Evaluation of CO 2 Ice rink heat recovery system performance SOTIRIOS THANASOULAS

Transkript

1 Evaluation of CO 2 Ice rink heat recovery system performance SOTIRIOS THANASOULAS Master of Science Thesis KTH School of Industrial Engineering and Management Energy Technology EGI TRITA-ITM-EX 2018:618 Division of Applied Thermodynamics and Refrigeration (ETT) SE STOCKHOLM

2 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 Commissioner Supervisor Jörgen Rogstam Contact person Master student: Sotirios Thanasoulas Studentbacken 25 Läg Stockholm Registration Number: Department Degree program Examiner at EGI: Supervisor at EGI: Energy Technology Sustainable Energy Engineering Samer Sawalha Samer Sawalha 2

3 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, Operation control, Warm climate 3

4 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. Nyckelord: Isrinkar, Koldioxid som kylmedel, Värmeåtervinning, Effektiva system, Driftstyrning, Varmt klimat 4

5 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. Finally, I would like to thank my family, for this long-term support. I could not forget my friend who were close to me in all this long educating trip. It was a very interesting experience to me, I met new people, I went out of my comfort zone multiple times, but they were close to me every time. 5

6 Contents 1. Introduction Background Objectives Methodology Scope and limitations Ice rink technology Types of Ice rinks Indoor Ice rinks Outdoor Ice rink Portable Ice rinks Ice rink classification Ice rink system Refrigerating system Theoretical assessment CO2 as refrigerant Properties of CO Thermal-physical properties Optimum Performance in trans-critical CO2 cycle Heat recovery in CO2 systems Advantage of heat recovery in CO2 trans-critical cycle Two stages heat recovery Theoretical study, De-superheater principle Explanation of heat transfer in the De-superheater Pinch Point Influence of subcooling Optimization of heat recovery Experimental measurements Ice rink Introduction (Definition, location) Methodology of measurement Results Discussion of Results Heat recovery evaluation Methodology Components Compressor

7 Evaporator Gas Coolers De-superheaters Scenarios Without Internal Heat Exchanger (IHE) One stage heat recovery Results One stage heat recovery (20ºC -70ºC) Results Two stages heat recovery (20ºC-40ºC) and (40ºC-70ºC) Results Scenarios comparison Comparison between theoretical data and real data (Gimo)- Proper controlling Warm climates With Internal Heat Exchanger (IHE) Real heat exchangers One stage heat recovery (20ºC-60ºC) One stage heat recovery (20ºC -70ºC) Two stages heat recovery (20ºC-40ºC) and (40ºC-70ºC) High Temperature De-superheater Low Temperature De-superheater Warm climate scenario Subcooling strategy Accuracy of Pinch Point Evaluation of higher pinch point temperature difference Statistics of existing Ice Arenas Gimo Hällevi ice rink Arena Umeå Conclusion Bibliography

8 Figure Table Figure 1 Regular ice rink demands Figure 2 Indoor ice rink at Dallas Galleria (MAKHNATCH, 2010) Figure 3 Europe s largest outdoor ice skating rink in City Park (MAKHNATCH, 2010) Figure 4 Portable ice rink for exhibition games in Tokyo, Japan (MAKHNATCH, 2010) Figure 5 Type of Arenas (MAKHNATCH, 2010) Figure 6 Energy demand per sector (Rogstam, J., 2010) Figure 7 Indirect and direct systems (Zhongyuan Zhang 2012) Figure 8 Ice rink refrigeration scheme with heat recovery in condensing loop Figure 9 Refrigerants saturation pressure vs saturation temperature (Sawalha2008) 18 Figure 10 Refrigerants volumetric refrigeration effect (Sawalha2008) Figure 11 CO2 refrigerating cycle in p-h diagram for different after gas cooler temperatures (Sawalha, 2008) Figure 12 Refrigerating cycle layout-1 stage heat recovery Figure 13 Heat recovery-water loop in Gimo (Bolteau, Rogstam and Tazi, 2016) Figure 14 Refrigerating cycle layout-2 stage heat recovery Figure 15 2 stages heat recovery-water loops Figure 16 CO2 unrecovered portion vs de-superheater outlet Temperature Figure 17 CO2 p-h diagram Figure 18 CO2 unrecovered portion vs de-superheater outlet Temperature Figure 19 Water s recovered portion vs de-superheater outlet Temperature Figure 20 Water s recovered portion vs de-superheater outlet Temperature Figure 21 Unrealistic CO2 and water Temperature profile Figure 22 CO2 and water Temperature profile Figure 23 de-superheater temperature profiles (Industrialheatpumps.nl, 2018) Figure 24 Approach Temperature Figure 25 CO2 and water Temperature profile, 40% recovery in the high temperature De-superheater Figure 26 CO2 and water Temperature profile, 30% recovery in the high temperature De-superheater Figure 27 CO2 profile for different Head Pressure Figure 28 CO2 Heat Capacity for different Pressure Figure 29 CO2 Density for different Pressure Figure 30 CO2 and water profile for different Head Pressure Figure 31 CO2 and water profile Temperature Figure 32 CO2 profile Temperature for different discharge Temperature, constant head Pressure (90 bar) Figure 33 CO2 and water profile Temperature,1 stage HR with and without IHE, 90 bar Head Pressure Figure 34 CO2 and water profile Temperature,1 stage HR with and without IHE, 90 bar Head Pressure Figure 35 Subcooling in CO2 cycle (Sawalha, 2013) Figure 36 Influence of subcooling in the discharge pressure (Sawalha, 2013) Figure 37 CO2 system optimum performance Figure 38 Influence of subcooling in the COP Figure 39 COP change for different subcooling

9 Figure 40 Gimo operation drawing Figure 41 Sensors placement using climacheck metod Figure 42 Temperature after De-superheater, season in Gimo Figure 43 Difference of enthalpy for different Temperature after De-superheater Figure 44 Recovered Energy, season in Gimo Figure 45 Temperature before expansion device, season in Gimo Figure 46 Power consumption, season in Gimo Figure 47 mass flow, season in Gimo Figure 48 COP_HR, season in Gimo Figure 49 Ventilation capaciry, season in Gimo Figure 50 Cooling load, season in Gimo Figure 51 Indoor temperature, season in Gimo Figure 52 volumetric flow in water loop, season in Gimo Figure 53 Supply and return Temperatures in water loop, season in Gimo Figure 54 Dorin CD 350H Figure 55 de-superheater temperature profiles principle Figure 56 Optimum operation Figure 57 Discharge and after de-superheater temparatures Figure 58 Optimum operation Figure 59 Discharge and after de-superheater temperatures Figure 60 de-superheater temperature profiles,discharge pressure lower than 80 bar 56 Figure 61 de-superheater temperature profiles,discharge pressure 94 bar Figure 62 Rejected energy in Gas cooler Figure 63 Optimun operation Figure 64 Discharge and after de-superheater temperatures Figure 65 1 st de-superheater 77 bar discharge pressure Figure 66 2 nd de-superheater 77 bar discharge pressure Figure 67 1 st de-superheater 81 bar discharge pressure Figure 68 2 nd de-superheater 81 bar discharge pressure Figure 69 Discharge pressure comparison between 1 and 2 stages heat recovery Figure 70 COPc comparison between 1 and 2 stages heat recovery Figure 71 COP_HR comparison between 1 and 2 stages heat recovery Figure 72 Global COP comparison Figure 73 Control of operation according Heat recovery demand Figure 74 2 stage optimum control Figure 75 Optimum operation with and without IHE, 1 stage (20-60) Figure 76 Global COP with and without IHE, 1 stage (20-60) Figure 77 Optimum operation with and without IHE, 1 stage (20-70) Figure 78 Global COP with and without IHE, 1 stage (20-70) Figure 79 Optimum operation with and without IHE, 2 stages (20-40,40-70) Figure 80 Global COP with and without IHE, 2 stages (20-40,40-70) Figure 81 de-superheater conditions Figure 82 Temperature profiles from CAS Figure 83 de-superheater options Figure 84 de-superheater conditions Figure 85 Temperature profiles from CAS Figure 86 de-superheater options Figure 87 1 st Desuperheaater conditions Figure 88 1 st de-superheater temperature profiles

10 Figure 89 1 st de-superheater options Figure 90 2 nd de-superheater conditions Figure 91 2 nd de-superheater temperature profiles Figure 92 2 nd de-superheater options Figure 93 Warm climate operation Figure 94 Refrigerating cycle layout-subcooler Figure 95 Subcooler-water loops Figure 96 Subcooling effect in warm climates (TGC,out=40ºC) Figure 97 2 stages heat recovery for different Pinch Point Figure 98 High Temperature de-superheater (3 ºC pinch point) Figure 99 High Temperature de-superheater options (3 ºC pinch point) Figure 100 Low Temperature de-superheater (3 ºC pinch point) Figure 101 High Temperature de-superheater options (3 ºC pinch point) Figure 102 Gimo Energy Data, season Figure 103 Gimo Energy Data before and after renovation Figure 104 Hallevi ice arena Figure 105 Connected water loops Figure 106 Umeå ice arena

11 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 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) There is no significant difference in the overall heating energy demand between winter, where the ambient temperature is low, and summer. So, heat recovery is useful during the whole operation term. When the refrigerating load is around 100 kw the heating demand is around 150 kw. Nowadays, the heating systems(ventilation) are more efficient, so the heating demand could be assumed around 130kW when the refrigerating load is around 100 kw. The high temperature heating demand (domestic hot water, dehumidification) is 1/3 of the total energy demand. 11

12 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 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? as well. 12

13 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 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; ) Outdoor Ice rink Under suitable weather conditions, rink is constructed outdoor. But the limitation for outdoor rink is that it cannot be in operation during warm outdoor conditions. Outdoor rinks can only be operated during winter season. 13

14 Figure 3 Europe s largest outdoor ice skating rink in City Park (PBase; ) 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; ) 2.2. Ice rink classification More than 350 indoor ice arenas are operated in Sweden (Countries ranked by number of ice hockey rinks, 2018). These arenas are classified according to their size, purpose 14

15 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 There are many energy sub-systems in an ice rink. The main energy user of these subsystems is the refrigerating system. In second place is the heating that it is needed for space heating (spectators and offices) and domestic hot water. Dehumidification and ventilation consume an important part of the overall energy consumption. 15

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

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

18 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 non-toxic, non-flammable and nonexplosive 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. Figure 9 Refrigerants saturation pressure vs saturation temperature (Sawalha2008) 18

19 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 In the trans-critical area, the cycle can have different cooling capacity for the same condensing Temperature. The main factor that affect the cooling capacity is the pressure. 19

20 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 Heat recovery in CO2 systems A heat exchanger, which is called de-superheater, could be used before condenser to recover energy which would be rejected to the ambient. Using just a heat exchanger, the system may not work properly because of many parameters which should be counted. 20

21 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 Gimo s heat recovery layout gives the idea of a proper heat recovery. The system could be more complicated having two stages of heat recovery. 21

22 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 It could be used 2 de-superheaters, which are connected in series instead of one desuperheater after the compressor. This makes the system more efficient, because more energy is recovered by achieving higher supply water temperature. 22

23 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 The quality of recovered energy varies in a refrigeration cycle. Using CO2 as refrigerant in trans-critical operation the quality becomes higher. The refrigerant can achieve high temperature and it is in gas phase. This gives a more efficient heat transfer in the heat 23

24 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. 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 = m (h discharge Temperature h 20 ºC ) But the recovered energy depends on the refrigerant s temperature after the desuperheater. Q recovered = m (h discharge Temperature h after 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 desuperheater temperature. Recoverd Energy ratio = Q recovered Q recovered,total 24

25 Desuperheater outlet Temperature CO2 88 bar % 20% 40% 60% 80% 100% Recovered Energy ratio 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 desuperheater. 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 25

26 Desuperheater outlet Temperature Desuperheater outlet Temperature 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 CO2 88 bar 0 0% 20% 40% 60% 80% 100% Recovered Energy ratio 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 water side 0 0% 20% 40% 60% 80% 100% Recovered Energy ratio 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 26

27 Desuperheater outlet Temperature Desuperheater outlet Temperature of water s two-phase conditions, but, the water does not exceed this temperature, in reality water side 0 0% 20% 40% 60% 80% 100% Recovered Energy ratio 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. CO2 88 bar % 20% 40% 60% 80% 100% Recovered Energy ratio 88 bar water side Figure 21 Unrealistic CO2 and water Temperature profile 27

28 Desuperheater outlet Temperature 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. CO2 88 bar % 20% 40% 60% 80% 100% Recovered Energy ratio 88 bar water side 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 desuperheater, 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. 28

29 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. So, it is needed a solution in which the pinch point is small in order to be achieved optimum heat transfer and the same time the approach temperature not to be so high. A solution in this problem is by splitting the transferable energy in two different heat exchangers. One which operates in high temperature and other which operates in lower temperature. In this way the temperature range is divided in smaller parts and the return temperature of each secondary flow could have even higher value. 29

30 Desuperheater outlet Temperature 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. CO2 88 bar- 2 stage heat recovery nd stage 1st stage 0% 20% 40% 60% 80% 100% Recovered Energy ratio 88 bar water side 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. 30

31 Desuperheater outlet Temperature Desuperheater outlet Temperature CO2 88 bar- 2 stage heat recovery 2nd stage 1st stage 0% 20% 40% 60% 80% 100% Recovered Energy ratio 88 bar water side 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 CO2 temperature profile in desuperheater bar 90 bar 80 bar 75 bar 0 0% 20% 40% 60% 80% 100% Recovered Energy ratio Figure 27 CO2 profile for different Head Pressure 31

32 Density(kg/m3) Cp (kj/kg K) 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 bar 75 bar 74 bar 73 bar CO2 Temperature (ºC) 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 bar 80 bar 75 bar 74 bar 73 bar CO2 Temperature (ºC) 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

33 Desuperheater outlet Temperature Desuperheater outlet Temperature CO2-water temperature profile in desuperheater % 20% 40% 60% 80% 100% Recovered Energy ratio 90 bar 75 bar Water side 90 bar Head Pressure Water side 75 bar Head Pressure 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. CO2-water temperature profile in desuperheater >10ºC 20 17% % 20% 40% 60% 80% 100% Recovered Energy ratio 75 bar water side 75 bar Head Pressure 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. 33

34 Desuperheater outlet Temperature 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. 120 CO2 temperature profile 90 bar % IHE use 56% IHE use 100% IHE use % 20% 40% 60% 80% 100% Recovered Energy ratio 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. 34

35 Desuperheater outlet Temperature CO2-water temperature profile 90 bar % IHE use 100% IHE use water side 100% IHE use water side 0% IHE use 0 0% 20% 40% 60% 80% 100% Recovered Energy ratio 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. 35

36 Desuperheater outlet Temperature CO2 temperature profile 90 bar % IHE use 100% IHE use water side 100% IHE use water side 0% IHE use 0 0% 20% 40% 60% 80% 100% Recovered Energy ratio 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 = m (h 2 h s ) The consequence is that the heat which could be recovered is lower as well, because of lower mass flow. 36

37 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) HRR = Q recovered Q cooling 37

38 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) When the need of heat recovery is low the subcooling can provide a better performance in the system but when more heat recovery is needed the subcooler doesn t affect the system. An important conclusion is that subcooling can be provided in case that the 38

39 COPc heating recovery is not needed and becomes less when the heating recovery demand is higher till the subcooling become zero HRR COP Figure 39 COP change for different HRR In this case the COP remains in a very good level for higher values of HRR. 39

40 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 The system changed through season for being more efficient using better ventilation units and the same time energy from the recovery part was provided to a close swimming pool for heating purposes. So, in this study case, it will be analysed the difference of the system s operation. 40

41 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 To analyse the performance of the ice rinks refrigeration system an internal method is used. This method is referred to as the ClimaCheck method. In the internal method the compressor is used as a mass flow meter and therefore there is no need installing an 41

42 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. Where: m : Refrigerant mass flow rate η el : Electric motor efficiency m = η el P el Q comptressor loss h comp,out h comp,in P el : Electric power to the compressor motors Q comptressor loss : Heat loss from compressor body which is estimated 7% of the Electric power consumption h comp,out : Enthalpy after compressor h comp,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 , where the system had some operation and equipment changes. 42 Figure 42 Temperature after de-superheater, season in Gimo The temperature after the renovation became 35 ºC, and in the rest season after February the temperature became 20 ºC.

43 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 CarbonDioxide P [kpa] C 38 C C C C C x h [kj/kg] 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. Q recovered = m (h 1k h desuper,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 = Q recovered E extra Where E extra is the extra energy consumption in the compressor in order to be achieved the certain Q recovered. 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 E extra 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). 43

44 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 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 in Gimo So, the enthalpy before the evaporator is higher, but the same time, the cooling demand is the same. Having lower difference of enthalpy in the evaporator the mass flow should 44

45 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 in Gimo Figure 47 mass flow, season in Gimo 4.4 Discussion of Results The system can recover higher amount of energy without any mass flow changing, when the water return temperature is lower. This means that the system could cover the total heating demand without any changes in the refrigerating cycle operation. So, the energy consumption remains the same, while the recovered energy becomes higher. 45

46 Figure 48 COP_HR, season 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. Figure 49 Ventilation capaciry, season in Gimo 46

47 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 in Gimo This has as a result a lower arena s indoor temperature. So, we have an arena which starts freezing. The reason is that less indoor temperature drives to less energy absorption in the ice. So, the cooling demand is becoming lower. The cold sink (ice rink) is saturated, and the system will start being inefficient. In other words, the energy is rejected out of the system using the gas cooler instead of being provided in the cold sink (which is the arena) via the ventilation system. The gas cooler produces subcooling which gives better performance, but it destroys the potential of the energy recirculation inside the arena s closed energy system. 47

48 Figure 51 Indoor temperature, season 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. 48

49 Figure 52 volumetric flow in water loop, season in Gimo 49 Figure 53 Supply and return Temperatures in water loop, season in Gimo The system used to recover 100% HRR, which means that the recovered energy is the same as the cooling load. An ice arena is designed according to specifications. For the certain arena the specification is that the cooling load is 130 kw and the same time the heating demand is 130 kw. When the cooling load becomes 90 kw the heating demand

50 remains 130 kw. So, the heat recovery sector should operate in higher HRR. In this case the HRR should be 140%. The system will find again its steady conditions and the HRR is the heat recovery sector will be again 100%. 50

51 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 desuperheater 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 desuperheaters 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 = E compressor,is η total E compressor,is = m (h 1k,is h suction ) E compressor = m (h 1k h suction ) The discharge enthalpy (h 1k ) is calculated from the data, as well as the suction enthalpy (h suction ). The isentropic enthalpy is calculated using the discharge pressure and the 51

52 total efficiency (%) 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 = P ratio P ratio P ratio Dorin CD350H Pressure ratio Evaporator Figure 54 Dorin CD 350H 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. 52

53 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 T CO2[i], T water[i] Q[i] 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 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 desuperheater characteristics, it is assumed that in the most demanding cases the desuperheater 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. 53

54 Temperature (ºC) Pressure (bar) COPc 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 stage heat recoverey (20-60) HRR(%) Discharge Pressure COPc Figure 56 Discharge pressure and COPc under optimum efficiency control with respect to HRR CO2 Temperature (20-60) HRR(%) Discharge Temperature Temperature after Desuperheater Figure 57 Discharge and after de-superheater temperatures 54

55 Pressure (bar) COPc 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 ºC water temperature. The difference is that the system operates in higher discharge pressure stage heat recoverey (20-70) HRR(%) Discharge Pressure COPc 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. 55

56 Temperature (ºC) 140 CO2 Temperature (20-70) Discharge Temperature Temperature after Desuperheater HRR(%) 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 T CO2[i] rejected Q[i] Figure 60 de-superheater temperature profiles,discharge pressure lower than 80 bar 56

57 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. T CO2[i] Q[i] 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 desuperheater. 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. 57

58 Energy (kw) Gas cooler capacity Rejected recoverable energy less subcooling HRR(%) 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. 58

59 Temperature (ºC) Pressure (bar) COPc Results stage heat recoverey (20-40),(40-70) HRR(%) Discharge Pressure COPc 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 desuperheaters solution brings the water profile temperature closer to CO2 profile temperature CO2 Temperature (20-40),(40-70) HRR(%) Discharge Temperature Temperature after 1st Desuperheater Temperature after 2nd Desuperheater Temperature after Gas Cooler Figure 64 Discharge and after de-superheater temperatures 59

60 As it is illustrated in Figure 64 the temperature difference in the inlet and the outlet of the 1 st de-superheater is lower than the same temperature difference in the 2 nd desuperheater, when the HRR is around 80%. This is because of the CO2 profile temperature in the pressure of 79 bar. 80 Twater, TCO Q[i] Figure 65 1 st de-superheater 79 bar discharge pressure 60

61 60 T water, T CO rejected Q[i] Figure 66 2 nd 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. 61

62 80 Twater, TCO Q[i] Figure 67 1 st de-superheater 81 bar discharge pressure 60 T water, T CO Q[i] Figure 68 2 nd de-superheater 81 bar discharge pressure 62

63 COPc Discharge Pressure (bar) 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 stage (20-70) 2 stage (20-40),(40-70) HRR (%) 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 Stage (20-70) 2 stage (20-40), (40-70) HRR (%) 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 a better perspective. COP HR = Q recovered (E consumption E consumption,floating condenser ), 63

64 COPGl COPHR E consumption,floating condenser is the compressor s consumption in the same conditions without heat recovery stage (20-70) 2 stage (20-40), (40-70) HRR (%) 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 = Q recovered + Q cooling E consumption HRR (%) Figure 72 Global COP comparison 1 stage (20-70) 2 stage (20-40), (40-70) 1 stage (20-60) The reference case of Gimo has a very good COPGL but in case that we demand even higher water supply temperature this coefficient is more or less 0.5 less. For a system 64

65 Pressure (bar) Temperature (ºC) which operates more than 8 months this difference could have a big impact in the total energy consumption. But having 2 stages heat recovery the COPGl becomes 1.1 higher, which means that the performance is even better than the Gimo case. So, higher quality energy is supplied for heating with better performance, in case of 2 stages heat recovery. Comparison between theoretical data and real data (Gimo)- Proper controlling The operation control is very important according to Gimo operation for the season Having a very high COP is not enough for a good operation. The temperature after the subcooler should not be as low as it could be stage heat recoverey (20-60) HRR(%) Discharge Pressure Temperature after Gas cooler Figure 73 Control of operation according Heat recovery demand According to evaluation using EES the temperature after the subcooler in Gimo should be 18 ºC instead of 10 ºC which has a result lower mass flow and less heat recovery. Having 10 ºC the HRR is less than 120%, but in order to be covered any heating demand on the HRR should be around 130%. Warm climates Finally, the warm climate is not a constraint for CO2 refrigerating ice rinks. The ice rink could become efficient even in high ambient temperature by using subcooling. In the end of the day, even if the refrigerating COP for CO2 is a bit lower than the use of a regular Chlorofluorocarbon the advantages are more. It can cover the heating demands using the same equipment, when in case of Chlorofluorocarbons, it may be needed auxiliary heating. A very important parameter is the environmental impact of each refrigerant, CO2 as refrigerant has a very small footprint in the environment. 65

66 Discharge Pressure (bar) Pressure (bar) Temperature (ºC) stage heat recoverey (20-40),(40-70) HRR(%) Discharge Pressure Temperature after Gas cooler Figure 74 2 stage optimum control With Internal Heat Exchanger (IHE) An internal heat exchanger in any scenario could provide a safer operation. Apart from the benefits that an IHE can provide in the refrigerating cycle and the compressor, the discharge pressure is lower. In the operation conditions the discharge temperature is at least 2 bar less. This means a safer operation for the system, even if the COPGL of the system is almost the same stage heat recoverey (20-60) HRR(%) without IHE with IHE Figure 75 Optimum operation with and without IHE, 1 stage (20-60) 66

67 Discharge Pressure (bar) COPGl 1 stage heat recoverey (20-60) without IHE with IHE HRR(%) Figure 76 Global COP with and without IHE, 1 stage (20-60) 1 stage heat recoverey (20-70) without IHE with IHE HRR(%) Figure 77 Optimum operation with and without IHE, 1 stage (20-70) 67

68 COPGl Discharge Pressure (bar) COPGl 1 stage heat recoverey (20-70) without IHE with IHE HRR(%) Figure 78 Global COP with and without IHE, 1 stage (20-70) stage heat recovery(20-40),(40-70) HRR(%) without IHE with IHE Figure 79 Optimum operation with and without IHE, 2 stages (20-40,40-70) stage heat recovery(20-40),(40-70) HRR(%) without IHE with IHE Figure 80 Global COP with and without IHE, 2 stages (20-40,40-70) 68

69 In the conclusion the IHE does not provide a better performance but provide a safer solution without any extra operation cost. Judging the benefits and the disadvantages of an IHE in the certain case, this is a profitable solution. 5.4 Real heat exchangers A temperature difference of 2 K could be acceptable, but the question is how close is in reality? There are heat exchangers which can achieve this high efficiency but there UA value could be very high. This means very large heat exchanger which drives to non-applicable solutions because of big space or very expensive solutions. So it may be acceptable a solution with a less efficient heat exchanger. CAS Alfa Laval package was used for a heat exchanger evaluation. In every scenario, the type of heat exchanger will be checked in conditions which give 130% HRR. It is a good condition for ice rinks in order the heat demand to be covered. CAS gives the opportunity for a better heat exchanger evaluation through the margin option. For instance, if the margin is 15% which is an acceptable value, the software will give solutions for 15% higher load. One stage heat recovery (20ºC-60ºC) The recovered energy is kw while the discharge pressure is 89 bar in the condition of 130% HRR. The discharge temperature is around 95ºC. Figure 81 de-superheater conditions 69

70 Figure 82 Temperature profiles from CAS According to CAS software, there are heat exchangers which can achieve these conditions as it is shown in Figure 83. It could be used one large heat exchanger or two in parallel that have the half size of it. Figure 83 de-superheater options The AXP112 heat exchanger has UA=2110 W/m2K and total area 15 m 2. But the other two heat exchanger have a bit better UA value (UA=2274 W/m2K) and the half size(7 m 2 ). Both of these cases are acceptable for these conditions, but the 2 heat exchangers are more profitable for higher energy transfer according to margin value. One stage heat recovery (20ºC -70ºC) The recovered energy is kw while the discharge pressure is 96 bar in the condition of 130% HRR. The discharge temperature is around 102ºC. The discharge temperature should be higher as to be achieved the higher water supply temperature. This means higher discharge pressure. 70

71 Figure 84 de-superheater conditions Figure 85 Temperature profiles from CAS Figure 86 de-superheater options As it visible in Figure 86, the heat exchangers should be larger in that case. The same style of heat exchanger (AXP112) should be have more plates as to achieve more heat transfer area (17 m 2 ) giving a smaller UA value (UA=1972 W/m2K). Another solution is the use of three heat exchanger in parallel that they have UA=1868 and heat transfer total area 6.2 m 2. 71

72 Two stages heat recovery (20ºC-40ºC) and (40ºC-70ºC). The conditions of a high temperature heat exchanger and a low temperature heat exchanger are shown in Figures 87 and 90, respectively. High Temperature De-superheater Figure 87 1 st Desuperheaater conditions Figure 88 1 st de-superheater temperature profiles This heat exchanger should not be that efficient because the temperature difference in the pinch point could be higher than 2 ºC, as to be achieved the desired effect. 72

73 Figure 89 1 st de-superheater options Indeed, AXP112 with only 54 plates could be used, which have 1514 W/m2K UA value and 6 m 2 total heat transfer area. There could be used two AXP27 heat exchangers which have 1305 W/m2K UA value and 3.25 m 2 total heat transfer area. Low Temperature De-superheater Figure 90 2 nd de-superheater conditions 73

74 Figure 91 2 nd de-superheater temperature profiles Even the lower temperature heat exchanger should not be that tight except for the outlet. This gives a very efficient solution as it is shown in Figure 91. One AXP112 heat exchanger could be used for low temperature de-superheater. It has 3071 W/m2K UA value and total heat transfer area 9.6 m 2. Figure 92 2 nd de-superheater options 5.5 Warm climate scenario CO2 looks an inefficient solution for warm climates because of each properties. When the outdoor temperature is around 35 ºC, the CO2 temperature after the gas cooler is expected to be 40 ºC, assuming an approach temperature of 5 ºC in the gas cooler. This drives directly to a trans-critical operation, because the temperature after the gas cooler is higher than 31.2 ºC (critical temperature). 74

75 COP Warm climate CO2 operation Pressure Figure 93 Warm climate operation For a pressure of 95 bar the COP becomes close to 1.8, but even in this case is very low. The use of a subcooler is needed for the cycle becoming more efficient. The subcooler could provide heating for low temperature demands. So, the one of the two desuperheaters is switched off and the subcooler is switched on, when the ambient temperature becomes higher. Figure 94 Refrigerating cycle layout-subcooler Ventilation, ice protection or water preheating are some applications which could be used. Moreover, after every resurfacing a high amount of ice is wasted. We could take advantage of this ice as to make the refrigerating performance higher. 75

76 subcooling (kw) COP Figure 95 Subcooler-water loops The subcooling could provide a high profit in a CO2 refrigerating system under these conditions. The effect is illustrated in the figure 96. Subcooling effect Pressure (bar) Q_sub 10 Q_sub 20 Q_sub 30 COP 10 COP 20 COP 30 Figure 96 Subcooling effect in warm climates (TGC,out=40ºC) Every curve in Figure 96 illustrates how much subcooling is needed for a certain temperature after the subcooler (10,20 or 30ºC) and a different discharge pressure. The same time, it is illustrated the COP in each condition. As it is visible the cooling demand is higher in lower discharge pressures because of the CO2 isothermal curves profile. Subcooling strategy The main problem is that the load in the subcooler cannot be constant. The resurfacing demand is around 10 times per day, which means that the need of hot water is 10 times per day plus the demand of showers which could be 2 more times. It is assumed that 76

77 the subcooler provides energy to the ventilation all day and to preheater when it is needed. It is assumed that the secondary flow in the subcooler is a closed water loop which at first provides heating to the ventilation. Afterwards, the loop is divided in two flows. The one flow pass through the melting pit having a temperature drop of 10 ºC (25ºC- 15ºC) and the other is connected with the preheater, which drives to a temperature drop of 8 ºC (25ºC-17ºC), and the ice protection which drives to a temperature drop of 2-3ºC (17ºC-15ºC). Assumptions 77 Ventilation demand is around 80 kw and it operates in temperatures between 35ºC and 25 ºC. According to Figure 49 the ventilation demand in Gimo is around kw in properly operation. Because of warmer climate and considering less losses to the outdoor the ventilation energy demand could be assumed around 80 kw. Ice protection demand is around 7 kw. According to Mazyar the ice protection demand for Sweden is around 7 kw. (Mazyar Karampour) The water preheating takes 20 minutes. It is assumed a tank of 500 liters and the temperature of water becomes from 15ºC to 25ºC. In regular ice rinks, new ice could be provided every 45 minutes, in the melting pit. After every resurfacing the wasted ice is kg, which drives in an assumption of 500 kg ice per resurfacing. (2014 ASHRAE Handbook-- Refrigeration (I-P), 2014) According to the assumption, the melting pit provides 46.5 kw cooling load for 45 minutes, while the preheating needs 34 kw for 20 minutes. Four different cases will try to cover all the different operations according to what is the cooling which could be provided. 1 st Case In this case the subcooler provides heat only for ventilation. The temperature after the ventilation is around 25 ºC and as a logical assumption the temperature after the subcooler could be 28 ºC. The cooling load in this case is 80 kw According to the evaluation in the figure 96 the discharge pressure in the refrigerating cycle could be 88 bar having a COP of 2.6 which is 0.8 higher (62% higher) than having no subcooling. 2 nd Case In this case the subcooler provides heating for ventilation, freeze protection and preheating. It is assumed that is an operation of just 20 minutes but it works in these conditions 12 times per day. So, it is assumed that the temperature after the subcooler is 18ºC, while the cooling load is 119 kw. According to the evaluation in the figure 96 the discharge pressure in the refrigerating cycle could be 78 bar, having a COP of 2.9 (81% higher than having no subcooling). 3 rd Case In this case the subcooler provides heating for ventilation and in the melting pit. It is assumed that is an operation of 45 minutes, but it works in these conditions 10 times

78 Pressure (bar) per day. So, it is assumed that the temperature after the subcooler is 16ºC, while the cooling load is 126 kw. According to the evaluation in the figure 96 the discharge pressure in the refrigerating cycle could be 75 bar, having a COP of 3.1 (94% higher than having no subcooling). 4 th Case In this case the subcooler provides heating for ventilation, freeze protection, preheating and in the melting pit. It is assumed that it is an operation of 20 minutes, but it could operate in these conditions 10 times per day. So, it is assumed that the temperature after the subcooler is 14ºC, while the cooling load is 140 kw. According to the evaluation in the figure 96 the discharge pressure in the refrigerating cycle could be 75 bar, having a COP of 3.3 (106 % higher than having no subcooling). 5.6 Accuracy of Pinch Point The scenario of 2 stage heat recovery was used as a reference scenario in order to be investigated if the 2 ºC of pinch point is worthy. According to alfa laval the pinch point could be 1 ºC. (Christesensen, R. 2014) 2 stages heat recovery (20-40, 40-70) HRR(%) Pressure DT=1 Pressure DT=2 Pressure DT=3 Pressure DT=4 COP DT=1 COP DT=2 COP DT=3 COP DT=4 Figure 97 2 stages heat recovery for different Pinch Point Having 1 ºC temperature difference in the pinch point gives the best solution from discharge pressure perspective. But this does not mean that the system operates in high efficiency. The COP difference in case of 1 ºC and 2 ºC difference is almost the same. This makes the 2 ºC pinch point temperature difference more acceptable solution. In order to be achieved this 1 ºC difference the heat exchanger should have better performance which means bigger size and better materials. 78

79 Evaluation of higher pinch point temperature difference When the pinch point temperature is higher than 3 ºC, the de-superheaters could be smaller, according to CAS alfa laval software. The high temperature de-superheater has the properties, which are visible in the figure bellow. Figure 98 High Temperature de-superheater (3 ºC pinch point) Figure 99 High Temperature de-superheater options (3 ºC pinch point) The AXP52 heat exchanger has 1380 W/m2K UA value and 5.4 m 2 total heat transfer area, which is 0.6 m 2 smaller than the one which could be used in the case of 2 ºC in the pinch point. Low temperature de-superheater s characteristics are shown in the figure bellow. 79

80 Figure 100 Low Temperature de-superheater (3 ºC pinch point) Figure 101 High Temperature de-superheater options (3 ºC pinch point) The AXP52 heat exchanger has 2913 W/m2K UA value and 7.6 m 2 total heat transfer area. which is 2 m 2 smaller than the one which could be used in the case of 2 ºC in the pinch point. As a conclusion the 3 ºC pinch point solution gives smaller de-superheaters options. The de-superheaters are 2.6 m 2 smaller in total than in the 2 ºC pinch point solution. This gives a cheaper and more compact solution. On the other hand, higher temperature difference in the pinch point options give much lower COP. The COP drop in 130% HRR is almost 0.3, when this temperature different, becomes 3 ºC, and the same time the COP becomes much worse if the HRR is a bit higher according to figure 97. Judging the benefits of each solution the 2 ºC solution could be more acceptable, even if the investment cost is lower. 80

81 6 Statistics of existing Ice Arenas 6.1 Gimo Gimo as it has already been mentioned is the average ice ring which recovers energy in one stage. Even in this case a proper operation will give a better performance Season Cooling (MWh) Recovered (MWh) Consumption (MWh) Figure 102 Gimo Energy Data, season The energy consumption was decreased 50% after the renovation. This gives a very fast payback period. Gimo used to need district heating as auxiliary energy, but after the renovation is able to provide energy to external systems (swimming pool), which makes it even more profitable. Figure 103 Gimo Energy Data before and after renovation 81