MEASUREMENT AND MODELLING OF ICE RINK HEAT LOADS

MEASUREMENT AND MODELLING OF ICE RINK HEAT LOADS

Mazyar Karampour

Master of Science Thesis KTH School of Industrial Engineering and Management Energy Technology Department EGI-2011-094MSC Division of Applied Thermodynamics and Refrigeration SE-100 44 STOCKHOLM

Measurement and Modelling of Ice Rinks Heat Loads

Master of Science Thesis EGI 2011/ETT:094MSC Measurement and modelling of ice rink heat loads Mazyar Karampour Approved

Examiner

Supervisor

Date

Björn Palm

Joachim Claesson

Commissioner

Contact person

Master student:

Mazyar Karampour Forskarbacken 19/1508 11415 Stockholm

Registration Number:

801012-6558

Department

Energy Technology

Degree program

Sustainable Energy Engineering

Examiner at EGI:

Prof. Dr. Björn Palm

Supervisor at EGI:

Dr. Joachim Claesson

Supervisor at Industry:

Eng. Lic. Jörgen Rogstam

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ABSTRACT Ice rinks are among the most energy intensive public buildings in developed and developing countries. According to a research on Swedish ice rinks; a typical ice rink consumes approximately 1185 MWh/year which leads to more than 300 GWh/year for the 342 Swedish indoor ice rinks. The refrigeration system is usually the largest consumer by 43% average share of the total energy consumption. To decrease the refrigeration system energy demand, there are a variety of energy efficiency techniques known and available but the key to select the best ones is finding the major heat loads on the ice sheet and refrigeration system, which is unique for each ice rink. To fulfil this objective and in addition to review literature, this study has two main approaches. The first approach is to measure and evaluate the performance of the refrigeration system in two ice rinks, called Norrtälje and Älta. The estimated cooling capacity is approximately equal to the total heat load on the ice plus the heat gains in the distribution system. This goal has been accomplished by using a performance analyser called “ClimaCheck” which is based on an “internal method” because it uses the compressor as an internal mass flow meter and consequently, there is no need for an external one. The refrigerant mass flow rate is calculated by an energy balance over the compressor. By knowing the mass flow, enthalpy of the refrigerant, etc. the cooling capacity and COP of the system can be calculated. While the total heat load is known by the first approach, the second approach tries to discover different heat loads shares by analytical modelling. The measured physical and thermodynamical parameters plus the ice rink geometrical characteristics are input to the heat transfer correlations to estimate the heat load magnitude. The results of the measurements show that the total energy consumption in Norrtälje is about two third of Älta. The main reasons for this less energy consumption are smarter control systems for compressors and pumps, better ventilation distribution design and 1°C-2°C higher ice temperature. Analytical modelling for a sample day has estimated that about 84% of the total heat loads is originated from the heat loads on ice sheet while the distribution system causes the remaining 16%. Moreover, calculations show that convection plus small portion of condensation (altogether 36%), radiation (23%), ice resurfacing (14%) and lighting (7%) are the largest heat loads in winter while in summer condensation is another significant heat load (10%). Comparing two six-hour periods, one without ice resurfacing and four resurfacings in the second one, 30% more cooling demand has been calculated for the second period. Furthermore, it has been shown that the evaporator to brine is the contributor for 66% of the heat transfer resistances from ice to evaporator while brine to bottom ice and bottom to top ice accounts for 27% and 7% respectively. To conclude, a parallel “performance analysis of the refrigeration system” and “heat loads estimation” proves to be a useful tool for adopting proper design and control for energy efficient operation. Key words: Ice Rink, Refrigeration, Heat Load, Power Consumption, Energy Efficiency, Modelling, Measurement 3


ACKNOWLEDGMENT I like to express my deep appreciation and respect to Jörgen Rogstam, for his kind support, valuable lessons and never getting tired of my endless questions. Special thanks to my supervisor at KTH, Dr. Joachim Claesson and Kenneth Weber at ETM Kylteknik AB for their helpful comments and discussions. Kenneth was really a generous person in sharing the valuable experiences with me. Thanks to Swedish Energy Agency for financing this research as part of the Stoppsladd project. I am grateful to Jakob Månberg (ClimaCheck), Ari Penttilä (Prorink), Antoni Gosalvez (MayekawaMYCOM), John Ekwall (Swedish Meteorological and Hydrological Institute), Torbjörn Thoresson (REFCALC) and Pavel Makhnatch for their help to provide me some required data. I should thank Matthias Dahlberg for making the company Energi & Kylanalys a pleasant and friendly atmosphere to work.

The last thanks to my family, for their lifelong support.

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CONTENTS 1.

Introduction ............................................................................................................................................... 9 1.1 Objectives...............................................................................................................................................10 1.2 Methodology ..........................................................................................................................................10 1.3 Scope and limitations............................................................................................................................11

2.

Ice rinks ....................................................................................................................................................12 2.1 Ice rink energy systems ........................................................................................................................12 2.2 Ice rink energy systems shares ............................................................................................................18 2.3 Ice rink refrigeration system ................................................................................................................19 2.3.1 Ice pad structure and piping arrangement .................................................................................21 2.4 Heat loads in ice rinks ..........................................................................................................................23 2.4.1 Convection-Condensation ...........................................................................................................23 2.4.2 Radiation .........................................................................................................................................23 2.4.3 Conduction .....................................................................................................................................23 2.5 Energy efficiency in ice rinks ..............................................................................................................25 2.5.1 Heat loads decrease .......................................................................................................................25 2.5.2 Refrigeration and distribution system performance improvement ........................................26 2.5.3 Ice/concrete slab quality enhancement......................................................................................27

3.

Experimental measurements .................................................................................................................28 3.1 Ice rinks ..................................................................................................................................................28 3.1.1 Norrtälje ice rink............................................................................................................................28 3.1.2 Älta ice rink ....................................................................................................................................30 3.2 Performance analyser - ClimaCheck ..................................................................................................31 3.2.1 Energy balance method ................................................................................................................32 3.2.2 ClimaCheck method modification ..............................................................................................34 3.3 Measurement results .............................................................................................................................36 3.3.1 “Cooling chain” temperatures .....................................................................................................36 3.3.2 Evaporation/Condensation temperatures .................................................................................37 3.3.3 Brine and coolant temperatures ..................................................................................................38 3.3.4 Air temperature and relative humidity over ice, indoor and outdoor....................................39 3.3.5 Electric power input and cooling capacity.................................................................................39 5

Measurement and Modelling of Ice Rinks Heat Loads 3.3.6 Total energy consumption ...........................................................................................................40 4.

Analytical modelling ...............................................................................................................................42 4.1 Heat loads...............................................................................................................................................42 4.1.1 Radiation .........................................................................................................................................43 4.1.2 Convection .....................................................................................................................................44 4.1.3 Condensation .................................................................................................................................45 4.1.4 Lighting ...........................................................................................................................................47 4.1.5 Ground Conduction .....................................................................................................................48 4.1.6 Brine headers..................................................................................................................................48 4.1.7 Ice resurfacing ................................................................................................................................49 4.1.8 Pump work .....................................................................................................................................50 4.1.9 Skaters .............................................................................................................................................50 4.1.10 Results for heat loads shares......................................................................................................51 4.2 Heat flux method ..................................................................................................................................53 4.2.1 Results .............................................................................................................................................54 4.3 Temperature resistances-differences ..................................................................................................55

5.

conclusion ................................................................................................................................................56

6.

Future work / suggestions.....................................................................................................................57 6.1

Future work ......................................................................................................................................57

6.2

Suggestions .......................................................................................................................................57

7.

Bibliograpgy .............................................................................................................................................58

8.

appendix ...................................................................................................................................................60 8.1 Compressor heat rejection sample calculations ................................................................................60 8.2 Cooling capacity sample calculations .................................................................................................63 8.3 Norrtälje and Älta ice rinks photo gallery .........................................................................................64

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FIGURES Figure 1: Refrigeration system energy vs. total purchased energy for ice rinks (Rogstam (c), 2011) .... 9 Figure 2: Energy systems in ice rink (Retscreen, 2005) ..............................................................................12 Figure 3: Configurations of heat rejection and heat recovery from a refrigeration system (Sawalha, 2010) ..................................................................................................................................................................13 Figure 4: A ventilation duct over the stands in an ice rink ........................................................................14 Figure 5: Dehumidification by water vapour condensation (IIHF, 2010)...............................................15 Figure 6: Dehumidification by Desiccant wheel (IIHF, 2010) ..................................................................16 Figure 7: An integrated ventilation and desiccant dehumidification system (Munters, 2011) ..............16 Figure 8: Energy systems consumption shares (Rogstam (a), 2010) ........................................................18 Figure 9: Refrigeration plant with recovery (IIHF, 2010) ..........................................................................19 Figure 10: Refrigeration components electricity consumption shares (Rogstam (b), 2010) .................20 Figure 11: Ice pad structure (How ice rink works?, 2011) .........................................................................21 Figure 12: Typical piping arrangement in a distribution network (Ingvar, 2007) ...................................22 Figure 13: Daily indoor ice rink heat loads (ASHRAE, 2010) ..................................................................24 Figure 14: Norrtälje Sportcentrum, a) Ice hockey hall, b) Bandy, c) Artificial soccer field, d) Indoor sport hall, e) Track and field pitch, f) Pool (to be constructed)...............................................................28 Figure 15: “Green soccer field” in “white winter” thanks to condenser waste heat .............................29 Figure 16: Norrtälje ice rink spectators stand (top-left), machinery room inside (top-right), compressors (bottom-left), and machinery room outside (bottom-right) ...............................................30 Figure 17: Älta ice rink hall and spectator stand (top-left), machinery room (top-right), flooded evaporator (bottom-left), and one of the two compressors (bottom-right)............................................31 Figure 18: ClimaCheck basic instrumentation configuration ....................................................................32 Figure 19: Energy balance over compressor ................................................................................................33 Figure 20: ClimaCheck flowchart for Norrtälje and Älta ice rinks ...........................................................34 Figure 21: Relative heat rejection versus RPM and tcond (teva=-10°C, superheat =7K, subcool = 5K)35 Figure 22: Average relative heat rejection versus tcond (teva=-10°C, superheat =7K, subcool = 5K) ...36 Figure 23: Ice, brine and evaporating refrigerant temperature fluctuations, 15 March 2011, Norrtälje ............................................................................................................................................................................37 Figure 24: Evaporation and condensation temperatures, 15 March 2011 ...............................................38 Figure 25: Brine and coolant temperatures, 15 March 2011 .....................................................................38 Figure 26: Air temperatures and relative humidity, 15 March 2011 .........................................................39 Figure 27: Electric power and cooling capacity, 15 March 2011 ..............................................................40 Figure 28: Total refrigeration system energy usage and outdoor temperatures in March 2011 ...........41 Figure 29: Heat loads in ice rinks and their impact points ........................................................................42 Figure 30: Angle factor between two aligned parallel faces (Çengel, 2007) ............................................44 Figure 31: Cooling capacity versus heat loads at midnight, 15 March 2011 ...........................................46 Figure 32: Air and ice temperatures and air relative humidity, 12 July 2010, Norrtälje.........................46 Figure 33: Condensation heat transfer coefficient, 12 July 2010, Norrtälje ............................................47 Figure 34: Lighting fixtures in Norrtälje .......................................................................................................48 Figure 35: Hourly average ice temperature, 15 March 2011, Norrtälje....................................................50 7

Measurement and Modelling of Ice Rinks Heat Loads Figure 36: Heat loads shares in the total heat load .....................................................................................52 Figure 37: Top and bottom temperature sensors embedded in the ice ...................................................53 Figure 38: Top and bottom heat transfer rate on days 14, 15 and 16 April 2011 ..................................54 Figure 39: Ice average temperature fluctuations on 14, 15 and 16 April 2011 .......................................54 Figure 40: Heat flow from ice to refrigeration plant ..................................................................................55 Figure 41: MYCOM software interface ........................................................................................................60 Figure 42: Cooling capacity calculation sample for 23:00-23:59 on March 15, 2011.............................63 Figure 43: Flooded evaporator (right) and brine pumps (left) - Älta .......................................................64 Figure 44: Coolant pumps and condenser (left corner) - Älta ..................................................................64 Figure 45: DX evaporator (right), desuperheater (top-left) and condenser (bottom-left) - Norrtälje 64 Figure 46: refrigeration system for outdoor bandy field - Norrtälje ........................................................64 Figure 47: heat recovery pump for ventilation - Älta .................................................................................64 Figure 48: District heating system - Norrtälje .............................................................................................64 Figure 49: desiccant wheel dehumidification and ventilation heat recovery unit - Älta ........................65 Figure 50: dehumidification piping (right) and heating/cooling coils (left) in ventilation ducts Norrtälje ............................................................................................................................................................65 Figure 51: ClimaCheck central control unit - Älta ......................................................................................65 Figure 52: Instruments for over ice temperature and humidity measurements- Älta ............................65 Figure 53: ventilation ducts and lighting - Älta ...........................................................................................65 Figure 54: ice resurfacing machine - Älta .....................................................................................................65

TABLES Table 1: Available lamps for ice rinks (IIHF, 2010) ...................................................................................17 Table 2: Ice pad and piping dimensions (ASHRAE, 2010)(IIHF, 2010) ................................................22 Table 3: Sample temperature control (Everything Ice, 2000) ...................................................................27 Table 4: Reasons for higher energy consumption in Älta ..........................................................................41 Table 5: Physical properties for brine headers for heat load calculation .................................................49 Table 6: Heat loads summary .........................................................................................................................51 Table 7: Daily heat load calculations, 15 March 2011 ................................................................................52 Table 8: Heat transfer resistances and temperature differences ...............................................................55 Table 9: Heat rejection calculation results by MYCOM software ............................................................61

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1. INTRODUCTION Imagination of a prosperous sustainable society without smart energy strategies seems impossible. While the amount of world energy consumption increases nonstop, there are only two solutions to avoid the fatal consequences: producing energy more sustainably and using the produced energy more efficiently. While there are big efforts to use renewable energies, it seems that still there is a long way to go and the dominating resources are still non-renewable energies. Keeping this in mind, using this energy more efficiently is the best answer to problems following by enormous energy consumption. Ice rinks are among the most energy consuming public areas which roots in simultaneous cooling, heating, ventilation and lighting demand. In small municipalities, ice rinks are the biggest energy consumers. Average annual energy consumed in a Swedish ice rink is 1185 MWh/year which this amount is supplied 82% by electricity and 18% by heat and the total energy consumption for Sweden ice rinks exceeds 300 GWh/year. The refrigeration system has the biggest share with 43%. (Rogstam (a), 2010) The relation between the refrigeration system energy usage versus the total purchased energy (including electricity and heat) is shown in Figure 1 for a number of ice rinks. It is expected that refrigeration energy consumption should increase with the total consumption increase but a considerable spread is seen. Very different operation and activities patterns and not considering the heat recovery potential in some ice rinks would be justifications for this spread. (Rogstam (c), 2011)

Refrigeration system energy usage (MWh/yr)

1200 1000 800 600 400 200 0 0

500

1 000 1 500 Total purchased energy ( MWh/yr)

2 000

Figure 1: Refrigeration system energy vs. total purchased energy for ice rinks (Rogstam (c), 2011)

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Measurement and Modelling of Ice Rinks Heat Loads As a comparison, Québec ice rinks average energy consumption is around 1 500 MWh/year, while the most efficient ones consumes 800 MWh/year and the least efficient ones 2 400 MWh/year (Nicolas, 2009). For the whole Canadian ice rinks the total electricity consumption is approximately 3 500 GWh/year. (Bellache, 2007) In Sweden there are about 341 indoor and 140 outdoor ice rinks, sized about 1800 m2 (60 m×30 m). Furthermore, 60 outdoor bandy rinks, around 8000 m2 each, exist in Sweden. Nine indoor bandy arenas are built during the recent years. The operating months for indoor ice rinks is 6-10 months, with an average of 8 months. For outdoor ice rinks, the winter period which lasts 3-5 months is the operating time. (Rogstam (a), 2010) While the total energy consumption by indoor ice rinks is more than 300 GWh/year, the ice rink numbers and the working periods is increasing continuously and it seems that Swedish people need more and more ice rinks, all year round. It means that the energy consumption will increase steadily if there are no policies adopted for better energy efficiency techniques. To find the best energy efficiency solutions, the first step is to know various heat loads and their shares on the load to the refrigeration system which is the largest energy consumer.

1.1 Objectives The objective of this study is to evaluate the heat loads in ice rinks. To obtain the best results the following steps are intended: a) Study literature on measurements and models on heat loads in ice rinks or similar applications. b) Evaluation of two ice rinks with ClimaCheck instrumentation enabling monitoring the cooling capacity/ice rink heat load. c) Build a simulation model with an appropriate tool for simulating the ice rink heat load in order to find the heat transfer mechanisms shares in the total heat load.

1.2 Methodology To fulfil the objectives of the research, three main steps are decided. The first step is to review the ice rink energy systems, the technology and different users of the input energy to the ice rinks, heat loads in the ice rinks and furthermore, to introduce the most promising energy efficient methods used to decrease the energy consumption in ice rinks. The second step is the experimental part; two ice rinks will be introduced. Moreover, a measurement system installed in these two ice rinks will be described. Finally, the results of the measurements will be presented. The most important output of theses measurement is to find the cooling capacity and, indirectly, total heat load which refrigeration system should compensate.

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Measurement and Modelling of Ice Rinks Heat Loads Trying to break the total heat load into various components will be discussed in the third step, analytical modelling. A correlation or estimate to calculate each of the major heat loads in the ice rinks will be presented and then the total heat loads will be compared with the provided cooling capacity.

1.3 Scope and limitations While the refrigeration system and heat loads of the indoor and outdoor ice rinks are to some extent similar, this research concentrates mainly on indoor ice rinks and in particular two indoor ice rinks in the Stockholm region. This makes the results and conclusions applicable the best for similar climate and built environment conditions, for example Scandinavian or North American locations above 50°-55° latitude. The interactions between the heat loads, as the driving forces, the ice, as the object of cooling, and the refrigeration system, as the responding/cooling system, is of main interest in this research. That is why other energy systems in ice rinks including the heating and ventilation systems are not discussed and analysed in any extend. The limitations of the research are few unknown parameters in the measurement and modelling process. Whenever such a parameter is encountered, it has been mentioned and the best possible logical assumption is made.

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2. ICE RINKS 2.1 Ice rink energy systems Ice rink energy system is comprised of several energy systems, indicated in Figure 2, because there are various demands in the ice rinks. What makes the ice rinks unique in comparison with other public buildings is the wide range of demands. For example, there is a permanent need for cooling and heating to provide temperatures ranging from -4°C (ice) to around +60°C (Domestic Hot Water) in the ice rinks, simultaneously and in a stable condition. There is a second difficulty as well; there are very few internal partitions to separate these energy systems targets. The energy systems that every ice rink should have are: refrigeration, heating, ventilation, dehumidification and lighting. The first three ones require distribution systems as well which are powered by pumps and fans for mass and energy transfer.

Figure 2: Energy systems in ice rink (Retscreen, 2005)

Refrigeration system is the most important energy system as it makes the ice and keeps it from melting. Considering the huge ice mass, for a typical ice rink the cooling capacity should be around 300-350 kW (IIHF, 2010). The most conventional refrigeration system used is electricity powered vapour compression indirect system. This system is explained more in section 2.3.

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Measurement and Modelling of Ice Rinks Heat Loads Heating system provides the required heat for space heating, ventilation, domestic hot water (DMW), ice resurfacing water, floor heating, snow melting, and subsoil heating. The heating system can be fed by fossil fuels, electricity or district heating but the most energy efficient, cost effective and environmentally friendly method is to use the heat rejected by the refrigeration system through the condenser and desuperheater (if available). The amount of heat pumped by the refrigeration system can cover a great share of the heating demand, sometimes even 100% of the need (ASHRAE, 2010). Using this integrated method is smart as the refrigeration system is used both in the evaporator side and the condenser side; hence it is not only a refrigeration system but also a heat pump. There are various ways available to exploit this waste heat. Some examples have been shown in Figure 3.

Figure 3: Configurations of heat rejection and heat recovery from a refrigeration system (Sawalha, 2010)

The middle layout is a refrigeration system without heat recovery and the refrigerant heat is rejected to atmosphere. It is called floating condensing as the condensation pressure follows the ambient temperature. In this case, all of the heating demands should be covered by a separate heating system including district heating, heat pump, etc. Top-left layout is a heat recovery system by a desuperheater. This system is suitable when the discharge temperature is relatively high. Refrigeration systems that use NH3 or CO2 can use this desuperheater heat recovery. The regulating valve after the condenser/gas cooler can adjust the discharge pressure and, consequently, the desuperheater heating capacity. Top-right and bottom-left figures are two heat pump cascade solutions. In the bottom-left layout heat is recovered from the condenser and delivered to a heat pump as the low grade heat. Then the heat pump upgrades it to higher temperatures for HVAC demands. This allows the refrigeration system to have lower discharge pressures. This system is called heat pump cascade. Top-right 13

Measurement and Modelling of Ice Rinks Heat Loads solution (heat pump cascade for subcooling) is similar to the heat pump cascade but the heat is recovered in a subcooler after the condenser. This increases the efficiency of the refrigeration system. The bottom-right system is a fixed-head pressure heat recovery system. In this solution the discharge pressure is adjusted according to the HVAC system demand. There is a coolant which transfers the heat from the condenser to the HVAC system. (Sawalha, 2010) Ventilation system delivers the fresh air to the inhabitants and provides the standard air change rate to avoid pollutant, smell, fog and biological disease sources concentration. During the design of the ventilation system, it can be divided into two zones; the ice rink and public areas. In the ice rink, spectators’ stand and emissions from ice resurfacing machine (if it is not an electric one) are to be considered. Moreover, direct air blown to the ice surface should be avoided. For public areas, air change required in the closed spaces including restaurant, offices, locker rooms for players, coaches, referees and linesmen, drying rooms, medical rooms and toilets/showers should be considered (IIHF, 2010).

Figure 4: A ventilation duct over the stands in an ice rink

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Measurement and Modelling of Ice Rinks Heat Loads Dehumidification system keeps the relative humidity of the indoor air up to a standard level. Too humid indoor air causes the corrosion of the metal and rotting of the wooden structures. Moreover, fungi and mould growth is more probable in a humid atmosphere. Another problem with too much water vapour content in the air is the fog created over the ice that makes it hard to play or control the movements. The last problem is the heat load on the ice due to condensation. To dehumidify the air, two primary solutions are available. The first one is to cool the humid air below its dew point. This leads to condensation of part of the air water content. For cooling the air, part of the cold brine can be used. Figure 5 illustrates this dehumidification process. Dehumidification by condensation can be integrated with ventilation or refrigeration system.

Figure 5: Dehumidification by water vapour condensation (IIHF, 2010)

The second method is to use water absorbing materials like silica gel. The most well-known equipment which uses this technique is called “desiccant wheel”. Desiccant wheel is the major component in a desiccant humidification system. It is a slow rotating wheel containing some absorbent chemicals (normally silica gel). When moist air passes one portion of the wheel, the moisture is absorbed. While it is rotating, in other portion of the wheel a drying air is blown to the wet absorbent to dry and “regenerate” it. In this system, the desiccant wheel plays a role of a “moisture transporter”; takes the moisture away from the supply air to the ice rink and transports it to the exhaust air. A simple desiccant dehumidification is shown in Figure 6.

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Figure 6: Dehumidification by Desiccant wheel (IIHF, 2010)

Desiccant dehumidification system can be integrated with an air handling unit of ventilation system. Figure 7 indicates an example of such a system. The return air from the ice rink (pink stream) is divided into two streams: one to be exhausted to the atmosphere and a portion is mixed with the fresh make-up air. The heat from the return exhaust air is recovered in an energy recovery wheel to preheat the make-up air. Then, a mixture of return and make-up air passes the desiccant wheel in the middle of the unit. The desiccant wheel removes some portions of moisture from this air. The desiccant wheel is reactivated (regenerated) by a hot air stream in the upper part of the wheel to be used as a moisture absorbent again in the lower part. The supply air after the desiccant wheel can be heated or cooled by heating/cooling coils. However, in ice rinks cooling coils are not used most of the year. The air enters the ice rink as the “supply air” from left side of the unit.

Figure 7: An integrated ventilation and desiccant dehumidification system (Munters, 2011)

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Measurement and Modelling of Ice Rinks Heat Loads Lighting is an energy system that provides a clear and pleasant indoor environment for the skaters and spectators as well. Different activities in an ice rink require different light intensities. Lighting intensity is normally measured in units of lumen (lux) or foot candle (FC). Each foot candle is equal to 10.76 lux. In general, figure or recreational skating requires10-15 foot-candles and curling 10-50 FC while ice hockey needs 80-150 foot candles. As a consequence, a method to decrease the energy consumption by the lighting system is to control the lighting intensity. (DOE, 1980)(Everything Ice, 2000) Another method to have an efficient and smart lighting system is to select energy-efficient lamps and lighting fixtures. Lamps can be categorized to incandescent and burst illuminates according to their operational principle. Generally, incandescent lamps are only suitable for general lighting, except for halogen lamps. Incandescent lamps consume relatively high electricity compared to the produced visible light. They have short life time but good controllability. Burst lamps, in contrast, have high efficiencies and long lifetime but poor controllability. Generally, Burst lights are more suitable for rink lighting (IIHF, 2010). Table 1 shows some of the more well-known lamps available for ice rinks. Luminous-efficacy of a lighting source is defined as the ratio of emitted visible light, in lux, to the total consumed electricity, in W. This parameter shows how much energy efficient the lighting device is. The luminous-efficacy for some incandescent lamps are 15-20% while some burst lamps including metal halide and florescent lamps can have 80-90% efficacy (Luminous, 2011). This means that, for example a 13 W compact florescent lamp can provides the same lighting of 800 lumens as a 60 W incandescent lamp (Everything Ice, 2000). Table 1: Available lamps for ice rinks (IIHF, 2010)

Type

Applicability

Power range

Life

Note

Compact fluorescent lamps

General lighting

5-55 W

8 000-12 000 h

Good energy efficiency

30-80 W

20 000 h

Good energy efficiency

Standard fluorescent lamps

General lighting

Metal halide lamps

Rink lighting

35-2000 W

6 000 - 20 000 h

Good for rink lighting

High pressure sodium lamps

Rink lighting

50-400 W

14 000 – 24 000 h

Poor colour rendering

Induction lamps

Rink lighting

55-165 W

60 000 h

Long life, expensive (so far)

Halogen lamps

Special lighting

20-2000 W

2 000 – 4 000 h

Rink lighting

Excellent colour rendering, good dimming capabilities

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2.2 Ice rink energy systems shares Through a statistical study of more than one hundred ice rinks in Sweden it is revealed that the refrigeration system has the largest share in total energy consumption, 43% (in average) as indicated in Figure 8. (Rogstam (a), 2010) Heating with 26% share is the second biggest energy consumer. In Sweden the main sources of heating are district heating and/or electricity plus the heat recovered from the refrigeration system’s high pressure side and the heat recovered from exhaust air in ventilation system. Lighting, ventilation system fans and dehumidification system are the next largest energy consumers.

1% 6%

5% Refrigeration

9%

Heating 43%

Lighting Ventilation Fans

10%

Dehumidification Miscellaneous Misc. Pumps 26%

Figure 8: Energy systems consumption shares (Rogstam (a), 2010)

There are several other researches confirmed that refrigeration system is the biggest energy consumer in majority of world ice rinks. In a research by CANMET, the research organization of Natural Resources Canada, refrigeration consumption is estimated to be 50% of the total energy consumption, by electricity or heat (AAQ, 2003). In a research by International Ice Hockey federation, refrigeration plant consumes 57% of the electricity input to a prototype ice rink in Munich, Germany (IIHF, 2010). A research published in ASHRAE journal estimates that while in an inefficient arena (1950 MWh/year consumption) refrigeration share is only 23%, in an efficient arena with heat recovery systems (840 MWh/year) refrigeration consumes about 42% of the total. (Nicholas, 2009)

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2.3 Ice rink refrigeration system The refrigeration system is known as the heart of the ice rink because it is the guard to keep the ice in its most desired form. A refrigeration system for the ice rink is direct, indirect, or a combination of them called partly indirect. In the direct system the refrigerant is pumped below the ice pad and the whole refrigerant distribution pipes serves as a large evaporator. This method is less used as there is a need for huge amount of refrigerant charge. R-22 and ammonia are the most used refrigerants for the direct systems but R-22 is banned now in many countries due to its global warming potential and ammonia has a charge limit according to its hazards and cannot be used in large systems including ice rink direct systems. Indirect system is the most conventional layout for ice rink refrigeration system. In this system a primary refrigerant cools a secondary refrigerant, known as “brine”, and then the distribution system circulates this secondary refrigerant below the ice pad and returns it back to evaporator. In Sweden, more than 97% of the ice rinks are indirect or partly indirect. (Makhnatch, 2010) Partly indirect are systems that either evaporator or condenser is connected to the source or sink by a secondary fluid for heat exchange. In partly indirect systems some portion of the cooling is provided by a direct system as well. (Melinder, 2009) A drawing of a typical ice rink with indirect system is demonstrated in Figure 9. As mentioned, refrigeration unit cools the brine in evaporator and the brine is sent to the embedded cooling pipes below the ice pad. The refrigeration system typically consists of a vapour compression cycle driven by electricity as the primary cycle. About 85% of the Sweden ice rinks use Ammonia as the refrigerant while the remaining use R404A, R134a or other HFC refrigerants. (Makhnatch, 2010)

Figure 9: Refrigeration plant with recovery (IIHF, 2010)

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Measurement and Modelling of Ice Rinks Heat Loads In the secondary loop there are one or more pumps that circulate the brine. Calcium Chloride (CaCl2) and Frezium are the most conventional brines in Sweden. In some ice rinks a small portion of the brine flow can be used in dehumidification units, as indicated in Figure 9. Compressors for ice rinks are traditionally reciprocating compressors while screw compressors are the other choices. Generally, more than one compressor is used for ice rinks and among the several compressors one of them can be selected to be ample capacity to adjust “cooling production” in harmony with fluctuating heat loads. In the ice making period or heavy activities on ice during high activity hours, the second compressor will be engaged to assist the first one. The rejected heat from condenser can be recovered to supply a number of heat demands in an ice rink including ventilation unit, floor heating, hot water storage (which is used for domestic hot water and/or ice resurfacing water) and ground frost protection. There are some other heat demands which are not shown in the sketch; for example, part of the heat can be used to melt the snow, produced during ice resurfacing, or a portion can be used for a nearby swimming pool. 75% to 100% of the space and water heating requirements can be supplied by smart heat recovery (ASHRAE, 2010). The excess unexploited heat is removed in outdoor cooling coils which are installed on ice rink roof, generally. The most conventional cooling units are dry air coolers and the most used coolants in Sweden are Glycol, Ethylene Glycol and Propylene Glycol (Makhnatch, 2010). Looking at the power consumption in refrigeration system as shown in Figure 10, it has been revealed that compressors account for 80% of the total electricity consumption while brine pumps, coolant pumps and dry cooler fans consume 10%, 5% and 5% respectively. (Rogstam (b), 2010)

5%

5%

10% Compressors Brine Pumps Coolant Pumps Dry Cooler Fans 80%

Figure 10: Refrigeration components electricity consumption shares (Rogstam (b), 2010)

20

Measurement and Modelling of Ice Rinks Heat Loads 2.3.1 Ice pad structure and piping arrangement Ice pad structure consists of a number of layers and each layer is designed in response to a requirement. As indicated in Figure 11, the topmost layer is ice. The most conventional structure for the second layer is concrete while in some ice rinks sand or asphalt is used as well. The distribution system of brine pipes is embedded in this “chilled concrete slab”. The level of concrete and pipes should be completely flat to have a well-distributed cooling and uniform ice thickness.

Figure 11: Ice pad structure (How ice rink works?, 2011)

Below the cold concrete layer there is an insulation pad to decrease the conduction heat gains from the ground. The next layer is “heated concrete”. It holds the weight of the top layers and it is heated to eliminate or minimize the hazards of ground heaving. Ground heaving which is contributed by soil freezing can lead to collapse of the whole ice pad structure. One of the most conventional methods to keep the concrete layer warm is to run warm water in a pipe system, similar to the brine pipes. The condenser heat recovery system can be the water heating source. All the layers are laid on a layer of sand and gravel. A ground water drain collects any water to avoid the water to be absorbed by the top layers. For the concrete and insulation layers being as dry as possible is a requirement. Figure 12 shows a typical piping arrangement for an ice rink. There are two brine pipe headers which one of them is the brine supply header and the other one is return header or collector. As it is shown, the brine distribution pipes are branched from the headers and have a U shape. This simple U shape is called two-pass arrangement but a four-pass arrangement is applied in some ice rinks as well. Four-pass layout has a W shape. Four-pass systems are claimed to be more energy efficient and decrease the energy consumption. (AAQ, 2003) As it can be seen in the figure, the brine header size is decreased gradually when the flow decreases as well. The reason is to have a uniform flow, and as a consequence, heat transfer distribution. Another method for uniform flow distribution is to use some small orifices in the inlet of the small brine distribution pipes. In this arrangement the brine header sizes remain constant.

21


Figure 12: Typical piping arrangement in a distribution network (Ingvar, 2007)

Dimensions of the ice pad structure and header and distribution pipes are indicated in Table 2. It can be seen the concrete thickness is 150 mm and the brine distribution pipes are located such that pipes top point is 25- 50 mm lower than ice bottom surface. Smaller the distance, less heat transfer resistance occurs. Furthermore, less concrete thickness leads to less load on the refrigeration system and there are some tries to decrease the 150 mm thickness to 125 mm. Header pipes are 6-8 inches (150-200 mm) typically and made of steel or Polyvinyl Chloride (PVC). Small brine distribution pipes are 25-32 mm and can be steel or polyethylene plastic pipes. There are some new materials including copper tubes which demonstrate a good heat transfer and flexibility properties (Shahzad, 2006). The distribution pipes are fixed in 100 mm spacing with some supports or spacers. Table 2: Ice pad and piping dimensions (ASHRAE, 2010)(IIHF, 2010)

parameter

size

Ice thickness

25-30 mm

Concrete thickness

150 mm

Insulation thickness

100 mm

Brine headers diameter

150-200 mm

Brine distribution pipes diameter

25-32 mm

Pipes spacing

75-125 mm (100 mm typically)

Top pipe – bottom ice distance

25-50 mm

22


2.4 Heat loads in ice rinks Heat loads in ice rinks can be categorized into three dominating heat transfer mechanisms:   

Convection - Condensation Radiation Conduction

Brief explanations of these will be given here but the detailed mathematical correlations used for calculations are discussed in chapter four, analytical modelling. 2.4.1 Convection-Condensation In convection heat transfer of air to the ice, air temperature, ice surface temperature and air velocity are important parameters. Higher temperature gradient between air and ice surface and higher air velocity lead to higher convective heat transfer, hence lower air velocity over ice and closer air and ice temperature are factors to decrease this heat load. Furthermore, water vapour in the humid air rejects its heat to the ice and condenses on the surface. This phenomenon is more serious in the ice rinks which are in operation during a humid summer climate. Dehumidification of the supply ventilation air is necessary to avoid condensation and keep the relative humidity low. Condensation could be a source of bad ice surface quality which brings the ice resurfacing requirement, hence has an indirect negative impact as well. 2.4.2 Radiation Two major sources of radiation are ceiling radiation and lighting. In several ice arenas radiation is reported having the largest share in heat loads on the ice (ASHRAE, 2010). The radiation from ceiling to ice could be estimated by the Boltzmann correlation which will be used later. One of the most important factors in ceiling radiation is emissivity index which is normally 0.85-0.95 for conventional materials used for roof ceiling construction. It is possible to do some coverings as aluminium foils or aluminium based clothes and paints which can reduce the ceiling radiation considerably, sometimes to 10%. Lighting is the second source of radiation to the ice as up to 60% of the light can be converted to heat and absorbed by ice (ASHRAE, 2010). A smart way to reduce this heat load is to control the light intensity according to the demand. In other words, it is not necessary to have lighting with full intensity all the working hours of the ice rink. When only few children play on the ice the lighting should be less intense in comparison with a professional match with hundreds of spectators. 2.4.3 Conduction Main contributors to conduction are ice resurfacing, brine pump work, brine headers, ground conduction and skaters. Ice resurfacing is a requirement to maintain the ice surface in a good condition. To fulfil this, ice resurfacing machine shaves the ice surface and then pours down a layer of hot water on the ice. The

23

Measurement and Modelling of Ice Rinks Heat Loads normal volume of water is 400-700 litres and the resurfacing water temperature is 30-80. The lower the water temperature, less heat load on the ice sheet would be dropped. (ASHRAE, 2010) Pump work causes an increase in the enthalpy of the secondary refrigerant passing the pump. Typical power consumption for these pumps is 15 kW. This 15 kW can be considered as a 15 kW heater in the brine circuit and that is why it should be tried to use variable speed pump to decrease this power consumption during unnecessary occasions including night shut down period. Headers are located along the length or the width of the ice rink. As they are colder than their environment, they should be insulated or covered with ice to decrease the heat gains (cooling losses) as much as possible. Ground conduction is a source of constant heat load on the system. The ground is separated by about 10cm insulation from the cold concrete but still there is some heat flux as the ground is heated to avoid the soil freezing. Skaters’ activities on the ice transfers heat through the ice surface but this is the only heat load that ice rink owners like to be as high as possible. To exemplify the heat loads shares, in Figure 13 there is a comparison of the shares of heat loads for two ice rinks in Canada and US, during three seasons. It can be seen that radiation, convection, pump work and ice resurfacing are the largest heat loads. Condensation has a significant effect during summer (humid season) but it is not considerable during winter (dry season).

100,00% 90,00%

skaters

Heat load shares (%)

80,00%

headers

70,00%

pump work

60,00%

ground conduction

50,00% 40,00%

ice resurfacing

30,00%

condensation

20,00%

lighting

10,00%

radiation

0,00%

convection Edmonton, Winter (95.4 W/m2)

Pittsburgh, Summer (135.5 W/m2)

Pittsburgh, Spring (114.3 W/m2)

Figure 13: Daily indoor ice rink heat loads (ASHRAE, 2010)

24


2.5 Energy efficiency in ice rinks There are several methods to use the input energy in ice rinks more efficiently but all of them can be categorized in one of the three classifications below:   

Heat loads decrease Refrigeration and distribution system performance improvement Ice/concrete slab quality enhancement

It should be noted that in energy efficiency applications instead of only focusing on the refrigeration system, it is better to study the whole building in an integrated approach. For example, maybe a method decreases the refrigeration system efficiency a little but in general, helps the waste heat recovery systems to supply the heat demands better. 2.5.1 Heat loads decrease 1. Low-e ceiling. Ceiling radiation is one of the largest heat loads in the ice rinks. The conventional materials for ceiling (wood, steel, etc.) have an emissivity index 0.85-0.95. The low-e ceiling concept referred to ceilings covered/painted with low-e aluminium based paints or suspended aluminium based clothes over the ceiling trusses. These low radiating ceilings have emissivity indexes in the ranges 0.05-0.2. By using low-e ceilings the radiation load can be decreased to 50% or less (Retscreen, 2003). It has an indirect impact as well which increases the light reflections and as a consequence less lighting is required in these arenas. The effect of less lighting is discussed in the lighting section. 2. Dehumidification. Humidity control is a way towards thermal comfort in all public areas but in ice rinks it has other significant effect as well. Water vapour in humid air over the ice rejects its heat to ice to condense on the surface. As a consequence, the humidity should be controlled and normally it is kept up to 50%-55%. Desiccant wheels are one of the most efficient dehumidification systems which fulfil the heat recovery from exhaust air and humidity regulation simultaneously. 3. Lighting. Less lighting has two impacts on the energy consumption, direct and indirect. The direct impact is less electricity consumption of the lamps. The indirect impact is according to the less heat transferred by radiation to the ice which decreases the refrigeration system electricity consumption. One way to decrease the lighting is to use more efficient lamps, for example using T5 or T8 fluorescent lamps instead of the metal halide lamps. Furthermore, the lighting intensity is not required to be max during the whole day and it can be adjusted according to the activity on the ice. 4. Resurfacing water. Ice resurfacing water quality, volume and temperature have significant effects on the heat loads. More purified/treated water makes a better ice with higher thermal conductivity. Moreover, less resurfacing water temperature and volume decreases the load. In Sweden 30°C-40°C is the normal temperature range (Makhnatch, 2010) while in North America 55°C-80°C is the case for many ice rinks. 5. Header pipes. Header pipes lying in the trenches should be insulated or frozen with an ice layer on them to reduce the cooling losses. Parts of the header pipes which are outside the trenches should as well be insulated. Other way to decrease the losses is to increase the brine temperature in pipes and 25

Measurement and Modelling of Ice Rinks Heat Loads then there is less temperature gradient between brine pipe headers and surrounding air. (Refer to section 2.5.3 and Table 3) 6. Air convection and air tightness. Temperature and velocity of the air moving above the ice surface has a great influence on the convection heat transfer. To control these parameters, direct air flows from ventilation diffusers should be avoided. At night, the ventilation system can be turned off or reduced down or the temperature of the supplied air during the less-crowded hours can be lowered. Pollution controllers is another way to adjust the ventilation demand as the big space inside the ice rinks buildings has too many leaks that sometimes ventilation looks unnecessary. Air tightness is to stop the uncontrolled movement of air into and out of a building which is not for a specific and planned purpose. Air tightness is another important factor to keep the building interior atmosphere isolated from outdoor conditions as in warm and humid seasons it can increase the convection and condensation heat loads severely. 7. Stands heating. To select the heating method for the spectators, its side effects as a heat load on the ice should be considered. One of the best solutions is infrared heater over the stands as it provides spot heating. If it is not possible and the ventilation system is used simultaneously for heating, it should be considered that the air should not be blown to the ice directly. 2.5.2 Refrigeration and distribution system performance improvement 1. Waste heat recovery. The refrigerant after the compressor is cooled in desuperheater and condenser. This heat can be recovered for heating demands. In this point of view, refrigeration system can be considered as a heat pump which the heat loads on the ice plays the role of the low level heat sources and the heat pump (refrigeration system) upgrades the heat level to distribute it in the required location and applications. According to the temperature degradation from desuperheater inlet to condenser outlet, the heat can be used for various applications including DHW, floor heating water, resurfacing water, snow melting, swimming pool heating, subfloor (soil) heating water, ventilation and space heating. 2. Brine pump. A full speed brine pumps working 24 hours a day can account for 15% (Retscreen, 2003) of the total electricity consumption in the refrigeration systems. Furthermore, this consumed electricity by brine pumps is converted as heat to the brine and therefore, it has direct and indirect negative effects on energy consumption. To decrease the consumption, variable speed pumps are one of the best solutions. They can be controlled by brine temperatures. 3. Brine pipe passes. The brine pipes are generally two pass configuration (as shown in Figure 12) but four pass configurations have been installed in several ice rinks over the world and no problem has been reported. Four pass layouts require less pumping power. (AAQ, 2003) 4. Compressor demand control. Similar to brine pumps, constant-speed compressors are not enough energy efficient as the amount of cooling required should be controlled by ice or brine temperature. At night or low-activity day hours it is not wise to have all the compressors on with full speed. For controlling the provided cooling capacity, electric motors equipped with frequency converters are used. In addition, using more than one compressor is a way to adjust the refrigeration with cooling demand. During the rush hours, the auxiliary compressor(s) can run in parallel with the first compressor. 26

Measurement and Modelling of Ice Rinks Heat Loads 2.5.3 Ice/concrete slab quality enhancement 1. Ice temperature and thickness. Considering the huge mass of ice (more than 40 ton for a typical 1800 m2, 25 mm ice sheet) each degree colder ice requires a significant amount of cooling. Therefore, any effort to keep the ice thickness as thin as possible or the ice temperature as high as possible will help using energy more efficiently. The recommended ice thickness is 25 mm. Recommended ice temperature varies according to the sport/activity. For hockey -6.5°C to -5.5°C, figure skating -4°C to -3°C and recreational skating -3°C to -2°C is satisfactory (ASHRAE, 2010). Overcooling the ice compared to these recommended values means energy waste and the ice rink owner should pay for it unnecessarily. During the night, when there is less heat load, these temperatures can be raised. Table 3 is a sample for ice temperature control. During midnight or early in the morning the ice temperature can be kept 2-3°C higher than the high load periods. It should be mentioned that the ice temperature can be manipulated by brine temperature adjustments and setting the control system for a schedule similar to Table 3, as an example. Table 3: Sample temperature control (Everything Ice, 2000)

Typical Daily Brine Cycle Period

Brine Temperature

Rink Function

0.00-6:00

-4⁰C

Night setback

6:00-8:00

-4⁰C

Ice maintenance

8:00-16:00

-6⁰C

Low load

16:00-18:00

-7⁰C

Figure skating

18:00-24:00

-8⁰C

Hockey

2. Concrete thickness and thermal conductivity. While the main object of cooling is ice, the heat transfer medium from ice to brine is concrete and less heat transfer resistance through concrete will cause less required cooling capacity. Concrete thickness in Swedish ice rinks is, typically, 150 mm which 25-50 mm is the distance from brine pipes top to the concrete surface (ice bottom). In parallel to concrete thickness, concrete thermal conductivity is very important. Better concrete quality leads to better heat transfer and less resistances.

27


3. EXPERIMENTAL MEASUREMENTS As the second step of this research, the performance of two refrigeration plants in two ice rinks is to be evaluated. To fulfil this objective, first, the ice rinks and their refrigeration systems are described in brief. Then, ClimaCheck, as a tool and method to analyse the performance of the refrigeration system is discussed and some modifications for more accurate calculations are suggested and applied. Finally, the results of the measurements (including power consumption and important temperatures) and estimated cooling capacity for some sample time periods are shown and discussed.

3.1 Ice rinks Two ice rinks which are studied in this research are situated in Norrtälje and Älta. Ice rink in Norrtälje belongs to the “Norrtälje Sportcentrum” which is located about one hour north-east of Stockholm. Älta ice rink is located in the Nacka district, Stockholm. 3.1.1 Norrtälje ice rink Norrtälje sport Centrum is comprised of two ice rinks, one indoor for hockey and figure skating and another one outdoor for bandy. Through this report only the indoor ice rink is considered and studied and for the outdoor rink another research project is in process. Moreover, there are an artificial soccer field, an indoor sport hall, a track and field pitch and a pool –to be constructed - in the sport facility. (Figure 14)

Figure 14: Norrtälje Sportcentrum, a) Ice hockey hall, b) Bandy, c) Artificial soccer field, d) Indoor sport hall, e) Track and field pitch, f) Pool (to be constructed)

28

Measurement and Modelling of Ice Rinks Heat Loads Among the neighbour sport fields, the artificial grass field is of particular interest as it uses part of the waste heat from the refrigeration system condenser to prevent freezing, even during the harsh winter climate (Figure 15). The heating piping system is similar to an ice rink but here the pitch is heated by the underground pipes containing ammonia 15% - water which is heated by condenser. In other words, the refrigeration system acts as a “heat pump” to keep the soccer field from freezing.

Figure 15: “Green soccer field” in “white winter” thanks to condenser waste heat

Norrtälje ice rink is about 1800 m2 (60 m×30 m) and its spectator capacity is 700-800 people, as shown in Figure 16 top-left. It is open all year round except for mid-April to mid-June. The refrigeration system is indirect with ammonia as refrigerant, calcium chloride 21% - water as secondary refrigerant and the coolant is ethylene glycol 35% - water. The refrigeration system is bought and shipped prefabricated from Finland. It is located outside the ice rink building as the inside old machinery room became useless after a price increase in district cooling water (Figure 16 bottom-right). Two MYCOM reciprocating compressors with nominal total cooling capacity of 300 kW and nominal 55 kW motor capacity are the driving forces of the refrigeration system. The evaporator is direct expansion and VAHTERUS shell and plate heat exchanger. The condenser and desuperheater are from the same manufacturer. There are two 15 kW brine pumps but only one is in operation during the measurements period of this research project. All the involved electrical motors mentioned are equipped with frequency converters. The humidity of the ice rink is controlled by a dehumidification system using a part of the cold brine to decrease the humidity ratio of the incoming air, mainly in the humid months. The heat rejected from the desuperheater is used to supply part of the heat required for space heating, ventilation, hot water and ice resurfacing water.

29


Figure 16: Norrtälje ice rink spectators stand (top-left), machinery room inside (top-right), compressors (bottom-left), and machinery room outside (bottom-right)

. 3.1.2 Älta ice rink The Älta ice rink is similar to Norrtälje ice rink comparing the size, spectator capacity and length of the season. The refrigeration system is indirect with ammonia as refrigerant, calcium chloride 24% water as secondary refrigerant and the coolant is propylene glycol 40% - water. Two GRAM reciprocating compressors with total nominal cooling capacity of 400 kW and 90 kW nominal motor capacity are the driving forces of the refrigeration plant in Älta ice rink, which are shown in Figure 17. These compressors were built in 1976 and brought from another ice rink, after being used a couple of years. It seems that they are over-sized for the required cooling capacity that might be a source of inefficiency. The evaporator is a flooded type plate heat exchanger (Figure 17 bottom-left). The condenser is an ALFA LAVAL plate heat exchanger and the desuperheater is a shell and tube heat exchanger. In Älta two nominal 15 kW brine pumps and two 11 kW coolant pumps run and none of the machineries are controlled by frequency converters. To control the humidity in Älta one desiccant wheel is installed in the hall.

30


Figure 17: Älta ice rink hall and spectator stand (top-left), machinery room (top-right), flooded evaporator (bottom-left), and one of the two compressors (bottom-right)

3.2 Performance analyser - ClimaCheck ClimaCheck is a tool to analyse the performance of refrigeration, air conditioning or heat pump systems. The motivation to use this tool is that from January 2009, an EU regulation requires that all air conditioning systems above 12 kW are to be “performance inspected”. The basic flowchart of ClimaCheck can be seen in Figure 18. 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. Furthermore, compressor electrical voltage and amperage are measured to know the electrical power input to the refrigeration system.

31


Figure 18: ClimaCheck basic instrumentation configuration

The ClimaCheck instrument can be a portable field kit or permanent fixed installation. Both of them can be connected to the internet to be monitored anywhere and the logged data can be processed through the ClimaCheck software to obtain the required calculated results. 3.2.1 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 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 according to Figure 19:

.

.

.

1

Where : Refrigerant mass flow rate : Electric motor efficiency : Electric power to the compressor motors

: Heat loss from compressor body and/or compressor cooling by oil/water

.

: Enthalpy after compressor

.

: Enthalpy before compressor 32


Figure 19: Energy balance over compressor

The enthalpy after the evaporator is known with two temperature and pressure sensors. The enthalpy before the evaporator is equal to the enthalpy before the expansion valve and this can be known by measuring the temperature before the expansion valve and the known high pressure. By knowing the refrigerant mass flow rate and the enthalpies before and after the evaporator, the cooling capacity can be calculated. The COP of the system can be estimated as the cooling capacity and electrical power are known. The heating capacity of the condenser and desuperheater are similarly possible to estimate as the refrigerant mass flow rate and temperatures in and out from condenser and desuperheater are measured. In addition to the measured data, two parameters are important for an accurate calculation; the electrical motor efficiency and the heat rejection from the compressor body. The precise electric efficiency is not easy to know since type and power of motor, motor age and motor full/part load will influence this. In addition, the frequency converter efficiency should be considered as well. For the Norrtälje compressors it is mentioned in their manual that the motor efficiency is 90% in standard conditions. The frequency converters efficiency normally ranges 97-99% and for the calculations here the worst case (97%) is considered. It means that in general the total electrical efficiency of the frequency converter and electric motor is 87.3% (=0.90×0.97). In Älta, there is no frequency converter and the compressor efficiency is not mentioned anywhere but according to a paper from US department of energy (motor challenge, 2010) typical compressor efficiency with this capacity ranges 90-92% where the worst case is chosen. It is worth mentioning that Älta compressors are very old and running most of the time part load that can influence the compressor efficiency significantly but in the calculations 90% is considered as the compressor motor efficiency. To estimate the heat losses from the compressor body 7% of the input power is suggested by ClimaCheck (Berglöf, 2010). Moreover, the amount of compressor cooling by oil and/or water should, if possible, be measured in the field or the compressor design-software can be used to find the amount of energy losses. For Älta, the compressors cylinder heads are cooled by water and the supply and return temperatures are measured. The water mass flow rate is estimated to be 0.1 kg/s and therefore, the amount of compressor cooling is calculated. In Norrtälje, this heat rejection by compressor body 33

Measurement and Modelling of Ice Rinks Heat Loads was unknown and assumed to be constantly 4 kW by ClimaCheck but it does not seem to give an accurate estimation of the cooling capacity. As a consequence, it was decided to use MYCOM compressors software for better estimates. The modification to ClimaCheck calculation is described in the next part.

Figure 20: ClimaCheck flowchart for Norrtälje and Älta ice rinks

In Norrtälje and Älta ice rinks, to monitor the heat loads on the ice sheet, the ice temperature, air temperature and humidity close to ice surface, indoor and outdoor temperatures are measured as well. Figure 20 indicates a flowchart of the measurement configuration. 3.2.2 ClimaCheck method modification 7% heat rejection assumption is good for hermetic and semi hermetic compressors but for the open compressors the amount of energy loss to input power ratio is more than this 7%. Normally, the open reciprocating compressors are cooled by oil or water. In Älta, as explained before, the mass flow rate and temperatures of the cooling water are measured and therefore the amount of energy losses by compressor and its cylinder heads can be calculated by:

∗

∗∆

%∗

∗

2

In Norrtälje, the compressor software is used to find the amount of oil and jacket heat rejection. Figure 21 shows the ratio of heat rejected to absorbed power versus condensing temperature for several compressor speeds (a sample calculation is shown in appendix, section 8.1). Heat rejection is the sum of oil and jacket heat rejection. The absorbed power means the shaft power and is equal to the electrical power multiplied by electrical efficiency. This graph is selected among four graphs drawn for four conditions: -10°C or -12°C evaporation temperature and zero degree 34

Measurement and Modelling of Ice Rinks Heat Loads subcooling/superheating or 5K subcooling and 7K superheating. After studying many conditions it seems that the below conditions (-10°C evaporation, 5K subcooling and 7K superheating) is the most frequent one.

(Heat rejection/Absored Power) %

25,00

20,00

RPM=900 RPM=1000 RPM=1100

15,00

RPM=1200 RPM=1300 10,00

RPM=1400 RPM=1500 RPM=1600

5,00

RPM=1700 RPM=1800

0,00 0

10

20

30

40

50

Condensing Temperature (°C) Figure 21: Relative heat rejection versus RPM and tcond (teva=-10°C, superheat =7K, subcool = 5K)

To have a correlation for compressor loss, average heat rejection over absorbed power ratio is drawn for various condensing temperatures as shown in Figure 22. As the Norrtälje compressors are cooled better with the cooling media, it is assumed that only 5% of the energy is lost from the compressor body. Therefore, the amount of compressor loss for Norrtälje can be estimated by:

∗

∗

.

.

%∗

∗

3

35

(Heat rejection/Absored Power) %

Measurement and Modelling of Ice Rinks Heat Loads 25,00 20,00 15,00

y = 0,2046x + 12,215

10,00 5,00 0,00 0

10

20

30

40

50

Condensing temperature (°C) Figure 22: Average relative heat rejection versus tcond (teva=-10°C, superheat =7K, subcool = 5K)

3.3 Measurement results The measurements have been done concentrating on the refrigeration system, the ice and its surrounding climatic conditions in order to analyse the performance of the refrigeration system in response to fluctuating heat loads, and in a further step, to find the major and minor heat loads. In both ice rinks, a majority of the ClimaCheck instruments are installed in the machinery room and in places which have been explained before. Furthermore, to record the ice temperature, two temperature sensors have been embedded in the top and bottom of Norrtälje ice, but in Älta there is only one ice temperature sensor. Moreover, some air temperature and relative humidity sensors record the data in three locations, a few centimetres above the ice, some meters above the ice (ice hall indoor conditions) and ice rink outdoor. As a great amount of data have been recorded every minute in each ice rink, there are vast amounts of logged data available and in following parts just samples of the most important ones have been shown and analysed. There is much more stability in the Norrtälje performance, hence majority of the graphs, to be discussed, belongs to Norrtälje. 3.3.1 “Cooling chain” temperatures The cooling chain temperatures show how the provided cooling is transmitted and distributed from evaporator to ice (in reality the direction of the energy flow is vice versa, from ice to evaporator). As an example, Figure 23 shows temperatures of the evaporation, brine in and out from evaporator and ice top/bottom for the day 15 March 2011 for Norrtälje ice rink. These profiles can help to check whether the system works properly and stable or not. The observation from this figure is that the refrigeration system seems to work in stable conditions as the evaporation temperature is stable. When one compressor is in operation, the evaporator temperature is around -10.4°C and when the parallel second compressor starts, it decreases to below -11°C. 36

Measurement and Modelling of Ice Rinks Heat Loads Another parameter to examine is the brine in and out temperature difference. It is kept constant to around 1.5-2°C thanks to the variable speed brine pump which regulates the brine mass flow rate. Moreover, it can be seen that the number of compressors working is governed by brine return temperature (brine in) with a set point of -4°C ± 0.5°C. Furthermore, ice top and bottom temperatures are also shown in Figure 23. The two temperatures are measured with PT1000 sensors embedded in the ice. The “bottom” sensor is put on the concrete surface before the ice build-up started and the second is placed approximately 15 mm up in the ice, which is in fact about 10 mm from the actual surface. It can be seen clearly that ice resurfacing with warm water takes place six times; once early in the morning before 7:00 start, once in the midday and four times in the evening during the intensive activity hours. Finally, this graph gives the opportunity to check the “the temperature waste” from the source, evaporator, to the user, ice. It can be seen that the highest temperature increase occurs during the heat transfer between the evaporator and brine. It might be as a result of a too small heat transfer area and/or the requirement for superheat in the evaporator..

Temperature (°C)

‐1 ‐3 ‐5

BRINE IN

‐7

BRINE OUT

‐9

ICE BOTTOM

‐11

ICE TOP

‐13

EVAPORATOR

00:00:00 00:52:00 01:44:00 02:36:00 03:28:00 04:20:00 05:12:00 06:04:00 06:56:00 07:48:00 08:40:00 09:32:00 10:26:00 11:18:00 12:10:00 13:03:00 13:55:00 14:47:00 15:40:00 16:32:00 17:24:00 18:16:00 19:08:00 20:01:00 20:53:00 21:45:00 22:37:00 23:29:00

‐15

Figure 23: Ice, brine and evaporating refrigerant temperature fluctuations, 15 March 2011, Norrtälje

3.3.2 Evaporation/Condensation temperatures Figure 24 shows the evaporation and condensation temperatures for the same day. It can be seen that at the start and end of the day two compressors run. During the rest of the day one compressor is enough to compensate for heat loads. It seems that two compressors are necessary during the first hour of midnight because the effects of the heat loads from last night are still remained on the ice. The reason for two compressors running during the last three-four hours is the heat loads of rush evening hours. There were some more activities on the ice and there are four ice resurfacings (Figure 23). As a consequence, the second compressor starts to work in parallel. It can be seen that there is a time lag between the ice resurfacing heat loads start (around 18:00) and the second compressor being on (around 20:00); the ice acts as a thermal storage body. 37

25,0

00:00:00 00:50:00 01:40:00 02:30:00 03:20:00 04:10:00 05:00:00 05:50:00 06:40:00 07:30:00 08:20:00 09:10:00 10:02:00 10:52:00 11:42:00 12:32:00 13:23:00 14:13:00 15:03:00 15:54:00 16:44:00 17:34:00 18:24:00 19:14:00 20:05:00 20:55:00 21:45:00 22:35:00 23:25:00


Temperature (°C)

20,0 15,0 10,0 5,0 0,0 -5,0

Two compressors

One compressor

-10,0 -15,0 -20,0

Evaporation temperature

Condensation temperature

Figure 24: Evaporation and condensation temperatures, 15 March 2011

3.3.3 Brine and coolant temperatures The temperatures of the brine and coolant have been indicated in Figure 25. Brine in and out temperature difference is kept constant 1.5-2°C, due to brine pump variable speed operation. The temperature difference of the coolant in and out is around 10°C when one compressor runs and during the two compressors operation simultaneously, the coolant out temperature rises about 5°C. 25

15 10

BRINE IN BRINE OUT

5

COOLANT IN

0

COOLANT OUT

‐5 ‐10 00:00:00 00:52:00 01:44:00 02:36:00 03:28:00 04:20:00 05:12:00 06:04:00 06:56:00 07:48:00 08:40:00 09:32:00 10:26:00 11:18:00 12:10:00 13:03:00 13:55:00 14:47:00 15:40:00 16:32:00 17:24:00 18:16:00 19:08:00 20:01:00 20:53:00 21:45:00 22:37:00 23:29:00

Temperature (°C)

20

Figure 25: Brine and coolant temperatures, 15 March 2011

38

Measurement and Modelling of Ice Rinks Heat Loads 3.3.4 Air temperature and relative humidity over ice, indoor and outdoor The results of the air temperature and relative humidity measurements in three locations (over ice, in the ice rink and outdoor) have been shown in Figure 26. It can be seen that the relative humidity in the ice rink is kept lower than 45-50%, thanks to the dehumidification system. The relative humidity close to the ice surface is higher, 50-60%, because the air temperature (light blue line) is colder in comparison to ice hall temperature (red line). It is worth mentioning that this graph is for 15 March 2011, a relatively cold and dry day. In describing the calculations for condensation, a humid warm summer day is chosen to demonstrate the significance of relative humidity impact on the condensation heat load.

70

25

60

20

50

15

40

10

30

5

20

0

10

‐5

0

Relative humidity %

30

rh ishall rh over ice temp ishall outdoor temp temp over ice

00:44:00 01:42:00 02:40:00 03:38:00 04:36:00 05:34:00 06:32:00 07:30:00 08:28:00 09:26:00 10:24:00 11:22:00 12:20:00 13:18:00 14:16:00 15:14:00 16:14:00 17:13:00 18:11:00 19:09:00 20:07:00 21:05:00 22:03:00 23:01:00 23:59:00

Temperature (°C)

Looking at the three temperatures shown in Figure 26, stable indoor conditions can be compared with outdoor temperature fluctuations of more than 10°C. It is a sign for stable energy systems (heating, ventilation, cooling) operation and good insulation.

Figure 26: Air temperatures and relative humidity, 15 March 2011

3.3.5 Electric power input and cooling capacity Measuring the electricity input to the refrigeration system is one of the most important logged data as it is the main requirement to be able having an energy balance over the compressor to calculate the refrigerant mass flow rate. To find this power, voltage and amperage of the electricity is measured. In Figure 27, the electricity consumption is shown for the same sample day. For this day one compressor starts to consume about 19 kW and in next hours it increases the consumption to 23 kW softly, due to frequency converter operation. During the two compressors parallel running, the total consumption is around 38 kW. It means that the compressors never run with their full speed in this day according to the control system functions. However, it has been observed that in warm 39

Measurement and Modelling of Ice Rinks Heat Loads summer days or for the “ice making” periods, the consumed electricity by the compressors is more than 70-80 kW.

250 200 150 100 50

Power -Cooling Capacity (kW)

300

00:46:00 01:49:00 02:52:00 03:55:00 04:58:00 06:01:00 07:04:00 08:07:00 09:10:00 10:15:00 11:18:00 12:21:00 13:25:00 14:28:00 15:32:00 16:35:00 17:38:00 18:41:00 19:45:00 20:48:00 21:51:00 22:54:00 23:59:00

0

Q cooling (kW)

Electric Power (kW)

Figure 27: Electric power and cooling capacity, 15 March 2011

When the input power has been measured, the ClimaCheck method makes it possible to calculate the system cooling capacity and COP. On 15 March 2011, COP by one compressor is 5.5-6 but by two compressors COP is around 5. This come from more difference between evaporation and condensation temperatures during the peak loads (Figure 24). The refrigeration system has to provide colder refrigerant, rejects the heat with higher temperature and the pressure ratio over compressors increase, as well. 3.3.6 Total energy consumption The total electricity consumption of the refrigeration systems and the outdoor temperatures of the two rinks for March 2011 are shown in Figure 28. The total energy consumption is the sum of the compressor electricity consumption and the auxiliary equipment including brine and coolant pumps and dry cooler fans. The energy consumption is clearly higher in Älta despite the same outdoor temperatures. The main reason for higher daily consumption in Älta (1550 kWh/day average) in comparison with Norrtälje (920 kWh/day average) is the auxiliary pumps and fans. The compressors in Älta are not capacity controlled and they are switched on-off several times per day but the brine and coolant pumps are running continuously with full rated powers of 15 kW and 11 kW respectively, even when the compressors are off. On the other side, in Norrtälje, these pumps change their speed according to the cooling load. 40


1800

29

1600

26

1400

23 20

1200

17

1000

14

800

11

600

8 5

400

2

200

Outdoor Temperature (°C)

Energy consumption (kWh/day)

The second reason for the higher electricity consumption may be related to the ice temperature. While the average ice temperature in Norrtälje is -4°C to -3°C, it is -6°C to -5°C in Älta. A lower ice temperature increases the energy consumption in two ways. First, lower ice temperature requires more cooling and hence, more electricity consumption. Secondly, colder ice increases the amount of absorbed heat through radiation-convection-condensation.

-1

0

-4 1

2

3

4

5

6

7

8

9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24 25 26 27 28 29 30 31

Älta consumption (kWh/day)

Norrtälje consumption (kWh/day)

Älta temp.

Norrtälje temp.

Figure 28: Total refrigeration system energy usage and outdoor temperatures in March 2011

The last reason for higher energy consumption in Älta is related to the ventilation ducts configuration. In Norrtälje there is one ventilation duct header over the spectators’ stand, while in Älta there are two ventilation headers; one over the stands and another one in the middle of the ice rink and over the ice sheet (refer to Figure 53). This may cause more air movements over the ice, with higher temperature and more velocity in comparison to Norrtälje. That is why convection is one of the major heat loads in Älta. Table 4 summarizes the main reasons for more energy consumption in Älta. Table 4: Reasons for higher energy consumption in Älta

Älta (1550 kWh/day)

Norrtälje(920 kWh/day)

Compressors

On – Off periodically

Capacity controlled by brine temperature

Brine & Coolant pumps

Without frequency converter

With frequency converter

2 Main distribution ducts, one over ice/one over stands

1 Main distribution duct over the stands

-6°C to -5°C

-4°C to -3°C

Ventilation

Ice temperature

41


4. ANALYTICAL MODELLING In the experimental measurement part it is described how the cooling capacity can be estimated by ClimaCheck method and the explained modifications. This cooling capacity is an indication of the total heat loads on the ice sheet plus the heat gains of the provided cooling in its way from evaporator to ice. In the analytical modelling part, the objective is to understand what the various heat loads are and how they can be calculated/estimated. In other words, the “total heat load” is broken down to find the magnitude and share of each heat load. This step helps any technician or ice rink owner to discover what the main sources of heat loads in the rink are. Furthermore, another method is introduced to guess the amount of heat loads on ice, just by knowing and using the ice top and bottom temperatures. It is not as precise as the ClimaCheck method, but is a simple useful gauge to present the magnitude of the heat loads. Finally, the heat transfer from ice to the evaporator is evaluated according to the temperature resistances. The objective is to discover where the biggest resistances against the heat flow are.

4.1 Heat loads In part 2.4 heat loads are classified according to their type of heat transfer mechanisms. In another point of view, these loads can be categorised into two groups according to their location of impacts: those which directly impact the ice, trying to melt it down and those heat gains that have impacts along the refrigerant-brine-concrete-ice path and increase the required cooling capacity. Figure 29 demonstrates these two groups of heat loads. Convection, condensation, radiation, lighting, ice resurfacing and skaters activities are heat loads for ice, directly and for the refrigeration system, indirectly. The ground conduction, brine headers heat gain and brine pump’s work belong to the second group which can be called “distribution system losses”. .

Figure 29: Heat loads in ice rinks and their impact points

42

Measurement and Modelling of Ice Rinks Heat Loads In the following sections the correlations and methods of estimating each heat load is introduced and a result sample for each calculation is presented and discussed. Then, for a whole day heat loads share is calculated to find the major and minor heat loads.

4.1.1 Radiation Radiation is one of the most significant heat loads in ice rinks as two large surfaces of cold ice and relatively warm ceiling face each other. To calculate the radiation the Stefan-Boltzmann equation can be applied:

4

Where σ is the Stefan-Boltzmann constant and equals to 5.67×10-8 W/(m2K4) and ƒci is gray body configuration factor for radiation from ceiling to ice surface and can be calculated by:

5

ɛ is the emissivity factor and for normal ceiling constructions and ice structure ranges from 0.8 to 0.95. Fci is the angle factor and depends on the geometric properties of the ceiling surface, ice surface and the ice rink height. The geometry inputs are the following, ice sheet 60×30 m, ceiling height and surface 9 m and 2200 m2 respectively. Graphs like what is shown in Figure 30 are used to avoid the complex correlation for calculating this angle factor. An example of a radiation calculation for one of the ice rinks is shown in Figure 31 as a sum of radiation, convection and condensation heat loads. This will be discussed after convection and condensation explanations. It should be mentioned that for ceiling temperature, as there was no installed temperature logger in the height, an average value of indoor and outdoor temperature is considered. As explained in “future work” chapter, measuring the ceiling temperature with infrared cameras is going to be done which can make the radiation calculations more accurate.

43


Figure 30: Angle factor between two aligned parallel faces (Çengel, 2007)

4.1.2 Convection The temperature of the air moving above the ice is a few degrees higher than zero and the ice temperature is few degrees lower than zero. This temperature difference is the driving force for convection. The amount of convection heat load can be calculated by equation (6):

6

αc is the convective heat transfer coefficient. It can be calculated by an experimental equation (ASHRAE, 2010) which uses the air velocity over ice to take into consideration both natural and forced convection: .

.

The heat transfer coefficient is assumed to be 5 W/m2·K as a reasonable assumption with air velocity 0.4-0.5 m/s over ice (DOE, 1980). 44

Measurement and Modelling of Ice Rinks Heat Loads 4.1.3 Condensation The driving force for water vapour condensation is the difference between the partial pressure of water vapour (in the air close to ice surface) and the saturation pressure of water at ice top surface temperature. The general condensation heat transfer equation is similar to convection and the difference is in the condensation heat transfer coefficient instead of the convective one. Condensation, which in some books known as diffusion, can be estimated by equation (8)

αd is the diffusion (condensation) heat transfer coefficient and be calculated by equation (9) (Granryd, 2005):

∗

is air relative humidity. P1 and P2 are water vapour saturation pressures in air and on ice top surface can be calculated by equations (11) and (12):

∗

.

∗

.

. .

It should be mentioned that equation (11) is appropriate for temperatures higher than zero while equation (12) is suitable for temperatures lower than zero. By knowing the correlations to calculate radiation, convection and condensation heat loads, a time period is chosen to examine these estimations. In order to eliminate the effects of lighting, ice resurfacing and skaters activities, the first six hours of March 15 midnight is selected. It has been tried to check these calculations with the cooling capacity minus ground and pump heat gains. The results for the calculations shown in Figure 31 indicate a good match. It is indicated that average net cooling capacity is 98.5 kW while the average heat loads on the ice is 92.5 kW. The slight 6 kW difference is due to brine headers heat loads which will be discussed later.

45


Q cooling (kW) ground-pump gains eleminated 110 105 100 95 90 85 80 75 70

00:00:00 00:12:00 00:24:00 00:36:00 00:48:00 01:00:00 01:12:00 01:24:00 01:36:00 01:48:00 02:00:00 02:12:00 02:24:00 02:36:00 02:48:00 03:00:00 03:12:00 03:24:00 03:36:00 03:48:00 04:00:00 04:12:00 04:24:00 04:36:00 04:48:00 05:00:00 05:12:00 05:24:00 05:36:00 05:48:00 06:00:00

Cooling-heat loads (kW)

heat loads (kW)radiation-convection-condensation

Figure 31: Cooling capacity versus heat loads at midnight, 15 March 2011

The chosen day was a cold dry day and as a consequence, condensation has a share of less than 2%. For this six-hour period, convection accounts for about 56% and radiation 42% of the total heat load. To study the condensation load and its effect better, the warmest day of the humid summer 2010 is studied below. As shown in Figure 32, on this day relative humidity above the ice and in the ice rink is 60-70%. It can be compared with Figure 26 to realize the difference. 80

20

60 50

10

40 5

30 20

0

10 -5

temp ice top temp ice hall temp over ice rh ice hall rh over ice

0 00:00:00 01:15:00 02:30:00 03:45:00 05:00:00 06:16:00 07:31:00 08:46:00 10:02:00 11:16:00 12:32:00 13:47:00 15:03:00 16:19:00 18:02:00 19:17:00 20:32:00 21:47:00 23:03:00

Temperature (°C)

15

Relative humidity %

70

Figure 32: Air and ice temperatures and air relative humidity, 12 July 2010, Norrtälje

46

Measurement and Modelling of Ice Rinks Heat Loads On March 15, 2011 the condensation heat transfer coefficient is very smaller than the convective heat transfer coefficient but on July 12, 2010 it is shown in Figure 33 that while αc is equal to 5

2 1,8 1,6 1,4 1,2 1 0,8 0,6 0,4 0,2 0

α_d

00:00:00 01:19:00 02:38:00 03:57:00 05:17:00 06:36:00 07:55:00 09:14:00 10:34:00 11:52:00 13:12:00 14:32:00 15:51:00 17:38:00 18:58:00 20:17:00 21:36:00 22:56:00

Heat transfer Coefficient (W/m2.K)

W/m2.k, average αd is 1.1-1.4 W/m2.k. In other words, for this humid day condensation heat load is about 25-30% of the convection heat load. This leads to a share of 10% in total heat load for condensation.

Figure 33: Condensation heat transfer coefficient, 12 July 2010, Norrtälje

4.1.4 Lighting Each lighting fixture in Norrtälje is composed of three lamps each consumes 48 watts electricity. There are 6 rows of lighting rails and each rail contains 38 lighting fixtures, then the total lighting electricity consumption should be 32.8 kW approximately. According to ASHRAE up to 60% of this input energy can be converted to the radiated heat but as fluorescent lamps are among the energy efficient lighting devices, 40% is assumed as the ratio of radiated heat from lamps to the ice, which is 13.1 kW.

47


Figure 34: Lighting fixtures in Norrtälje

4.1.5 Ground Conduction To calculate the amount of heat conduction from ground to the concrete, which increases the amount of required cooling capacity, it is assumed that the concrete temperature is -5°C as the brine in and out average are -4°C and -6°C, respectively. Ground temperature is kept +5°C by soil heating pipes, and then there is 10°C temperature difference between the ground and the concrete. The distance between these two heat transfer sheets is filled with 10cm concrete (brine pipe top to concrete bottom) and 10cm thick polystyrene insulation. Thermal conductivity for the concrete is 1.7 W/mK and polystyrene is 0.04 W/mK (Makhnatch, 2010 and FAO, 2003). These parameters lead to 7 kW conduction heat gains from ground. 4.1.6 Brine headers Supply and return brine headers are located along the 30 m width of the ice rink. These two 150 mm pipes are in a trench and they are covered with 2 cm thick ice as a heat transfer resistance. The distance between the trench and the evaporator, inside the machinery room is 10m and the pipes have 2 cm Armaflex insulation, 0.033 W/mK thermal conductivity. These data can be used to calculate the amount of heat gain by the brine pipe headers. Whether the resistance to heat transfer is the pipe wall, the 2-cm ice ring around pipe or the insulation, the conduction heat transfer is equal to:

Using equation (13) and the parameters summarized in Table 5 leads to 6.7 kW heat load on the refrigeration plant due to brine headers heat gains.

48

Measurement and Modelling of Ice Rinks Heat Loads Table 5: Physical properties for brine headers for heat load calculation

Thermal conductivity (W/m.K)

20m pipe-insulation

Pipe Di = 77

Pipe = 43

60m pipe-ice

Size (mm)

Pipe Do = 84

Ice = 2.25

Ice thickness = 20 Pipe Di = 77

Pipe = 43

Pipe Do = 84

Insulation = 0.033

Insulation thickness = 20

Temperature (°C) Brine = -5 Ice surface = -3 Brine = -5 Ambient = 10

4.1.7 Ice resurfacing Ice resurfacing adds a significant amount of heat to the ice. The amount of resurfacing heat can be estimated by equation (14): (ASHRAE, 2010).

Qresurfacing 1000 Vf .

tf

t i

Where Vf is the flood water volume (m3), tf and ti are the water and ice temperatures (°C). Flooded water is typically 0.4 to 0.7 m3. In Norrtälje flood water volume and temperature are 600 litres and 35°C, respectively. Ice temperature is considered to be -1.5°C according to Figure 23. These input data result in about 81 kWh heat addition to ice for each ice resurfacing. To study the significance of ice resurfacing heat load an exercise has been done. On March 15, two six-hour period has been chosen, 12:00-18:00 when there is no ice resurfacing and 18:00-24:00 when there are four ice resurfacings. The produced cooling capacity for the first six hours is 701 kWh while for the second six-hour period it is 912 kWh. It means that there is a need for 30% more cooling in the second period because there are several ice resurfacings. It was mentioned before that each ice resurfacing heat addition is 81 kWh. It means that four ice resurfacings have 324 kWh total load. From this 324 kWh, 211 kWh (=912-701) is compensated by the refrigeration system and the remaining is stored in ice body, which later and during the next midnight will be transferred to the refrigeration system as well. This heat absorption causes an increase in ice temperature during the evening and has been shown in Figure 35. It is discussed before that when the ice becomes warmer and there is a risk for melting, the second compressor starts to work in parallel with the first one. Furthermore, during the less crowded hours of midday ice is in rest and the temperature decreases slightly.

49


0

Temperature (°C)

-0,5 -1

hourly avergae ice temperature

-1,5 -2 -2,5 -3

two compressors

-3,5 Figure 35: Hourly average ice temperature, 15 March 2011, Norrtälje

4.1.8 Pump work The brine pump nominal power in Norrtälje is 15 kW and the speed is controlled by a frequency converter. When the compressors are off, this pump is jogging with a low frequency. This is necessary because the system control is based on the brine return temperature. When one compressor works, this pump operates at 40Hz and consumes about half of the nominal capacity, 7.5 kW. During the peak load conditions the brine pump frequency is 50Hz and the electricity consumption is close to nominal capacity, 15 kW. As a consequence, the heat load from the brine pump has a step function with three values; “near zero”, 7.5 kW and 15 kW according to the cooling capacity demand.

4.1.9 Skaters Skaters heat load is very hard to estimate precisely as there should be detailed information of activities, their duration and type (heavy, light, etc...) and even if this information is available still there is no scientific methods introduced to calculate the heat load, for example by skates friction on the ice or skaters body radiation. ASHRAE suggests assuming 4% of the total heat loads regarding to skaters activities on the ice.

50

Measurement and Modelling of Ice Rinks Heat Loads 4.1.10 Results for heat loads shares In the previous nine sections it has been described how to calculate, estimate or for a few cases have a logical good guess for each heat load in the ice rinks. Table 6 shows a summary for six heat loads other than radiation, convection and condensation. The other three ones do not have a relatively constant value as these six ones have, during their period of impacts. By “period of impact” it is meant that for example lighting is a heat load during the working hours of the ice rink, around 18 hours a day. Table 6: Heat loads summary

Heat source

Heat Gain (kW/kWh)

Notes/assumptions

Lighting

13 kW

Heat/emitted light = 40%

Ground

7 kW

+5°C ground and -5°C concrete, 10 cm polystyrene, 10 cm concrete

7.5 kW (1 Compressor)

40Hz

15 kW (2 Compressors)

50Hz

Ice resurfacing

81 kWh per resurfacing

35°C resurfacing water, 600 litres flood water

Brine header pipes

6.7 kW

Skaters

4% of total heat load

Brine pump

20m insulated pipe headers 60m frozen pipe headers -

In order to find the share and weight of the heat loads in total heat load on the ice sheet and refrigeration system, daily heat loads estimation has been fulfilled for 15 March 2011 in Norrtälje ice rink and the results are demonstrated in Table 7 and Figure 36. Average radiation heat load is 33 kW and daily average convection-condensation is 51.5 kW. These heat loads applied during the whole 24 hours. Lighting is applied from 6:00 in the morning to 24:00, and then the 13 kW load is not applied on the ice during six hours of midnight. Brine header and ground conduction have their loads all the day. Brine pump average heat load is calculated to be 9 kW as in some hours it runs by 7.5 kW power and the remaining with 15 kW. Furthermore, there are six ice resurfacings. All the above heat loads account for 96% of the total daily heat load and the remaining 4% is skaters’ heat load. As it is shown in the table, the total heat load is equal to 3 430 kWh. The refrigeration system should compensate this heat load. To examine this, the average cooling capacity is calculated to be 142 kW as in many hours the cooling capacity is 120-130 kW and in few hours it is 180-190 kW (refer to Figure 27) and then the total daily cooling is 3 408 kWh which shows a good match with the heat loads sum.

51

Measurement and Modelling of Ice Rinks Heat Loads Table 7: Daily heat load calculations, 15 March 2011

heat load

kW

kWh

radiation

33

792

lighting

13

234

convection - condensation

51.5

1236

ice resurfacing

81 kWh/res.

486

header pipe

6.7

160.8

brine pump

9 (Avg.)

216

ground

7

168

skaters

4% of total

137.2

total daily heat load

3430

total daily cooling capacity

142 (Avg.)

3408

Heat loads shares are indicated in Figure 36. Convection, radiation, ice resurfacing and lighting are the four biggest heat loads in this ice rink while condensation (less than 3% of convection) and skaters are the smallest ones. Conduction heat gains of the distribution system – ground, header pipe and brine pipe - have almost equal shares. This figure demonstrates that about 84% of the total heat loads are the loads on the ice and 16% are the losses in the distribution system. It should be reminded that condensation has a bigger share in summer, as discussed before. ground brine pump 5% 6%

skaters 4%

header pipe 5%

radiation 23%

ice resurfacing 14%

lighting 7%

convection +condensation 36%

Figure 36: Heat loads shares in the total heat load

52


4.2 Heat flux method Top and bottom ice temperature fluctuations can be used to estimate the amount of heat load on ice sheet and the amount of heat transferred to the cold concrete slab. To measure these, two temperature sensors are installed in the ice as indicated below. From Fourier-Biot, the general transient conduction equation, the ice sheet case can be simplified to one dimensional transient conduction (diffusion equation), as the thickness of the ice is only 0.1-0.2% of the ice length and width and hence equation (15) governs the heat transfer process in ice:

T represents the ice temperature, t for time; z is the ice thickness axis and α is the thermal diffusivity of ice which equals to k/ρ·Cp of the ice (thermal conductivity over the density and specific heat capacity which are considered 2.25 W/mK, 912 kg/m3 and 2087 J/kgK respectively). The temperature profile is T=Az2+Bz+C which “A”, “B” and “C” would be known by boundary conditions and temperature as a function of time. If the z axis starts from the top ice and goes downward to ice bottom, it is possible to say z=0 for ice top and z=depth difference between two sensors for ice bottom (15mm in Norrtälje). Then “C” is equal to ice top temperature at z=0. “A” can be known by using equation (15) and twice differentiating the temperature profile equation. When “A” and “C” are known, “B” can be calculated by using ice bottom temperature. To conclude, as the temperature fluctuations over time and place are measured and known, heat flux to the top and bottom surface of the ice can be achieved by equation (16) and then by multiplying to ice surface area (1800 m2) heat transfer rate can be calculated:

top ice temp. sensor

Ice bottom ice temp. sensor

Figure 37: Top and bottom temperature sensors embedded in the ice

53


4.2.1 Results The results of the calculation by this method are shown in Figure 38 for the days 14-16 April 2011. The reason for selecting these three days is that at 20:30 on 15 April the refrigeration system was switched off and at 6:30 of the third day the compressor was switched on again. The effect of this switch off-on is seen as the lowest heat transfer rate. Top Heat Transfer Rate (kW)

Bottom Heat Transfer Rate (kW)

Heat transfer rate (kW)

140 120 100 80 60 40 20 0-1 2-3 4-5 6-7 8-9 10-11 12-13 14-15 16-17 18-19 20-21 22-23 0-1 2-3 4-5 6-7 8-9 10-11 12-13 14-15 16-17 18-19 20-21 22-23 0-1 2-3 4-5 6-7 8-9 10-11 12-13 14-15 16-17 18-19 20-21 22-23

0

Figure 38: Top and bottom heat transfer rate on days 14, 15 and 16 April 2011

0,00

Shut down period

‐1,00 ‐2,00 ‐3,00

21‐22

18‐19

15‐16

12‐13

9‐10

6‐7

3‐4

0‐1

21‐22

18‐19

15‐16

12‐13

9‐10

6‐7

3‐4

0‐1

21‐22

18‐19

15‐16

12‐13

9‐10

6‐7

3‐4

‐4,00 0‐1

Temperature (°C)

The difference between the top and bottom heat flow rate may show the ice temperature rise or decrease. For instance, during the compressor-off period, as the brine becomes warmer, less heat is transferred to the concrete slab and the absorbed heat causes the ice temperature to rise. Furthermore, the steep decrease in heat gain during the compressor-off period seems to indicate the significant effect of ice temperature on radiation and convection heat gains. In other words, colder ice absorbs more heat and vice versa.

Figure 39: Ice average temperature fluctuations on 14, 15 and 16 April 2011

54


4.3 Temperature resistances-differences Some other calculations can be done to analyse the performance of refrigeration and distribution system. The objective of these calculations is to find the places which are the main resistances to the heat transfer from the ice to the refrigeration system. As it was mentioned before, Figure 40 demonstrates the heat flow from the heat loads above the ice to the refrigeration system. In this energy flow, concrete and brine plays a medium role and as indicated ground, header and pump heat loads are added to the “above ice” loads through this path.

Figure 40: Heat flow from ice to refrigeration plant

Three temperature differences are considered to be calculated; evaporator-brine, average brine in and out temperatures- bottom ice and bottom-top ice temperature difference. Table 8 summarizes the results for March 15, 2011 in the Norrtälje ice rink. As it is shown the major heat transfer resistance is in the heat transfer from brine to evaporator. The reason may be related to the difference in the mechanism of the heat transfer and the actual surfaces used for the heat transfer. In the evaporator convection occurs while for the two other situations conduction is the prevailing heat transfer mechanism. The solution to have less temperature resistance might be a heat exchanger with more heat transfer area and to not be dependent on evaporator superheat control. Table 8: Heat transfer resistances and temperature differences

Average resistance (K/kW)

average ΔT (K)

evaporator-brine

0.0400

5.63 (LMTD)

Avg. brine-bottom ice

0.0168

1.97

bottom-top ice

0.0041

0.47

It should be mentioned that for calculating the first temperature difference, log mean temperature difference is used as in evaporator the two-phase refrigerant temperature is almost constant while the brine temperature decreases. Looking at the resistances and differences, it can be concluded that two third of the total resistance resulted from the evaporator-brines while other shares are 27% and 7%.

55


5. CONCLUSION Methods to analyse the refrigeration system performance have been studied by using the ClimaCheck instrumentation plus the compressor heat rejection calculation. Furthermore, ice and indoor hall conditions are studied to understand the driving factors for the heat loads. The results of the measurements show that the total energy consumption in Norrtälje is about two third of Älta. The main reasons for this less energy consumption are smarter control for compressors and pumps, better ventilation distribution design and 1°C-2°C higher ice temperature. Methods to estimate the heat loads and their shares in the ice rinks total heat load are introduced. These estimates are useful to select the proper energy efficiency measures. Analytical modelling for a sample day has estimated that about 84% of the total heat loads is originated from the heat loads on ice sheet while the distribution system causes the remaining 16%. Moreover, calculations show that convection plus small portion of condensation (altogether 36%), radiation (23%), ice resurfacing (14%) and lighting (7%) are the largest heat loads in winter while in summer condensation is another significant heat load (10%). Comparing two six-hour periods, one without ice resurfacing and four resurfacings in the second one, 30% more cooling demand has been calculated for the second period. A heat flux method to calculate the heat transfer in the ice is introduced. It has been shown that with the help of two temperature sensors embedded in ice, it is possible to have an estimation of the heat loads on the ice. Moreover, the effect of ice temperature on the “heat absorbing” process is discussed. Finding the heat transfer resistances, it has been shown that evaporator to brine contributes with 66% of the heat transfer resistances from ice to evaporator while brine to bottom ice and bottom to top ice accounts for 27% and 7% respectively. To conclude, a parallel “performance analysis of the refrigeration system” and “heat loads estimation” proves to be a useful tool for adopting proper design and control for energy efficient operation.

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6. FUTURE WORK / SUGGESTIONS 6.1 Future work This thesis was part of a greater study for “energy efficiency in ice rinks”. There are some more researches which are in process or will be started in future including: 

  

A paper has been presented in 23rd IIR International Congress of Refrigeration titled: “Experimental cooling load analysis of ice rinks”. Another paper can be prepared to reflect the latest results. Energy Usage Prediction model comparing Indoor vs. Outdoor Ice Rinks Indoor ice rink more detailed temperature measurements including the ceiling temperature and ice surface by IR camera Research on ice physical properties and its impacts on the total heat load

6.2 Suggestions There are some ideas that may worth thinking: 





 

In Älta ice rink a low-e ceiling is installed during summer 2011. It can be a valuable study to compare the energy consumptions before and after the installation and it can show the significant effect of this technique (or its weakness). As radiation and convection are the “two big” heat loads in ice rinks, it seems that more detailed calculations and modelling is required. For example for radiation, a better knowledge of ceiling temperature and its ɛ value is necessary. Furthermore, the ceiling shape by ASHRAE (and in this study) is assumed to be flat while in reality, many ceilings are curved and a broader study is required on this topic. For convection, the heat transfer coefficient is taken from a paper of 1970’s and it seems that better correlations should be derived by new researches. Ice rinks are mostly used in northern countries with cold winters including North American, Russia and Scandinavian countries. There could be some systems to use the “very cold outdoor air”, for example during midnights, to help the refrigeration system works with a lower capacity. Covering the ice during nights can lower the radiation and convection, probably to a great extent. The advantages and disadvantages of such a system can be studied. Body heat losses of open compressors and the input electricity losses are sources of uncertainty in cooling capacity calculations and they are not easy to obtain. More research and accuracy is necessary for ClimaCheck evaluation of refrigeration systems with open compressors.

57


7. BIBLIOGRAPGY • • • • • • • •

• • • • • • • • •

• •

AAQ, 2003, Technical Fact Sheets on the Impacts of New Energy Efficiency Technologies and Measures in Ice Rinks, The Association des Arenas du Québec ASHRAE. 2010, ASHRAE Refrigeration Handbook, chapter 44:ice rinks, ANSI/ASHRAE standard Bellache et al. 2007, Calculation of refrigeration loads by convection, radiation and condensation in ice rinks using a transient 3D zonal model Berglöf K. 2010, Optimization through performance inspections, Proc. Stockholm Conference, IIF/IIR. Çengel, Yunus.A et al. 2007, Fundamentals of Thermal-Fluid Sciences, Chapter 22: Radiation heat transfer, McGraw-Hill DOE. 1980, energy conservation in ice skating rinks, prepared by B.K.Dietrich and T.J.MacAvoy, U.S. Department of energy, Washington, D.C. Everything Ice, 2000, Recreation Facilities Design Manual, publish authorized by “Manitoba Hydro”, Canada FAO, 2003, The use of ice on small fishing vessels, food and agriculture organization of the united nations, Rome, 2003, available at http://www.fao.org/docrep/006/Y5013E/y5013e00.htm#Contents Granryd E. 2005, Refrigerating Engineering, Part II, Chapter 15: Processes in moist air, frosting and defrosting, KTH, Department of Energy Technology How ice rink works?, 2011, available at http://entertainment.howstuffworks.com/ice-rink2.htm IIHF, 2010. Technical guidelines of an ice rink, international ice hockey federation guide book, chapter 3 Ingvar, 2007, Ice skating rink-movable, available at http://www.ingvar.is/Plants/IceSkatingRink/IceSkatingRink.html Luminous. 2011, Wikipedia: Luminous Efficacy available at: http://en.wikipedia.org/wiki/Luminous_efficacy Makhnatch P. 2010, Technology and Energy Inventory of Ice Rinks, Master thesis, KTH Melinder Å. 2009, Handbook on indirect refrigeration and heat pump systems, Effsys2 program, KTH Motor challenge, 2010, determining electric motor load and efficiency, US department of Energy Munters, 2011, Munters DryCoolTM Dehumidification Systems, Engineering Catalogue available at: http://webdh.munters.com/webdh/BrochureUploads/Engineering%20Catalog%20DDS.pdf Nicolas L. 2009, Improving efficiency in ice hockey arenas, ASHRAE Journal June 2009 Retscreen, 2005, Energy Efficiency Project Analysis for Supermarkets and Arenas, presentation for “Clean energy project analysis course”, Canada 58

Measurement and Modelling of Ice Rinks Heat Loads • • • • •

Rogstam J. (a) 2010, Energy usage statistics and saving potential in ice rinks, Proc. Stockholm Conference, IIF/IIR. Rogstam J. (b) 2010, Ice rinks using carbon dioxide as secondary refrigerant, Proc. Stockholm Conference, IIF/IIR. Rogstam J. (c)/Karampour M. 2011, Experimental cooling load analysis of ice rinks, The 23rd IIR International Congress of Refrigeration, Prague Sawalha S. and Chen Y. 2010, Investigations of Heat Recovery in Different Refrigeration System Solutions in Supermarkets, Effsys2 Project final report, KTH Shahzad K. 2006, an ice rink refrigeration system based on CO2 as secondary fluid in copper tubes, Master Thesis, IUC/KTH.

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8. APPENDIX To bring a clearer vision of some important calculations, in sections 8.1 and 8.2 two sample calculations are shown. The method to estimate the percentage of heat rejected from compressor body in comparison to the shaft power is described in section 8.1. Furthermore, in section 8.2 the calculation to find the cooling capacity is shown for one hour, as a sample. The last section shows some figures of the two ice rinks including their refrigeration, heating, ventilation and dehumidification systems.

8.1 Compressor heat rejection sample calculations The heat that is rejected from the compressor is estimated by using MYCOM software. It is a compressor selection and analysis software. The interface, as shown in Figure 41, includes the refrigerant type, thermodynamics of the cycle, compressor specifications and calculation results. The compressors used in Norrtälje are “K” series with 8 cylinders. Various RPMs and condensing temperatures have been studied (refer to section 3.2.2) to find a general correlation for the heat rejection.

Figure 41: MYCOM software interface

60

Measurement and Modelling of Ice Rinks Heat Loads Table 9: Heat rejection calculation results by MYCOM software

RPM=1400

RPM=1300

RPM=1200

RPM=1100

RPM=1000

RPM=900

teva = -10°C, superheat = 7K, subcool = 5K tcond (°C) Cooling Capacity Absorbed Power Oil Heat Rejection Jacket Heat Rejection Total heat Rejection (Tot. Heat/abs. Power)% Cooling Capacity Absorbed Power Oil Heat Rejection Jacket Heat Rejection Total heat Rejection (Tot. Heat/abs. Power)% Cooling Capacity Absorbed Power Oil Heat Rejection Jacket Heat Rejection Total heat Rejection (Tot. Heat/abs. Power)% Cooling Capacity Absorbed Power Oil Heat Rejection Jacket Heat Rejection Total heat Rejection (Tot. Heat/abs. Power)% Cooling Capacity Absorbed Power Oil Heat Rejection Jacket Heat Rejection Total heat Rejection (Tot. Heat/abs. Power)% Cooling Capacity Absorbed Power Oil Heat Rejection Jacket Heat Rejection Total heat Rejection (Tot. Heat/abs. Power)%

kW kW kW kW kW kW kW kW kW kW kW kW kW kW kW kW kW kW kW kW kW kW kW kW kW kW kW kW kW kW

10 109.9 12.4

15 104.5 14.6

20 99 16.6

25 93.3 18.5

30 87.5 20.3

35 81.5 21.8

40 75.3 23.2

0.58 1.58 2.16 17.42 122.1 14

0.58 2.13 2.71 18.56 116.2 16.4

0.58 2.69 3.27 19.70 110.2 18.7

0.58 3.24 3.82 20.65 103.9 20.8

0.58 3.79 4.37 21.53 97.5 22.7

0.58 4.33 4.91 22.52 91 24.5

0.58 4.88 5.46 23.53 84.2 26

0.58 1.79 2.37 16.93 134.2 15.7

0.58 2.35 2.93 17.87 127.8 18.4

0.58 2.9 3.48 18.61 121.3 20.8

0.58 3.46 4.04 19.42 114.5 23.2

0.58 4.02 4.6 20.26 107.6 25.3

0.58 4.57 5.15 21.02 100.5 27.2

0.74 5.13 5.87 22.58 93.2 28.9

0.58 1.97 2.55 16.24 146.1 17.5

0.58 2.54 3.12 16.96 139.3 20.3

0.58 3.1 3.68 17.69 132.3 23

0.58 3.67 4.25 18.32 125.1 25.5

0.58 4.24 4.82 19.05 117.6 27.9

0.69 4.8 5.49 20.18 110 30

0.91 5.37 6.28 21.73 102.2 31.8

0.58 2.13 2.71 15.49 158 19.3

0.58 2.71 3.29 16.21 150.7 22.3

0.58 3.29 3.87 16.83 143.2 25.2

0.58 3.87 4.45 17.45 135.5 27.9

0.64 4.44 5.08 18.21 127.6 30.5

0.86 5.02 5.88 19.60 119.5 32.7

1.09 5.6 6.69 21.04 111.1 34.7

0.58 2.27 2.85 14.77 169.7 21.1

0.58 2.86 3.44 15.43 162 24.4

0.58 3.45 4.03 15.99 154 27.5

0.58 4.05 4.63 16.59 145.9 30.4

0.81 4.64 5.45 17.87 137.5 33.1

1.04 5.23 6.27 19.17 128.9 35.5

1.28 5.82 7.1 20.46 120.1 37.7

0.58 2.38 2.96 14.03

0.58 3 3.58 14.67

0.58 3.61 4.19 15.24

0.74 4.22 4.96 16.32

0.99 4.83 5.82 17.58

1.23 5.44 6.67 18.79

1.48 6.05 7.53 19.97

61

RPM=1800

RPM=1700

RPM=1600

RPM=1500

Measurement and Modelling of Ice Rinks Heat Loads Cooling Capacity Absorbed Power Oil Heat Rejection Jacket Heat Rejection Total heat Rejection (Tot. Heat/abs. Power)% Cooling Capacity Absorbed Power Oil Heat Rejection Jacket Heat Rejection Total heat Rejection (Tot. Heat/abs. Power)% Cooling Capacity Absorbed Power Oil Heat Rejection Jacket Heat Rejection Total heat Rejection (Tot. Heat/abs. Power)% Cooling Capacity Absorbed Power Oil Heat Rejection Jacket Heat Rejection Total heat Rejection (Tot. Heat/abs. Power)%

kW kW kW kW kW kW kW kW kW kW kW kW kW kW kW kW kW kW kW kW

181.1 22.9

173.1 26.4

164.7 29.8

156.1 32.9

147.3 35.7

138.2 38.4

129 40.7

0.58 2.48 3.06 13.36 192.4 24.8

0.58 3.11 3.69 13.98 183.9 28.6

0.65 3.75 4.4 14.77 175.2 32.1

0.92 4.38 5.3 16.11 166.2 35.4

1.17 5.01 6.18 17.31 156.9 38.4

1.43 5.64 7.07 18.41 147.4 41.2

1.69 6.27 7.96 19.56 137.8 43.7

0.58 2.56 3.14 12.7 203.4 26.8

0.58 3.22 3.8 13.3 194.6 30.7

0.82 3.87 4.69 14.6 185.5 34.4

1.1 4.53 5.63 15.9 176.1 37.9

1.37 5.18 6.55 17.1 166.4 41.1

1.65 5.84 7.49 18.2 156.5 44.1

1.92 6.49 8.41 19.2 146.5 46.7

0.58 2.62 3.2 11.94 214.2 28.8

0.71 3.3 4.01 13.06 205 32.9

1 3.99 4.99 14.51 195.5 36.8

1.29 4.67 5.96 15.73 185.8 40.5

1.58 5.35 6.93 16.86 175.8 43.9

1.87 6.03 7.9 17.91 165.5 47

2.15 6.71 8.86 18.97 155 49.8

0.58 2.66 3.24 11.25

0.88 3.38 4.26 12.95

1.19 4.09 5.28 14.35

1.49 4.8 6.29 15.53

1.79 5.51 7.3 16.63

2.1 6.23 8.33 17.72

2.4 6.93 9.33 18.73

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8.2 Cooling capacity sample calculations Figure 42 shows the cooling capacity calculation for the last hour of March 15, 2011. Day

time

2011-03-15 2011-03-15 2011-03-15 2011-03-15 2011-03-15 2011-03-15 2011-03-15 2011-03-15 2011-03-15 2011-03-15 2011-03-15 2011-03-15 2011-03-15 2011-03-15 2011-03-15 2011-03-15 2011-03-15 2011-03-15 2011-03-15 2011-03-15 2011-03-15 2011-03-15 2011-03-15 2011-03-15 2011-03-15 2011-03-15 2011-03-15 2011-03-15 2011-03-15 2011-03-15 2011-03-15 2011-03-15 2011-03-15 2011-03-15 2011-03-15 2011-03-15 2011-03-15 2011-03-15 2011-03-15 2011-03-15 2011-03-15 2011-03-15 2011-03-15 2011-03-15 2011-03-15 2011-03-15 2011-03-15 2011-03-15 2011-03-15 2011-03-15 2011-03-15 2011-03-15 2011-03-15 2011-03-15 2011-03-15 2011-03-15 2011-03-15 2011-03-15

23:59:00 23:58:00 23:57:00 23:56:00 23:55:00 23:54:00 23:53:00 23:52:00 23:51:00 23:50:00 23:49:00 23:48:00 23:47:00 23:46:00 23:45:00 23:44:00 23:43:00 23:42:00 23:41:00 23:39:00 23:38:00 23:36:00 23:35:00 23:34:00 23:33:00 23:32:00 23:31:00 23:30:00 23:29:00 23:28:00 23:27:00 23:26:00 23:25:00 23:24:00 23:23:00 23:22:00 23:21:00 23:20:00 23:19:00 23:18:00 23:17:00 23:16:00 23:15:00 23:14:00 23:13:00 23:12:00 23:11:00 23:10:00 23:09:00 23:08:00 23:07:00 23:06:00 23:05:00 23:04:00 23:03:00 23:02:00 23:01:00 23:00:00

Electric Enthalpy Enthalpy ref comp ref exp in ref comp Hp Lp t_cond Power ExpVal in Comp in in (°C) (°C) out (°C) (KPa_a) (KPa_a) (°C) (kW) (kJ/kg) (kJ/kg) -3,9 8,4 79,6 886 273,3 37,7 21,03 239,25 1467,72 -3,9 8,4 79,5 886,3 273,2 37,8 21,04 239,25 1467,73 -3,9 8,4 79,5 886 273 38,1 21,03 239,25 1467,75 -3,9 8,4 79,4 886,6 273,2 38 21,05 239,25 1467,73 -3,9 8,4 79,4 886,6 273,1 38,2 21,05 239,25 1467,74 -3,9 8,3 79,4 886 273 37,7 21,03 238,78 1467,75 -3,9 8,3 79,4 885,7 273 37,8 21,02 238,78 1467,75 -3,9 8,4 79,4 886 273 37,7 21,03 239,25 1467,75 -3,8 8,4 79,4 885,7 273 37,5 21,02 239,25 1468,00 -3,8 8,4 79,4 886,8 273,1 37,9 21,06 239,25 1467,99 -3,8 8,4 79,5 887,1 273,2 37,7 21,07 239,25 1467,98 -3,8 8,4 79,5 888,2 273,4 37,6 21,11 239,25 1467,96 -3,8 8,4 79,5 887,4 273,7 37,6 21,08 239,25 1467,93 -3,8 8,4 79,5 887,1 273,5 37,9 21,07 239,25 1467,95 -3,8 8,4 79,6 886,6 273,4 38,1 21,05 239,25 1467,96 -3,8 8,4 79,6 886,3 273,4 38 21,04 239,25 1467,96 -3,7 8,4 79,6 886,3 273,6 37,7 21,04 239,25 1468,18 -3,7 8,4 79,8 886,3 273,4 38 21,04 239,25 1468,20 -3,7 8,4 79,9 886 273,5 37,9 21,03 239,25 1468,19 -3,7 8,4 79,9 886 273,1 37,8 21,03 239,25 1468,23 -3,7 8,4 79,9 886,6 273,3 38,2 21,05 239,25 1468,21 -3,7 8,4 79,8 886,8 273,6 37,7 21,06 239,25 1468,18 -3,8 8,5 79,8 887,4 273,5 37,7 21,08 239,72 1467,95 -3,8 8,5 79,8 888,2 274 37,9 21,11 239,72 1467,90 -3,8 8,5 79,8 888,2 273,4 37,8 21,11 239,72 1467,96 -3,8 8,5 79,8 888,5 273,8 37,9 21,12 239,72 1467,92 -3,8 8,5 79,8 888,8 273,7 38,1 21,13 239,72 1467,93 -3,8 8,5 79,8 889,3 273,6 37,7 21,14 239,72 1467,94 -3,8 8,5 79,7 889,3 273,9 37,8 21,14 239,72 1467,91 -3,8 8,5 79,7 889,3 273,6 38,3 21,14 239,72 1467,94 -3,8 8,6 79,7 890,1 274 37,8 21,17 240,18 1467,90 -3,8 8,5 79,7 891 274 37,9 21,20 239,72 1467,90 -3,8 8,5 79,7 890,1 273,8 38 21,17 239,72 1467,92 -3,8 8,6 79,7 890,4 273,8 37,9 21,18 240,18 1467,92 -3,8 8,6 79,8 890,4 274,1 38,1 21,18 240,18 1467,89 -3,7 8,7 79,8 891,5 273,9 37,8 21,22 240,65 1468,15 -3,7 8,7 79,9 892,1 274,1 37,8 21,24 240,65 1468,13 -3,7 8,7 79,9 892,3 274,4 38,1 21,25 240,65 1468,10 -3,7 8,7 79,9 893,4 274,5 38,2 21,29 240,65 1468,10 -3,7 8,8 80 894,3 274,5 38,2 21,32 241,12 1468,10 -3,7 8,8 80 894,8 274,9 38 21,34 241,12 1468,06 -3,7 8,7 80 894,3 274,6 38,2 21,32 240,65 1468,09 -3,7 8,6 80 893,4 274,6 38,3 21,29 240,19 1468,09 -3,7 8,6 80,1 889,9 274,4 37,8 21,17 240,18 1468,10 -3,7 8,6 80,2 890,4 274,1 38 21,18 240,18 1468,13 -3,6 8,6 80,1 890,4 274,1 38,3 21,18 240,18 1468,38 -3,7 8,6 80,1 891 274,5 38 21,20 240,18 1468,10 -3,7 8,6 80,1 890,7 274,6 38,2 21,19 240,18 1468,09 -3,7 8,6 80,1 891 274,1 37,6 21,20 240,18 1468,13 -3,7 8,6 80,1 890,4 274,1 37,6 21,18 240,18 1468,13 -3,7 8,6 80,1 891,8 274,4 38,3 21,23 240,19 1468,10 -3,7 8,6 80 891,2 274,1 38,3 21,21 240,18 1468,13 -3,7 8,6 80 891,5 274,5 38,1 21,22 240,19 1468,10 -3,7 8,6 80 891,8 274,3 38,3 21,23 240,19 1468,11 -3,7 8,6 80 891,8 274,3 37,9 21,23 240,19 1468,11 -3,7 8,7 79,9 892,3 274,4 37,7 21,25 240,65 1468,10 -3,7 8,7 79,8 893,2 274,6 37,8 21,28 240,65 1468,09 -3,7 8,7 79,8 894,3 274,6 37,9 21,32 240,65 1468,09

Enthalpy Q_rejection Comp out (kW) (kJ/kg) 1637,06 7,08 1636,80 7,10 1636,81 7,16 1636,54 7,14 1636,54 7,18 1636,56 7,08 1636,57 7,10 1636,56 7,08 1636,57 7,04 1636,53 7,12 1636,77 7,08 1636,72 7,07 1636,75 7,07 1636,77 7,12 1637,03 7,16 1637,04 7,14 1637,04 7,08 1637,54 7,14 1637,80 7,12 1637,80 7,10 1637,77 7,18 1637,52 7,08 1637,50 7,09 1637,47 7,12 1637,47 7,11 1637,46 7,13 1637,44 7,16 1637,42 7,09 1637,18 7,11 1637,18 7,20 1637,15 7,11 1637,11 7,13 1637,15 7,15 1637,14 7,13 1637,38 7,17 1637,34 7,11 1637,57 7,12 1637,56 7,17 1637,52 7,19 1637,73 7,20 1637,71 7,16 1637,73 7,20 1637,76 7,21 1638,14 7,11 1638,37 7,15 1638,13 7,21 1638,10 7,15 1638,11 7,19 1638,10 7,07 1638,13 7,07 1638,07 7,21 1637,85 7,21 1637,84 7,17 1637,82 7,21 1637,82 7,13 1637,56 7,10 1637,28 7,12 1637,23 7,14

m_dot (kg/s)

Q_cooling (kW)

0,1525 0,1532 0,1544 0,1542 0,1550 0,1530 0,1534 0,1530 0,1524 0,1541 0,1530 0,1526 0,1526 0,1538 0,1544 0,1540 0,1530 0,1537 0,1531 0,1527 0,1543 0,1525 0,1523 0,1531 0,1528 0,1531 0,1540 0,1524 0,1530 0,1550 0,1530 0,1534 0,1538 0,1534 0,1540 0,1530 0,1528 0,1540 0,1544 0,1542 0,1533 0,1542 0,1545 0,1523 0,1529 0,1545 0,1531 0,1539 0,1515 0,1515 0,1543 0,1545 0,1537 0,1545 0,1529 0,1523 0,1530 0,1534

187,39 188,19 189,70 189,47 190,48 188,05 188,54 187,98 187,29 189,30 188,02 187,52 187,46 188,98 189,70 189,19 187,98 188,95 188,16 187,71 189,69 187,45 187,09 188,04 187,62 188,08 189,09 187,13 187,86 190,38 187,80 188,39 188,89 188,33 189,01 187,81 187,52 188,98 189,49 189,17 188,14 189,22 189,77 186,96 187,72 189,79 187,96 188,93 186,03 186,02 189,48 189,78 188,74 189,77 187,79 187,00 187,76 188,29

Figure 42: Cooling capacity calculation sample for 23:00-23:59 on March 15, 2011

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8.3 Norrtälje and Älta ice rinks photo gallery

Figure 43: Flooded evaporator (right) and brine pumps (left) - Älta

Figure 44: Coolant pumps and condenser (left corner) Älta

Figure 45: DX evaporator (right), desuperheater (top-left) and condenser (bottom-left) - Norrtälje

Figure 46: refrigeration system for outdoor bandy field Norrtälje

Figure 47: heat recovery pump for ventilation - Älta

Figure 48: District heating system - Norrtälje

64


Figure 49: desiccant wheel dehumidification and ventilation heat recovery unit - Älta

Figure 50: dehumidification piping (right) and heating/cooling coils (left) in ventilation ducts - Norrtälje

Figure 51: ClimaCheck central control unit - Älta

Figure 52: Instruments for over ice temperature and humidity measurements- Älta

Figure 53: ventilation ducts and lighting - Älta

Figure 54: ice resurfacing machine - Älta

65