divided into three types. When ice rink is going to be used for longer period during a year and also environment is not suitable then the rink must be build indoor. Most of the ice rinks are now constructed indoor.

Figure 2 Indoor ice rink at Dallas’ Galleria (Travelocity; 1996-2005)

Outdoor Ice rink

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Figure 3 Europe’s largest outdoor ice skating rink in City Park (PBase; 1999-2005)

Portable Ice rinks

These rinks can be laid down where is necessary. The only requirement is to prepare the ground and level it carefully. Portable ice rinks can be transferred anywhere.

Figure 4 Portable ice rink for exhibition games in Tokyo, Japan (Los Tres Papagayos; 1999-2005)

2.2. Ice rink classification

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spectators’ capacity and so on. There are three different classifications of ice arenas in Sweden now (MAKHNATCH, 2010).

 Training Arenas. They are the least advanced arenas used for games in lower series and boys/girls matches.

 Spectator arenas are the ones which are built for the different activities up to Championship games. Therefore, these ice rinks imply existence of various equipment and service facilities.

 Event Arenas. The “Eliteserien” games are normally held in Event arenas. However, everything from large music events to exhibitions and motor shows could be held in this type of arenas.

Figure 5 Type of Arenas (Svenska Ishockeyförbundet 2009 )

2.3. Ice rink system

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Figure 6 Energy demand per sector (Rogstam, J., 2010)

2.4. Refrigerating system

There are basically two kinds of refrigeration systems in the ice rinks: direct and indirect refrigeration systems. In direct refrigeration systems the rink pipe is working as the evaporator while a heat exchange is used between the evaporator in refrigeration unit and rink pipe in indirect refrigeration system. The ice board is maintained by the secondary coolant in rink pipe.

Although a direct system is simpler with higher efficiency, the refrigerant in the direct system is limited by the health and safety risk of refrigerant leakage as for example the ammonia in several countries. Furthermore, a standard industrial refrigeration unit can be used in the indirect system which gives more flexibility to the choice of refrigerants (IIHF, 2011). So, the ice rinks now are built with an indirect refrigeration system mostly.

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3. Theoretical assessment

3.1. CO2 as refrigerant

Properties of CO2

CO2 was a very popular natural refrigerant because of its safe properties in comparison

to ammonia. The production of new synthetic chemical refrigerant, which had the ability to work in higher ambient temperatures in lower pressure, drove to a stop of CO2

usage. However, till 1990 CO2 became an alternative of CFC and HCFC replacement.

These artificial refrigerants have a high GWP and more specifically they have ozone depleting ability. (Padalkar A.S.2010). CO2 is toxic, flammable and

non-explosive gas while ammonia is mild flammability, has a pungent smell, and low threshold limit value.

Thermal-physical properties

Trans-critical operation with CO2 has high saturation pressure, as shown in figure 9, it

is 6 to 7 times higher than NH3 at 0 ºC. At the saturation temperature -10 ºC, the

operating pressure of CO2 is around 25 bar, so operating pressure of CO2 is often in

range between 25 bar when it is evaporating around -10 ºC and it can reach to 120 bar at high pressure side in the cycle.

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Other important properties of CO2 resulting from high operating pressure is a higher

volumetric refrigeration effect and lower vapor pressure drop than other refrigerants. Figure 10 shows the volumetric refrigeration effect of CO2 in comparison with other refrigerants. It is around 5 times higher than R22 and NH3 at 0 ºC

Figure 10 Refrigerants’ volumetric refrigeration effect (Sawalha2008)

CO2 has good thermophysical, transport properties and is safe for both environment and human in comparison to ammonia. The disadvantage of low performance of operating in trans-critical cycle motivate for a more efficient heat recovery system.

Optimum Performance in trans-critical CO2 cycle

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Figure 11 CO2 refrigerating cycle in p-h diagram for different Temperature after gas cooler temperatures (Sawalha, 2008)

So, if the cycle operates in high pressure, the compressor needs more energy to achieve this high difference of pressure without a significant higher cooling profit, which is inefficient. The reason is that isothermal carve of CO2 becomes almost vertical. In case

of low pressure, according to the diagram above the cooling potential could be significant higher by a small increasing in the high pressure which does not affect the energy consumption in the compressor dramatically.

According to Samer Sawalha the pressure for the optimum COP of a CO2 refrigerating

cycle is calculated by the equation above. (Sawalha, 2008)

𝑃_{𝑑𝑖𝑠𝑐ℎ,𝑜𝑝𝑡} = 2,7 ∗ 𝑇_{𝑔𝑐,𝑜𝑢𝑡}− 6,1

𝑇_{𝑔𝑐,𝑜𝑢𝑡} is the temperature after the gas cooler which is equal the ambient temperature plus the temperature approach of the gas cooler. This correlation gives the optimum operation for cooling purpose. But in the case of ice rinks, the optimum operation depends on an efficient heating and cooling load production.

3.2. Heat recovery in CO2 systems

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Figure 12 Refrigerating cycle layout-1 stage heat recovery

The quality of the energy that is recovered is very important. Each application needs a different temperature. (Bolteau, Rogstam and Tazi, 2016)

 Domestic hot water’s temperature should not be less than 55 °C.  Dehumidifier needs a temperature between 55 and 58 °C .

 Space heating demands a temperature from 35 till max 55 °C in very low outdoor temperature.

 Re-surfacing water should have a target temperature of 40 °C .  Ventilation needs 35 °C.

 Subfloor heating is between 20 and 30 °C .

Advantage of heat recovery in CO2 trans-critical cycle

 Higher discharge temperature, which drives to higher water supply temperature  Higher amount of energy that could be recovered

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Figure 13 Heat recovery-water loop in Gimo (Bolteau, Rogstam and Tazi, 2016)

The de-superheater (heat exchanger B) is connected to the heat recovery system which is divided in stages in order to provide the wanted temperature to each application. The first stage is the tank A for the domestic hot water (DHW) which has the highest temperature demand. After the DHW the required heat for the dehumidifier is supplied in a secondary circuit via a heat exchanger marked as C and next function is the radiator system D. Re-surfacing water is heated in the tank E and the ventilation is connected in the F circuit. The lowest temperature demand is achieved in the heat exchanger H before the return in the de-superheater.

Important solution is the preheating of the water in a tank G before the heating in tanks A and E, which gives the opportunity to be optimized the heat recovery even in low temperatures.

Two stages heat recovery

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Figure 14 Refrigerating cycle layout-2 stage heat recovery

Figure 15 2 stages heat recovery-water loops

In this case the first de-superheat is used to heat the high temperature water which is used for dehumidification and domestic hot water. The second de-superheater is used to heat the medium and low temperature tanks respectively. The medium temperature tank is used for space heating and the low for other lower temperature demand like ventilation or subfloor heating.

Theoretical study, De-superheater principle

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exchanger (de-superheater), but the temperature of the second medium (usually water) does not depend only on the refrigerants discharge temperature. The temperature profile of CO2 is a very important factor.

It is used the software “Simple CO2 one stage plant”, for the certain evaluation. This

software gives a very good evaluation of a refrigerating cycle using CO2. It is assumed

that the system operates in regular Swedish conditions. Evaporation temperature: -8 °C

Temperature after gas cooler: 10 °C

Compressor’s efficiency: 75 % (Dorin CD 350H) External superheat: 2 K

Explanation of heat transfer in the De-superheater

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Refrigeration side

It is assumed that 100% of recovery is achieved when the refrigerant has temperature 20ºC, because the lowest water return temperature could be around 20ºC, according to heat demand. The highest temperature of the de-superheater is the discharge temperature which is the refrigerant’s temperature immediately after the compressor. So, the total energy that could be recovered in a refrigeration system is equal to mass flow of the refrigerant multiplied by the difference of enthalpy between discharge temperature and 20ºC.

𝑄_{𝑟𝑒𝑐𝑜𝑣𝑒𝑟𝑒𝑑,𝑡𝑜𝑡𝑎𝑙} = 𝑚̇ ∗ (ℎ_{𝑑𝑖𝑠𝑐ℎ𝑎𝑟𝑔𝑒 𝑇𝑒𝑚𝑝𝑒𝑟𝑎𝑡𝑢𝑟𝑒−}ℎ_{20 º𝐶})

But the recovered energy depends on the refrigerant’s temperature after the de-superheater.

𝑄𝑟𝑒𝑐𝑜𝑣𝑒𝑟𝑒𝑑 = 𝑚̇ ∗ (ℎ𝑑𝑖𝑠𝑐ℎ𝑎𝑟𝑔𝑒 𝑇𝑒𝑚𝑝𝑒𝑟𝑎𝑡𝑢𝑟𝑒−ℎ𝑎𝑓𝑡𝑒𝑟 𝑑𝑒𝑠𝑢𝑝𝑒𝑟ℎ𝑒𝑎𝑡𝑒𝑟)

The recovered energy ratio is used to be projected how much energy of the total recoverable energy has been recovered, for a specific refrigerant’s outlet from de-superheater temperature.

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Figure 16 CO2 unrecovered portion vs de-superheater outlet Temperature

The graph in in Figure 16 illustrates the remaining unrecovered energy per outlet temperature. Outlet temperature is the temperature of the refrigerant after the de-superheater. The line is not linear because of different enthalpy steps for the same temperature drop.

Figure 17 CO2 p-h diagram

The difference of enthalpy for temperature drop from 50 ºC to 40 ºC is lower than the drop from 40 ºC to 30 ºC for the same pressure, as it is marked with red line in Figure 17. This is the reason why the CO2 temperature profile curve is almost horizontal

around 40 ºC for the case of 88 bar pressure. As it will be explained bellow this condition makes the heat recovery inefficient.

An example which could make the graph more understandable is the case in Figure 18. In this case, it has been recovered 30% of the total recoverable energy, so, the remaining unrecovered energy is 70%, as it is illustrated in the graph. In this case, the temperature

0 20 40 60 80 100 0% 20% 40% 60% 80% 100% D esu p er h ea ter ou tle t Temp er at u re

Recovered Energy ratio

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of CO2 after the de-superheater would be around 50 ºC. When more energy is recovered

the temperature after the de-superheater becomes even lower.

Figure 18 CO2 unrecovered portion vs de-superheater outlet Temperature

Water side

If it is assumed that the water has the same temperature difference. The water is continuously in liquid phase in 1 bar pressure.

Figure 19 Water’s recovered portion vs de-superheater outlet Temperature

The graph has been created in the same logic as the CO2 side. In this case the water’s

inlet temperature is 20 ºC and the total energy recovering has been achieved when the water’s temperature is around 100 ºC. The line is assumed to be linear because of the water’s liquid phase. If the temperature is higher than 100 ºC the line changes because

0 20 40 60 80 100 0% 20% 40% 60% 80% 100% D esu p er h ea ter ou tle t Temp er at u re

CO2 88 bar

0 20 40 60 80 100 120 0% 20% 40% 60% 80% 100% D esup erheater o u tlet Te m p erature

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of water’s two-phase conditions, but, the water does not exceed this temperature, in reality.

Figure 20 Water’s recovered portion vs de-superheater outlet Temperature

The graph shows the temperature of the water after the de-superheater, which is the water supply temperature. In the case of Figure 20, it has been recovered 40% of the total recoverable energy and the water supply temperature is around 50ºC.

Combination of two de-superheater sides

The water’s temperature should be higher than refrigerant’s temperature as the water supply temperature to have the same value as the discharge temperature. This is wrong and opposite of the second thermodynamic law. The water’s temperature should be lower than refrigerant’s temperature, every moment. So, the Figure21 case is wrong.

Figure 21 Unrealistic CO2 and water Temperature profile 0 20 40 60 80 100 120 0% 20% 40% 60% 80% 100% D esup erheater o u tlet Te m p erature

water side

0 10 20 30 40 50 60 70 80 90 100 0% 20% 40% 60% 80% 100% D esu p er h ea ter ou tle t Temp er at u re

CO2 88 bar

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This condition has as result the water supply temperature and the discharge temperature to have a difference of 30ºC or more. This situation makes the system inefficient and sometimes unable to cover temperature level demand. In some cases, the water supply temperature demand is higher than the achievable.

Figure 22 CO2 and water Temperature profile

Pinch Point

Pinch point refers to the closest point between the CO2 temperature profile and the

water temperature profile. It is very important for the heat transfer through the de-superheater, because the water temperature profile should be adapted according to the pinch point and the CO2 temperature profile. This means that the pinch point creates

some limitations in the heat transfer and disability of proper energy recovery (less recovered energy or lower supply water temperature).

According to heat exchanger manufacturers (Alfa Laval) the pinch temperature between the CO2 and the water loop could be 1 ºC. (Christesensen, 2014) This gives

very efficient solutions but the same time the heat exchanger should have very large size. In this study case, different pinch point temperatures will be analyzed.

The high approach temperature is a consequence of the pinch point in a heat exchanger. Approach temperature is called the temperature difference between water inlet and refrigerant outlet in the de-superheater.

CO2 88 bar

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Figure 23 de-superheater temperature profiles (Industrialheatpumps.nl, 2018)

Figure 24 Approach Temperature

In the figure above the approach temperature is 20 ºC. This means that the water return temperature should have a low value (10 ºC) for a good heat transfer in the heat exchanger. If the temperature does not have this low value, the heat exchanger is even more inefficient.

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Two stage heat recovery

The benefit of this solution is that the temperature in the water’s profile is tighter to refrigerant’s. This solution gives the opportunity of a higher water supply temperature. As it is illustrated in Figure 25, for the same condition the supply water could reach 80ºC instead of 60ºC (Figure 22) in case of one stage heat recovery.

Figure 25 CO2 and water Temperature profile, 40% recovery in the high temperature De-superheater It could be achieved even higher temperature in case that less energy was recovered in the first de-superheater (high temperature). But usually the high temperature energy demand covers the 40% of the total recovered energy.

CO2 88 bar- 2 stage heat recovery

88 bar water side

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Figure 26 CO2 and water Temperature profile, 30% recovery in the high temperature de-superheater

Discharge pressure impact

When the discharge pressure becomes higher, the discharge temperature becomes higher as well. This affects the CO2 temperature profile through the de-superheater. The

curve of the temperature profile is moved upwards when the head pressure is higher. This drives to a smoother curve without parts which approaches horizontal incline.

Figure 27 CO2 profile for different Head Pressure 0 10 20 30 40 50 60 70 80 90 100 0% 20% 40% 60% 80% 100% D esu p er h ea ter ou tle t Temp er at u re

CO2 88 bar- 2 stage heat recovery

88 bar water side 2nd stage 1st stage 0 20 40 60 80 100 120 0% 20% 40% 60% 80% 100% D esu p er h ea ter ou tle t Temp er at u re

CO2 temperature profile in desuperheater

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According to Figure 27 the horizontal part of the curve for lower head pressures is appeared when the temperature is around 30ºC. The reason is that CO2 is close to critical point. Around the critical point the specific heat capacity becomes too high, so it is needed more energy transfer in order the temperature to change value. This makes the enthalpy step bigger for the same temperature drop in figure 17.

Figure 28 CO2 Heat Capacity for different Pressure

Moreover, the CO2 density change becomes sharp for temperature 31 ºC. But when the refrigerant operates in head pressure higher than 74 bar, the density change is smoother.

Figure 29 CO2 Density for different Pressure

In that case the lower discharge pressure drives to a lower water supply temperature. It is used one stage of heat recovery, as it is illustrated in Figure 30.

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Figure 30 CO2 and water profile for different Head Pressure

The water supply temperature is lower, but this is not the only problem. An important problem is that the water return temperature should be much lower than 20 ºC in order to be recovered 100% of the recoverable heat. Otherwise some energy will be rejected in the gas cooler, as it is presented in Figure 31.

Figure 31 CO2 and water profile Temperature

The unrecovered heat is 17% of the total recoverable heat and the approach temperature in the de-superheater is more than 10 ºC. This makes the system significant inefficient. This is the reason why the system should operate in higher pressure even if the COPc

(cooling coefficient of performance) is lower. In case that the head pressure is 75 bar the COPc is 3.6 and in case of 90 bar head Pressure the COPc is 3.4, which is not a big

difference but in the second case the achieved water supply temperature is more than 10ºC higher and it is achieved 100% of recovery.

CO2-water temperature profile in desuperheater

90 bar

75 bar

Water side 90 bar Head Pressure

Water side 75 bar Head Pressure 0 10 20 30 40 50 60 70 80 90 0% 20% 40% 60% 80% 100% D esu p er h ea ter ou tle t Temp er at u re

CO2-water temperature profile in desuperheater

75 bar

water side 75 bar Head Pressure 17%

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Internal Heat exchanger

One of the modifications to the basic CO2 cycle is to add internal heat exchanger(IHE) in the system. IHE is a heat exchanger which sub-cools/further-cools the refrigerant stream inlet to the expansion valve and increases the superheating of the vapor inlet to the compressor. This type of heat exchanger is often used in systems with long runs to convert the non-useful heat gains of cold return lines with providing some subcooling for high pressure supply lines. The raised temperature of the return lines also helps to decrease the water vapor condensation on these pipes. The same time provides superheating which makes safer the compressor’s operation. (KARAMPOUR and SAWALHA, 2014)

The effectiveness of all the IHEs is assumed to be 50% (Sawalha, 2008). Some other experimental measurements show a range of 34-46% effectiveness for an IHE including pressure drop and heat losses (Torrella et al., 2011). So, an IHE with effectiveness of 40% is a good assumption for this study.

In a scenario with an internal heat exchanger, COPc is almost constant even if the lower

mass flow, which is produced because of the produced subcooling. The reason is the pressure drop, which is created by IHE. This pressure drop could reach 2 bar before the compressor, when this is in totally use. This lower mass flow could drive in less heat recovery, but the higher difference of enthalpy equalizes this drop. Eventually, the recoverable energy is 2% more when the IHE is 100% in use.

In this scenario the CO2 profile curve changes for temperature higher than 40ºC. The result is a smoother curve.

Figure 32 CO2 profile Temperature for different discharge Temperature, constant head Pressure (90 bar) The water supply temperature is just 3 ºC higher in the case of the IHE, even if the difference between the discharge temperatures, with and without IHE, is around 11 ºC,. In the first case the temperature is 67 ºC and in the second 64 ºC, respectively.

0 20 40 60 80 100 120 0% 20% 40% 60% 80% 100% D esu p er h ea ter ou tle t Temp er at u re

CO2 temperature profile 90 bar

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Figure 33 CO2 and water profile Temperature,1 stage HR with and without IHE, 90 bar Head Pressure IHE provides a small advantage in the system, but it provides superheating which makes the system safer.

Internal Heat exchanger+ two stages heat recovery

This scenario looks to fit better. The difference between these two water supply temperatures is 7ºC. In case of IHE the temperature could reach 90 ºC, but in scenario without IHE the temperature is 83 ºC. In both scenarios the temperature after the high temperature de-superheater is higher than 40 ºC.

CO2-water temperature profile 90 bar

0% IHE use 100% IHE use

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Figure 34 CO2 and water profile Temperature,1 stage HR with and without IHE, 90 bar Head Pressure It is interesting enough that 40% of energy could be recovered in the high temperature de-superheater, in the case of the IHE. But in the scenario without IHE the same quota is around 30%.

Influence of subcooling

The refrigerating cycle becomes more efficient after the use of subcooling, because the hs (enthalpy in expansion devise) becomes lower so the cycle can achieve the same

cooling capacity (Qcooling) having lower mass flow, which means less energy

consumption in the compressor. 𝑄_{𝑐𝑜𝑜𝑙𝑖𝑛𝑔} = 𝑚̇(ℎ₂− ℎ_𝑠)

The consequence is that the heat which could be recovered is lower as well, because of lower mass flow.

CO2 temperature profile 90 bar

0% IHE use 100% IHE use

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Figure 35 Subcooling in CO2 cycle (Sawalha, 2013)

The pressure should be higher as it is visible in the figure bellow, in order to be recovered the same heat.

Figure 36 Influence of subcooling in the discharge pressure (Sawalha, 2013)

Optimization of heat recovery

There are periods where more heating energy or more cooling load is needed. So, the refrigerating cycle should be manipulated, in order to be achieved the energy demands. The COPc doesn’t represent how efficient a cycle is. The reason is that the heating

recovery which is produced by a low COPc cycle can be more profitable than using two

different cycles which the one produce heating load and the other cooling load. This low COP cycle can perform even better than using extra auxiliary energy for heating. Samer Sawalha insert the definition of Heat Recovery Ratio in order to define the limits of the effectiveness of the COP decrease. (Sawalha, 2013)

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Figure 37 CO2 system optimum performance(Sawalha, 2013)

The HRR can be divided in 3 regions. The first one is when the CO2 cycle works in

sub-critical conditions and the recovery is not so effective because of low discharge temperature, but the cycle has a high COPc. COPc becomes lower when the Pressure

exceed the critical point but the same time the heat recovery is becoming significant higher. The heat recovery becomes significant high without high sacrifice of COPc for

pressure lower than 88 bar. After 88 bar the COPc decreases sharply and the cycle is

becoming inefficient. Moreover, the cycle is becoming unsafe because of high temperature. (Sawalha, 2013)

The principle of this optimization is the subcooling provision before the expansion valve. The subcooling can be provided with many ways like by using boreholes or by manipulating the capacity of the gas cooler’s fan.

Figure 38 Influence of subcooling in the COP(Sawalha, 2013)

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heating recovery is not needed and becomes less when the heating recovery demand is higher till the subcooling become zero.

Figure 39 COP change for different HRR

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4 Experimental measurements

4.1 Ice rink Introduction (Definition, location)

GIMO ICE RINK

In February 2013 the roof of the local Gimo ice rink in the community of Östhammar, Sweden collapsed. It left the community without one of its two ice rinks and the trouble to fit the popular ice hockey activity in to the remaining ice rink. In 2014, the building was restored; complete with a new roof and an innovative upgrade of the energy systems. The heart of the energy management being a trans-critical CO2 unit, putting the new and improved Gimo ice rink on the map as the first ice rink in Europe which uses pure CO2 technology.

Figure 40 Gimo operation drawing

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4.2 Methodology of measurement

Gimo is equipped with sensors around the arena as well as in the heat pump unit. The sensors are connected to IWMAC. IWMAC has a data base, so the data collection was achieved via this platform. The refrigeration cycle data were analysed using the clima check method. IWMAC provides temperature and pressure measurements in the refrigeration side. The data were converted into enthalpy, which is needed for the loads calculations, using REFPROP.

ClimaCheck method

The basic flowchart of ClimaCheck can be seen in Figure 41. For a simple basic refrigeration cycle, seven temperature sensors, two pressure sensors and one electrical power meter are used to determine the performance of the system from a thermodynamic point of view. The data which are measured are refrigerant temperatures and pressures before and after the compressor(s), air/water temperatures in and out from evaporator/condenser and refrigerant temperature before the expansion valve.

Figure 41 Sensors’ placement using climaCheck metod

Energy balance method

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external mass flow meter. The refrigerant mass flow rate is calculated by an energy balance over the compressor (Berglöf, 2010). By measuring the pressure and temperature before and after the compressor and the electricity input to the compressor it is possible to calculate the mass flow rate.

𝑚̇ =𝜂𝑒𝑙∗ 𝑃𝑒𝑙− 𝑄𝑐𝑜𝑚𝑝𝑡𝑟𝑒𝑠𝑠𝑜𝑟 𝑙𝑜𝑠𝑠 ℎ𝑐𝑜𝑚𝑝,𝑜𝑢𝑡 − ℎ𝑐𝑜𝑚𝑝,𝑖𝑛

Where:

𝑚̇: Refrigerant mass flow rate 𝜂_𝑒𝑙: Electric motor efficiency

𝑃𝑒𝑙: Electric power to the compressor motors

𝑄𝑐𝑜𝑚𝑝𝑡𝑟𝑒𝑠𝑠𝑜𝑟 𝑙𝑜𝑠𝑠: Heat loss from compressor body which is estimated 7% of the

Electric power consumption

ℎ𝑐𝑜𝑚𝑝,𝑜𝑢𝑡: Enthalpy after compressor

ℎ𝑐𝑜𝑚𝑝,𝑖𝑛: Enthalpy before compressor

4.3 Results

The water return temperature in a heat recovery system is used to be more or less 35 ºC. Assuming that the approach temperature in the de-superheater is around 5 K, the CO2 temperature after the de-superheater should be around 40 ºC. In case of Gimo ice

hockey arena, this temperature used to have this value until the middle of season 2017-2018, where the system had some operation and equipment changes.

Figure 42 Temperature after de-superheater, season 2017-2018 in Gimo

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How does it change the system?

It is expected that the system can recover more energy having the same mass flow and as a result the same compressor’s energy consumption.

Figure 43 Difference of enthalpy for different Temperature after de-superheater

The difference of enthalpy in case of 20 ºC CO2 temperature after the de-superheater is

higher than in case of 30 ºC, according to the above figure. 𝑄𝑟𝑒𝑐𝑜𝑣𝑒𝑟𝑒𝑑 = 𝑚̇(ℎ1𝑘− ℎ𝑑𝑒𝑠𝑢𝑝𝑒𝑟,𝑜𝑢𝑡)

The system has no changes and the discharge temperature is constant as well as the mass flow.

This gives a high Heat recovery COP. Heat recovery COP (COPHR) defines how

efficient is the heat recovery in the system.

𝐶𝑂𝑃_𝐻𝑅 = 𝑄𝑟𝑒𝑐𝑜𝑣𝑒𝑟𝑒𝑑 𝐸_{𝑒𝑥𝑡𝑟𝑎}

Where 𝐸𝑒𝑥𝑡𝑟𝑎 is the extra energy consumption in the compressor in order to be achieved

the certain 𝑄𝑟𝑒𝑐𝑜𝑣𝑒𝑟𝑒𝑑. The consumption in the compressor is higher in the beginning

as to be provided the best conditions for a heat recovery, like proper discharge pressure. This extra consumption is constant until the system changes, so 𝐸𝑒𝑥𝑡𝑟𝑎 is constant and

the COPHR becomes higher.

In which case, should the operation change?

The mass flow in the refrigerating system is manipulated according to cooling demand and the outdoor conditions. The enthalpy before the evaporator depends on the Temperature of the refrigerant before the expansion device. In case of no subcooler or internal heat exchanger this Temperature depends on the ambient temperature and the operating conditions of the gas cooler(condenser).

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In case of heat recovery, sometimes the recovered energy cannot cover the heating demand. The system should have higher mass flow, in order this higher demand to be achieved. This means that the mass flow depends on the heating demand. The conditions in the gas cooler are changing in order to be achieved this higher mass flow.

Figure 44 Recovered Energy, season 2017-2018 in Gimo

By changing the speed in the gas cooler’s air flow, the efficiency of the gas cooler becomes lower and as a result the temperature after the gas cooler becomes higher. Less efficiency means lower heat transfer and less condensing of the refrigerant.

Figure 45 Temperature before expansion device, season 2017-2018 in Gimo

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be higher as to be balanced the system. More power in the compressor is needed in order to be achieved this higher mass flow.

Figure 46 Power consumption, season 2017-2018 in Gimo

Figure 47 mass flow, season 2017-2018 in Gimo

4.4 Discussion of Results

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Figure 48 COP_HR, season 2017-2018 in Gimo

When the mass flow of the system must become higher the refrigerant’s temperature after the de-superheater is around 40 ºC and the COPHR is around 4,5. In case of constant

mass flow, when the refrigerant’s temperature after the de-superheater is around 20 ºC, the COPHR is around 7.

Consequences:

According to Figure 44, the recovered energy, when the system operates in higher COPHR , is lower. It is used less energy for the ventilation, but it is consumed the same

amount of energy, for the rest sub-systems.

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This destroys the energy chain of the arena, so the cooling capacity is not the same as it was supposed.

Figure 50 Cooling load, season 2017-2018 in Gimo

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Figure 51 Indoor temperature, season 2017-2018 in Gimo

Consequences in the Heat recovery

The water loop pump tries to balance the heating demand, so the mass flow is becoming lower. This lower water flow in the de-superheater drives to less efficient process. The supply water temperature is lower. But because of the energy demand for the other subsystems is the same, every subsystem operates in lower inlet and outlet temperatures. This could affect the efficiency of these subsystems.

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Figure 52 volumetric flow in water loop, season 2017-2018 in Gimo

Figure 53 Supply and return Temperatures in water loop, season 2017-2018 in Gimo

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5 Heat recovery evaluation

The best conditions for an ice arena will be analysed, according to the cooling and heating demand. Different scenarios will be analysed as to be achieved the most efficient. It is expected that the results will give an optimum system controlling per scenario. There are ice arenas which could operate more optimum even if their equipment is not that efficient.

EES has been used in order different operation cases to be simulated. The system should be as much closer to the reality. For that reason, data from existing ice arenas has been used in order the components to be defined. Gimo ice hockey arena has been used as reference case.

5.1 Methodology

The energy could be recovered in either one or two stages. In case of one stage the de-superheater is placed just before the gas cooler. This used to be common solution, but the same time is an inefficient solution. The reason is the temperature profile of CO2.

As it has already been explained, the difference of temperature in the pinch point should not be lower of a limitation, otherwise there is no heat transfer in the heat exchanger. The case of the two stages heat recovery gives a solution in that problem. Two de-superheaters are placed in serial as the recovered energy to be divided in two parts. So, the high temperature stage usually covers the 1/3 of the total requirement energy and the low temperature stage recovers the rest 2/3 of the demand.

5.2 Components Compressor

The total efficiency of the compressor is used, in order the compressor to be simulated. Different kind of efficiencies could be used. But total efficiency is an easy and accurate way to define the compressor.

The total efficiency presents how far from an isentropic operation is the current operation. Losses to the ambient are counted as well.

𝐸_{𝑐𝑜𝑚𝑝𝑟𝑒𝑠𝑠𝑜𝑟} = 𝐸𝑐𝑜𝑚𝑝𝑟𝑒𝑠𝑠𝑜𝑟,𝑖𝑠 𝜂_{𝑡𝑜𝑡𝑎𝑙}

𝐸𝑐𝑜𝑚𝑝𝑟𝑒𝑠𝑠𝑜𝑟,𝑖𝑠 = 𝑚̇ ∗ (ℎ1𝑘,𝑖𝑠− ℎ𝑠𝑢𝑐𝑡𝑖𝑜𝑛)

𝐸𝑐𝑜𝑚𝑝𝑟𝑒𝑠𝑠𝑜𝑟 = 𝑚̇ ∗ (ℎ1𝑘− ℎ𝑠𝑢𝑐𝑡𝑖𝑜𝑛)

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entropy for suction conditions. The compressor’s power consumption is known for every moment.

The efficiency in the compressor is related to the pressure ratio between the warm and the cold side. The compressor operates more efficient in specific pressure ratio and not for a very low or a very high pressure ratio. Dorin CD350H is used in case of Gimo. After evaluation of manufacturer’s data and counting 8% of heat losses the total efficiency of this type of compressor is:

𝜂_𝑡𝑜𝑡 = 0.017 ∗ 𝑃_{𝑟𝑎𝑡𝑖𝑜}3− 0.1705 ∗ 𝑃_{𝑟𝑎𝑡𝑖𝑜}2+ 0.5267 ∗ 𝑃_{𝑟𝑎𝑡𝑖𝑜}+ 0.1486

Figure 54 Dorin CD 350H

Evaporator

The evaporator has the same conditions for every different scenario. According to Gimo, which is an average ice rink in Sweden the average cooling demand is around 130 kW and the evaporator temperature around -8 ºC. It is assumed 1.2 ºC internal superheating, which is profitable for the evaporator and 2ºC external superheating, which creates losses in the system.

Gas Coolers

The energy, which isn’t recovered, is rejected via the gas cooler. The temperature after the gas cooler is estimated to be at least 10ºC. Regardless of the ambient temperature, the gas cooler’s fans operate as to cool the refrigerant down to 10 ºC. Obviously, the gas cooler works like subcooler in the system. The temperature of the refrigerant decreases which drives to lower enthalpy in the inlet of the evaporator. The evaporator, as it mentioned before, has constant capacity, so the mass flow and as a result the energy consumption is becoming lower. The temperature after the gas cooler is controlled by changing the fans speed (air mass flow). When the heating demand is becoming higher the need for subcooling is becoming lower. So the refrigerant’s temperature after the subcooler could be higher as to approach the de-superheater’s outlet temperature, which means that the gas cooler is out of operation.

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De-superheaters

The de-superheaters operate as to provide the certain water temperatures. The difference of temperature in the pinch point should not be lower than 2 ºC. The principle behind the de-superheater’s operation is illustrated in the figure bellow.

Figure 55 de-superheater temperature profiles principle

When the temperature difference between the refrigerant (red line) and the water (blue line) is not higher than 2 ºC the CO2 outlet is becoming higher as to be achieved this

limitation. This is how a high approach temperature, in a heat exchanger, could be explained.

5.3 Scenarios

5.3.1 Without Internal Heat Exchanger (IHE) One stage heat recovery

Using Gimo ice rink as reference scenario, the water return temperature could be 20 ºC and the water supply temperature should achieve 60 ºC. But regardless of the real superheater characteristics, it is assumed that in the most demanding cases the de-superheater could have 2 ºC difference of temperature in the pinch point. This may be an ideal condition, but it is acceptable in the reality, less efficient heat exchanger will be used later in this study.

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54 Results

The simulation shows that Gimo operates in the best conditions (pressure and subcooling). The refrigeration cycle operates in constant discharge pressure around 88 bar which follows the trend of the simulation which shows that the 87 bar of discharge pressure gives the best performance.

The system becomes less efficient, having the same discharge pressure but the subcooling in the gas cooler is less. The COPc becomes lower when the HRR is higher,

as it is visible in the figure bellow. Sacrificing the performance of the system a lit bit the heat recovery gain is higher.

Figure 56 Discharge pressure and COPc under optimum efficiency control with respect to HRR

Figure 57 Discharge and after de-superheater temperatures 1.5 2 2.5 3 3.5 4 4.5 5 50 60 70 80 90 100 110 120 20 40 60 80 100 120 140 160 C OP c Pr essu re (ba r) HRR(%)

1 stage heat recoverey (20-60)

Discharge Pressure COPc 0 20 40 60 80 100 120 140 20 40 60 80 100 120 140 160 Temp er at u re (º C ) HRR(%)

CO2 Temperature (20-60)

Discharge Temperature

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One stage heat recovery (20ºC -70ºC)

In some cases, the supply water temperature should be more than 60 ºC, because of higher demand on high temperature domestic hot water or other applications. So, in this scenario, the heat exchanger will be boosted to recover higher temperature energy. The temperature between the hot and cold side should be as much close as it is possible in higher operation range. This condition drives to a larger de-superheater, because of the need of a better UA value. The size analysis will be achieved later in this report.

Results

The trend of the system is similar to the case of 20-60 ºC water temperature. The difference is that the system operates in higher discharge pressure.

Figure 58 Optimum operation

In case of lower heat recovery ratio, the discharge pressure is low which drives to lower than 70 ºC discharge temperature. The discharge temperature should be higher than 73 ºC as to be achieved the heat transfer in the de-superheater. Even in this case the heat exchanger operates in very tight conditions. The temperature after the de-superheater is much higher than we would like to have. So, the system must reject energy that could be useful. 1.5 2 2.5 3 3.5 4 4.5 50 60 70 80 90 100 110 120 20 40 60 80 100 120 140 160 C OP c Pr essu re (ba r) HRR(%)

1 stage heat recoverey (20-70)

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Figure 59 Discharge and after de-superheater temperatures

In an ideal condition, the approach temperature in the de-superheater could be 2 ºC, so the CO2 temperature after the de-superheater should be around 22 ºC. But as it is

illustrated in Figure 60 the temperature after the de-superheater is more than 30 ºC until the discharge temperature becoming 100 ºC.

Figure 60 de-superheater temperature profiles,discharge pressure lower than 80 bar 0 20 40 60 80 100 120 140 20 40 60 80 100 120 140 160 Temp er at u re (º C ) HRR(%)

CO2 Temperature (20-70)

Discharge Temperature

Temperature after Desuperheater

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The temperature profiles of CO2 with red and water with blue are illustrated in the figure

above, in case where the discharge temperature is too low for an efficient operation. It is obvious that useful energy is rejected in the ambient via gas cooler.

Figure 61 de-superheater temperature profiles,discharge pressure 94 bar

In Figure 61, it is a more efficient case, where almost all the recoverable energy is recovered. In this case the discharge temperature should be 94 bar, which drives in a less efficient refrigerating operation with a COPc around 2.6. Even in this case the

difference in temperature between the discharge line and the water supply line is more than 20 ºC.

In the figure below, it is illustrated the conditions in the gas cooler that could explain how much recoverable energy is rejected because it cannot be recovered in the de-superheater. The system stops rejecting useful energy after the 120 % of HRR, in that point the rejected energy is 20 kW which produce subcooling and helps the performance of the refrigerating cycle. But in the previous conditions except for the subcooling, which demands almost always 20 kW, it is rejected more than 100 kW.

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Figure 62 Rejected energy in Gas cooler

Two stages heat recovery (20ºC-40ºC) and (40ºC-70ºC)

Using 1 stage heat recovery a lot of energy is rejected to the ambient. So, the solution of 2 stages heat recovery will show how profitable is the second de-superheater. In this scenario it is considered that 1 part of the energy is recovered in the first de-superheater and 2 parts in the second one. Usually the demand on low temperature energy is two times the demand on high temperature energy.

0 20 40 60 80 100 120 140 40 60 80 100 120 140 160 En er gy (k W) HRR(%)

Gas cooler capacity

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Results

Figure 63 Optimun operation

The system can recover energy, when the HRR is higher than 40 % in this scenario, but it could recover energy for lower HRR in the low stage (20-40 ºC), using the one of the two de-superheaters. It could be achieved the same result as the previous scenario in much lower discharge pressure (81 bar instead of 87 bar). The reason is that the two de-superheaters solution brings the water profile temperature closer to CO2 profile

temperature.

Figure 64 Discharge and after de-superheater temperatures 1.5 2 2.5 3 3.5 4 4.5 50 60 70 80 90 100 110 120 20 40 60 80 100 120 140 160 COPc Pr essu re (ba r) HRR(%)

2 stage heat recoverey (20-40),(40-70)

Discharge Pressure COPc 0 20 40 60 80 100 120 140 40 60 80 100 120 140 160 Temp er at u re (º C ) HRR(%)

CO2 Temperature (20-40),(40-70)

Discharge Temperature Temperature after 1st Desuperheater Temperature after 2nd Desuperheater

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As it is illustrated in Figure 64 the temperature difference in the inlet and the outlet of the 1st de-superheater is lower than the same temperature difference in the 2nd de-superheater, when the HRR is around 80%. This is because of the CO2 profile

temperature in the pressure of 79 bar.

Figure 65 1st_{de-superheater 79 bar discharge pressure}

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Figure 66 2nd_{de-superheater 79 bar discharge pressure}

But the heat transfer is better in both de-superheaters when the discharge pressure is higher than 81 bar. The 1st de-superheater should not be that efficient, in order to respect the 2nd de-superheater. As a result, the overall heat recovery operation will be improved, because heat rejection to the ambient will not be needed. The temperature difference in the pinch point could be higher than 2 ºC.

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Figure 67 1st_{de-superheater 81 bar discharge pressure}

Figure 68 2nd_{de-superheater 81 bar discharge pressure}

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Scenarios comparison

The two stages heat recovery has many benefits from every perspective. At first of all, the refrigerating cycle doesn’t need to operate in high discharge temperature, while the heat transfer is more efficient in two stages.

Figure 69 Discharge pressure comparison between 1 and 2 stages heat recovery

This lower discharge pressure drives to the higher COPc which is illustrated in the figure bellow.

Figure 70 COPc comparison between 1 and 2 stages heat recovery

The COP depends on the refrigerating operation, this does not give a clear view of the extra energy consumption of heating purposes. Heating recovery COP (COPHR) gives

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𝐸_{𝑐𝑜𝑛𝑠𝑢𝑚𝑝𝑡𝑖𝑜𝑛,𝑓𝑙𝑜𝑎𝑡𝑖𝑛𝑔 𝑐𝑜𝑛𝑑𝑒𝑛𝑠𝑒𝑟} is the compressor’s consumption in the same conditions without heat recovery.

Figure 71 COP_HR comparison between 1 and 2 stages heat recovery

In both cases the COPHR is high because the same device produces double work.

Otherwise, 2 different devices will reject a lot of energy in the ambient. But the system operates even better in 2 stages heat recovery. The difference of COPHR is around than

1.5 in 100 % HRR, which is a usually heating demand.

The COPHR is not enough to judge the performance of one system. The best way for

this comparison is the Global COP (COPGl), which defines the whole performance of

these systems.

𝐶𝑂𝑃_𝐺𝑙 = 𝑄𝑟𝑒𝑐𝑜𝑣𝑒𝑟𝑒𝑑+ 𝑄𝑐𝑜𝑜𝑙𝑖𝑛𝑔 𝐸_{𝑐𝑜𝑛𝑠𝑢𝑚𝑝𝑡𝑖𝑜𝑛}

Figure 72 Global COP comparison

The reference case of Gimo has a very good COPGL but in case that we demand even

Evaluation of CO2

Evaluation of CO

Ice rink heat recovery

system performance

SOTIRIOS THANASOULAS

ABSTRACT

SAMMANFATTNING

Acknowledgment

Contents

Figure Table

1. Introduction

2. Ice rink technology

3. Theoretical assessment

Explanation of heat transfer in the De-superheater

.

CO2 88 bar

water side

CO2 88 bar

CO2 88 bar

CO2 88 bar- 2 stage heat recovery

CO2 temperature profile in desuperheater

CO2-water temperature profile in desuperheater

CO2-water temperature profile in desuperheater

CO2 temperature profile 90 bar

CO2-water temperature profile 90 bar

CO2 temperature profile 90 bar

4

Experimental measurements

ClimaCheck method

Energy balance method

5

Heat recovery evaluation

1 stage heat recoverey (20-60)

CO2 Temperature (20-60)

1 stage heat recoverey (20-70)

CO2 Temperature (20-70)

Gas cooler capacity

2 stage heat recoverey (20-40),(40-70)

CO2 Temperature (20-40),(40-70)