HUGO DELLO SBARBA
HEAT RECOVERY SYSTEMS IN UNDERGROUND MINE VENTILATION SYSTEMS AND NOVEL MINE COOLING SYSTEMS
Mémoire présenté à la Faculté des études supérieures et postdoctorales de l’Université Laval dans le cadre du programme de maîtrise en génie des mines pour l’obtention du grade de Maître ès sciences (M.Sc.)
DÉPARTEMENT DE GÉNIE DES MINES, DE LA MÉTALLURGIE ET DES MATÉRIAUX FACULTÉ DES SCIENCES ET DE GÉNIE UNIVERSITÉ LAVAL QUÉBEC
2012 ©Hugo Dello Sbarba
Résumé L’exploitation minière souterraine dans les régions froides du monde nécessite le chauffage de l’air frais de ventilation et des bâtiments de surface. L’air vicié est habituellement rejeté dans l'atmosphère à des températures beaucoup plus élevées que l'air ambiant. Un logiciel informatique a été développé afin d'évaluer la faisabilité de récupérer la chaleur de l'air vicié des mines. Le logiciel estime la quantité de chaleur d’air vicié récupérable dans une mine souterraine. Il déterminera ensuite les économies annuelles potentiels d'énergie et un coût capital du système pour obtenir le retour sur l’investissement initial. Le logiciel considère un circuit de glycol en boucle fermée avec des échangeurs de chaleur à tubes et ailettes situées à l'extrémité des installations de ventilations à la surface (à l’entrée et l’échappement d’air). Différents concepts des systèmes de récupération de chaleur sont énoncés. La plupart des sources de chaleurs habituelles trouvées sur un site minier sont répertoriés. Quelques concepts innovateurs qui exploitent le froid de l'hiver comme un atout pour refroidir l'air d'entrée sont exposés. Mots clés : Sources de chaleurs, air vicié, récupération de chaleur, faisabilité, chauffage, refroidissement
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Abstract Underground mining in cold regions of the world requires heating of surface buildings and intake fresh air. Exhaust return air is usually discharged to the atmosphere at much higher temperatures than the ambient air. A computer software application has been developed in order to evaluate the feasibility of recovering heat from return exhaust air. The software approximates the amount of heat that can be recovered on surface from the exhaust ventilation shaft of an underground mine. It will then determine the annual energy cost savings and a capital cost of the system. This software considers a closed-loop glycol circuit with tube and fins heat exchangers located at the extremity of the exhaust and intake shaft surface installations. Different concepts of the heat recovery system are as well described. Most common heat sources that can be found on mine sites are listed. Several innovative designs that exploit cold winter weather as an asset to cool mine intake air are explained. Key words: heat sources, return air, heat recovery, feasibility, heating, cooling
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Acknowledgements This master’s thesis could have never been completed without the help of the following individuals and industries. I am grateful to everyone that helped me get through this incredible journey that I will never forget. I would first like to thank CEMI the Centre for Excellence in Mining Innovations which funded this research project and have always supported me for which ever needs I had. Their dedication for supporting researchers in the mining world is overwhelming and I will always be thankful for the opportunity they gave me. I would like to thank my Director, Dr. Fytas, who has always helped, taught and guided me in the right direction to complete this project. Most of all I recognize the efforts and his caring to prepare me for my future career which I already begun and which I will enjoy for the years to come. I would like to thank my co-director, Dr. Paraszczak, who always had his door open for any help I needed and also guided me to make this project what it is now. Fajnie było I would like to thank my dear friend, Georges Bedros, who helped in key aspects of my project. I would like to thank the following people and industries for giving me their precious time to make this project possible.
Nick Newman and Bill Thomas from Industrial Heat Transfer Inc.
Charles Gagnon from Genivar
Édith Lafontaine, Christian Quirion and Rosaire Émond from Agnico-Eagle
Stéphane Dubois from Wesdome
Jérôme Massonat from Thermofin
Alexandre Lacasse from LysAirMecanic
Charles Kade from SNC-Lavallin
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Contents CHAPTER 1.INTRODUCTION .................................................................................................... 1 1.01
Overview ...................................................................................................................... 1
1.02
Underground mine ventilation ...................................................................................... 2
1.03
Heat sources in underground mines: ............................................................................ 3
1.04
Exhaust air heat recovery ............................................................................................. 4
1.05
Heating in underground mines ..................................................................................... 5
1.06
Thermodynamics of underground mine air for cold periods ........................................ 6
1.07
Closed loop glycol circuit ........................................................................................... 10
CHAPTER 2.EXISTING UNDERGROUND MINE HEAT RECOVERY PROJECTS AND STUDIES ...................................................................................................................................... 12 2.01
Introduction ................................................................................................................ 12
2.02
Heat recovery from abandoned mines ........................................................................ 12
2.03
Exhaust air heat recovery studies ............................................................................... 13
2.04
Existing heat recovery projects................................................................................... 14
2.04.1 Heat recovery system from mine water .................................................................. 14 2.04.2 Heat recovery from mine air compressors .............................................................. 14 2.04.3 Creighton mine heat recovery system ..................................................................... 14 2.04.4 Strathcona mine heat recovery system .................................................................... 14 2.04.5 Kiena mine exhaust air heat recovery system ......................................................... 15 2.04.6 Williams mine heat recovery system ...................................................................... 15 2.05
Conclusions ................................................................................................................ 16
CHAPTER 3.PROBLEMATIC AND RESEARCH OBJECTIVES ............................................ 17 3.01
Introduction ................................................................................................................ 17
3.02
Evaluating the economics of exhaust air heat recovery system ................................. 17
3.03
Novel cooling systems ................................................................................................ 19
3.04
Research objectives .................................................................................................... 20
CHAPTER 4.FEASIBILITY STUDY SOFTWARE; ENERGY CALCULATIONS ................. 21 4.01
Introduction ................................................................................................................ 21
4.02
Calculation of heat capacity rate of air at exhaust ...................................................... 22
4.03
Predicting air conditions at HE outlet (humidity and dry bulb temperature) ............. 23
4.03.1 First case: Dry cooling of exhaust air ..................................................................... 24 4.03.2 Second case: Cooling and dehumidifying of air ..................................................... 24 4.04
Calculation of HE efficiency ...................................................................................... 27 vi
4.05
Calculation of mean wall temperature ........................................................................ 30
4.06
Calculation of the fluids mean temperatures .............................................................. 30
4.07
Glycol temperature above water freezing point ......................................................... 31
4.08
Calculation of nominal pipe size diameter ................................................................. 31
4.09
Pressure drop across piping system ............................................................................ 32
4.10
Calculation of additional fan power cost .................................................................... 35
4.11
Calculation of energy savings per year....................................................................... 36
4.12
Calculation of maximum pipe heat loss to surroundings ........................................... 37
4.13
Calculation examples for heat losses .......................................................................... 37
4.14
Calculation of mass flow rate of condensate .............................................................. 39
4.15
Ethylene glycol mixture thermophysical properties ................................................... 39
4.16
Conclusions ................................................................................................................ 39
CHAPTER 5.CAPITAL COST AND DESIGN CONSIDERATIONS ....................................... 41 5.01
Introduction ................................................................................................................ 41
5.02
Heat exchangers and its installations .......................................................................... 43
5.03
Cost of the tube and fin heat exchangers .................................................................... 44
5.04
Manifolds .................................................................................................................... 44
5.05
Ventilation Building extension cost ........................................................................... 46
5.05.1 Foundations for coil supports and walls ................................................................. 49 5.06
Building ...................................................................................................................... 50
5.06.1 Slab on grade for the HE building. ......................................................................... 50 5.06.2 Insulation of building .............................................................................................. 50 5.06.3 Coils support ........................................................................................................... 50 5.06.4 Coils arrangement ................................................................................................... 50 5.07 5.07.1
Main piping system .................................................................................................... 52 Underground piping ................................................................................................ 52
5.07.2 Pipe supports: .......................................................................................................... 54 5.07.3 Pipe insulation......................................................................................................... 56 5.08
Pumps and electric motor ........................................................................................... 57
5.08.1 Pump ....................................................................................................................... 57 5.08.2 MCC and motor feeder ........................................................................................... 58 5.09 5.09.1
Piping accessories ....................................................................................................... 59 Strainer .................................................................................................................... 60
5.09.2 Air bleed lines ......................................................................................................... 60 vii
5.09.3 Reducer ................................................................................................................... 60 5.09.4 Tee........................................................................................................................... 60 5.09.5 Check valve ............................................................................................................. 60 5.09.6 Butterfly valves ....................................................................................................... 60 5.09.7 Gate valve ............................................................................................................... 61 5.09.8 Flange ...................................................................................................................... 61 5.09.9 Expansion tank ........................................................................................................ 61 5.09.10
Expansion joints .................................................................................................. 62
5.10
Ethylene glycol ........................................................................................................... 63
5.11
Bypass valve ............................................................................................................... 63
5.12
Automated Washing System ...................................................................................... 64
5.12.1 Cost and geometry of the flexible hose and nozzles ............................................... 65 5.12.2 Piping system of spray nozzles branches ................................................................ 66 5.12.3 Pump and electric motor ......................................................................................... 66 5.12.4 Valves ..................................................................................................................... 67 5.12.5 Trench for Piping from main water supply to exhaust ........................................... 67 5.12.6 Piping from main water supply to exhaust ............................................................. 67 5.13
Other systems not included......................................................................................... 67
5.14
Economy of the project............................................................................................... 68
5.15
Case studies ................................................................................................................ 70
5.16
Conclusions ................................................................................................................ 73
CHAPTER 6.ALTERNATIVE DESIGNS OF THE HEAT RECOVERY SYSTEM ................. 74 6.01
Introduction ................................................................................................................ 74
6.02
Recovering heat from the depths of the mine ............................................................. 74
6.03
Heat pump, evaporator at exhaust and condenser at intake ........................................ 75
6.04 Spray chambers at exhaust, tube-fin HE at intake, plate heat exchangers to transfer the heat.. .................................................................................................................................... 75 6.04.1 Direct contact HE .................................................................................................... 76 6.04.2 Filtration system...................................................................................................... 79 6.04.3 Plate Heat Exchanger .............................................................................................. 80 6.05 Heat pump from a refrigeration plant; Direct-contact HE at exhaust, glycol tube and fin HE at intake ......................................................................................................................... 82 6.06
Re-circulation of return air ......................................................................................... 83
6.07
Heat sources other than exhaust mine air ................................................................... 83
6.07.1 Mine water heat recovery........................................................................................ 83 viii
6.07.2 Heat recovery of mine air compressors .................................................................. 84 6.07.3 Geothermal ground heat pump................................................................................ 84 6.07.4 Recovering heat from tailings ................................................................................. 84 6.08
Heating the surface buildings ..................................................................................... 85
6.09
Conclusions ................................................................................................................ 88
CHAPTER 7.TUBE AND FIN HE TECHNOLOGY; SOFTWARE FOR HEAT EXCHANGER DESIGN ........................................................................................................................................ 90 7.01
Introduction ................................................................................................................ 90
7.02
Geometrical parameters calculations .......................................................................... 92
7.03
Air conditions calculations ......................................................................................... 95
7.04
Inlet water variables.................................................................................................... 95
7.05
Fluid properties and velocity calculations .................................................................. 97
7.06
Heat transfer calculations ........................................................................................... 97
7.06.1 j factors correlations ................................................................................................ 98 7.07
HE efficiency design study ....................................................................................... 101
7.07.1 Air face velocity effect on the efficiency .............................................................. 102 7.07.2 Air face velocity effect on the pressure drop ........................................................ 103 7.08
Conclusions .............................................................................................................. 104
CHAPTER 8.MEANS OF REDUCING THE ADVERSE EFFECTS OF ADIABATIC COMPRESSION (EXCLUDING NATURAL AND MECHANICAL COOLING) ................. 105 8.01
Introduction .............................................................................................................. 105
8.02
Use of turbines instead of regulators ........................................................................ 108
8.03
Brattice wall .............................................................................................................. 110
8.04
Conclusion ................................................................................................................ 110
CHAPTER 9.NOVEL COOLING SYSTEMS; APPLICATIONS TO CANADIAN MINES .. 111 9.01
Introduction .............................................................................................................. 111
9.02
Vapour compression cycle........................................................................................ 111
9.02.1 Ideal cycle ............................................................................................................. 112 9.02.2 Actual vapour compression cycle ......................................................................... 114 9.03
New design proposal: producing work from the refrigeration plant ........................ 116
9.04
Questioning the use of surface air cooling for Canadian mines ............................... 116
9.05 New design proposal: Closed-loop glycol circuit for refrigeration plants on surface at sub-zero glycol temperatures .................................................................................................. 117 9.06
Natural heating and cooling system (NHEA) ........................................................... 117
9.07
Ice stopes .................................................................................................................. 118 ix
9.08
Ice conveying to the underground levels .................................................................. 119
9.09 New design proposal: surface ice formation, ice storage for ice conveying to the underground levels .................................................................................................................. 119 9.10
Conclusions .............................................................................................................. 121
CHAPTER 10.GENERAL CONCLUSIONS AND RECOMMENDATIONS FOR FURTHER WORK ........................................................................................................................................ 123 10.01
Overview .................................................................................................................. 123
10.02
Main goals and objectives ........................................................................................ 123
10.03
Recommendations for further work .......................................................................... 124
10.03.1
Energy calculations and tube and fin HE: ......................................................... 124
10.03.2
Capital cost approximation and design recommendations ................................ 125
10.03.3
Alternative designs ............................................................................................ 125
10.03.4
Cooling .............................................................................................................. 125
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List of Tables Table 4-1: IHT Inc. data HE data for condensation ...................................................................... 28 Table 4-2: HE efficiency with respect to relative humidity.......................................................... 28 Table 4-3: Nominal pipe size ........................................................................................................ 32 Table 5-1: Couplings material cost and labour hours with respect to NPS .................................. 42 Table 5-2: Manifold system components ...................................................................................... 45 Table 5-3: Assumed depth of trench with respect to the NPS diameter ....................................... 53 Table 5-4: Depth of bedding with respect to NPS ........................................................................ 54 Table 5-5: Maximum span of pipe supports ................................................................................. 56 Table 5-6: Rated head chosen with respect to flow rate for pumps in series ................................ 58 Table 5-7: Total linear thermal expansion for carbon steel pipes ................................................. 63
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List of Figures Figure 1-1: Laronde mine ventilation installations ......................................................................... 2 Figure 1-2: Heat sources in deep underground mines .................................................................... 3 Figure 1-3: Geothermal gradients for different regions of the world ............................................. 4 Figure 1-4: Ice block accumulations in intake raise ....................................................................... 5 Figure 1-5: Intake air burners ......................................................................................................... 6 Figure 1-6: Mine ventilation system heat gain ............................................................................... 7 Figure 1-7: Air enthalpy as it flows through underground mine workings during cold periods .... 7 Figure 1-8: Psychrometric chart, cooling and dehumidifying process ........................................... 9 Figure 1-9: Cross flow tube and fin HE .......................................................................................... 9 Figure 1-10: Heat recovery system schematic closed loop glycol circuit. ................................... 10 Figure 2-1: Heat pump coupling with an inactive mine shaft ....................................................... 13 Figure 2-2: Conventional heat pipe............................................................................................... 14 Figure 3-1: Canadian average natural gas and propane price since 2001 ..................................... 18 Figure 4-1: First case: Dry cooling of exhaust air ........................................................................ 24 Figure 4-2: Cooling and dehumidifying of exhaust air ................................................................. 24 Figure 4-4: Outlet conditions predicted above saturation line ...................................................... 25 Figure 4-3: Outlet conditions predicted below saturation line...................................................... 25 Figure 4-5: HE efficiency with respect to relative humidity ........................................................ 29 Figure 4-6: Friction fanno flows ................................................................................................... 33 Figure 4-7: Tee of manifolds ........................................................................................................ 33 Figure 5-1: Manifold schematic .................................................................................................... 45 Figure 5-2: IPS hole cut ................................................................................................................ 46 Figure 5-3: Surface building extension ......................................................................................... 46 Figure 5-4: Surface building extension building top and side view ............................................. 47 Figure 5-5: Exhaust ventilation collar at surface .......................................................................... 48 Figure 5-6: Diffuser efficiency ..................................................................................................... 49 Figure 5-7: Assumed building shape for cost estimation, top view .............................................. 50 Figure 5-8: Airflow bypassed at the exhaust building, top view .................................................. 51 Figure 5-9: Train wheels for coils support .................................................................................... 51 Figure 5-10 : Pipe support schematic............................................................................................ 55 Figure 5-11: Adjustable saddle with stanchion..............................................................................55 Figure 5-12: Strainer tee type ....................................................................................................... 60 Figure 5-13: Bypass valve ............................................................................................................ 64 Figure 5-14: Automated washing system schematic .................................................................... 65 Figure 5-15: flexible hose for automated washing system ........................................................... 65 Figure 5-16: Cost distribution of components case 1 ................................................................... 69 Figure 5-17: Cost distribution of components case 2....................................................................69 Figure 5-18: Cost distribution of components case 3 ................................................................... 69 Figure 5-19: Cost distribution of components case 4....................................................................69 Figure 5-20: Cost distribution of components case 5 ................................................................... 69 Figure 5-21: Cost distribution of components case 6....................................................................69 Figure 5-22: Cost distribution of components case 7 ................................................................... 69 Figure 5-23: Cost distribution of components case 8....................................................................69 Figure 5-24: Cost distribution of components case 9 ................................................................... 70 Figure 5-25: Heat recovery system economics, case 1 ................................................................. 71 xii
Figure 5-26: Heat recovery system economics, case 2..................................................................71 Figure 5-27: Heat recovery system economics, case 3 ................................................................. 72 Figure 5-28: Heat recovery system economics, case 4..................................................................71 Figure 5-29: Heat recovery system economics, case 5 ................................................................. 72 Figure 5-30: Heat recovery system economics, case 6..................................................................72 Figure 5-31: Heat recovery system economics, case 7 ................................................................. 72 Figure 5-32: Heat recovery system economics, case 8..................................................................72 Figure 5-33: Heat recovery system economics, case 9..................................................................72 Figure 6-1: Schematic of heat recovery system from the depths of the mine............................... 75 Figure 6-2: Diagram of heat recovery system with plate heat exchanger..................................... 76 Figure 6-3: Crossflow horizontal spray chambers: low water loading ......................................... 77 Figure 6-4: Schematic of two-stage cross flow horizontal spray high water loading ................... 78 Figure 6-5: Schematic of filtration system.................................................................................... 80 Figure 6-6: Plate heat exchanger................................................................................................... 80 Figure 6-7: Heat recovery system with the use of a refrigeration plant........................................ 82 Figure 6-8: Fresh air heat demand greater or equal than total heat recovered, no building heating ....................................................................................................................................................... 86 Figure 6-9: Fresh air heat demand lower than total heat recovered, building heating from remaining heat recovered .............................................................................................................. 87 Figure 6-10: No fresh air heat demand, portion of the total heat recovered for building heating, building heat demand fully satisfied ............................................................................................. 87 Figure 6-11: Bypass for building heating ..................................................................................... 87 Figure 6-12: Bypass for building heating for no intake air heating .............................................. 88 Figure 7-1: Longitudinal fins heat exchanger ............................................................................... 92 Figure 7-3: Longitudinal and transversal tube pitch ..................................................................... 93 Figure 7-2: Tube arrangements ..................................................................................................... 92 Figure 7-4: HE efficiency vs Face air speed ............................................................................... 102 Figure 7-5: Pressure drop vs Face air speed ............................................................................... 104 Figure 8-1: Compressor adding positive work to air .................................................................. 105 Figure 8-2: Change in pressure with negative altitude ............................................................... 106 Figure 8-3: Temperature change with depth, polytropic equations and linear relationship ....... 107 Figure 8-4: Turbine coupled with generator with variable load for flow regulation .................. 109 Figure 9-1 Schematic of the vapour compression-cycle ............................................................. 112 Figure 9-2: Ammonia vapour ln(P)-h diagram ........................................................................... 114 Figure 9-3: Schematic of vapour compression cycle with state points ....................................... 114 Figure 9-4: Temperature-enthalpy diagram of cycle with heat exchange .................................. 115 Figure 9-5: Diagram of natural heating system .......................................................................... 118 Figure 9-6: Diagram of ice stopes ............................................................................................... 119 Figure 9-7: Surface ice formation for heating, ice storage for ice conveying to the underground levels ........................................................................................................................................... 120
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CHAPTER 1. INTRODUCTION Summary This Chapter describes usual mine ventilation systems, heat sources in underground mines, heating of the fresh air and an introduction to exhaust air heat recovery.
1.01
Overview
Energy systems in underground mines are very important as they can significantly increase production cost and therefore affect the economics of a mining project. In most mining operations, ventilation cost is an important part of mining cost, but as underground mines become deeper air cooling systems are required to increase safety and comfort at deep levels. Air cooling systems must be optimized as their operating and capital costs are elevated. In Northern regions mines, there are additional challenges as heating ventilation fresh air is required for a significant portion of the year. Heating cost for a usual shallow mine (< 1000 m) can be over a million dollar per year. These additional energy systems are as important as the ventilation and must be optimized to minimize energy costs as much as possible. In this thesis, heating and cooling applications for Northern operations will be covered. New ideas as well as existing designs will be described. Economics of a heat recovery system is explained into detail and a software to perform a feasibility study is made available. The first chapter describes usual mine ventilation systems, heat sources in underground mines, heating of the fresh air and an introduction to exhaust air heat recovery. Chapter two lists existing heat recovery projects and studies related to underground mines. Chapter three presents the research objectives and the problematic of this thesis. Chapter four describes the energy calculations performed in the feasibility software of the closed-loop glycol circuit return air heat recovery system. Chapter five describes the detailed capital cost calculation of each of the components of the closed-loop glycol circuit. Furthermore, chapter five outlines the design considerations and recommendations of the system. Chapter six presents designs of heat recovery systems other than the closed-loop glycol circuit.
©Hugo Dello Sbarba
Chapter seven outlines the calculations of the tube and fin heat exchanger design performances. A design software tool is also available. Chapter eight explains the effect of adiabatic compression and means of reducing its adverse effects other than mechanical or natural cooling. Chapter nine describes existing and novel cooling system designs with some recommendations for implementation in Northern climates.
1.02
Underground mine ventilation
Ventilation in underground mines is essential since both personnel and mining equipment require oxygen to operate. Fresh air is carried to the underground levels with the use of mechanical driven fans usually located on surface. The fans are driven by electric motors. Fresh air is required to provide oxygen but also to dilute the dusts and gases within the mine that are discharged by the different activities that a mining operation involves such as diesel equipment and blasting. There are several mine regulations regarding ventilation to ensure the comfort and safety of the workers. In order to maintain a flow rate of fresh air across the workings; an inlet and outlet are connected to the atmosphere. The inlet is called the downcast ventilation shaft (or intake shaft) and the outlet is the upcast ventilation shaft (or exhaust shaft). On surface, building installations are usually in place to accommodate the fans at the extremity of the upcast and downcast shaft. The ventilation building installations of the Laronde mine located in Cadillac, Quebec, Canada are shown in Figure 1-1. The fans can be centrifugal or axial, in both cases, an elbow will be required to re-direct the airflow as it reaches surface. There are several configurations of fans: a pull system indicates that the fans are located downstream of the upcast shaft and a push system that the fans are located upstream of the downcast shaft. There also exist push and pull systems in which fans are located at both inlet and outlet. The electricity cost of the ventilation system can usually represent from 30 to 50% of the total electricity cost of the mine (Fytas, 2007). The significant energy cost in underground mining is a major issue, energy systems must be optimized to ensure the sustainability of underground mining operations as energy prices will keep on rising.
Figure 1-1: Laronde mine ventilation installations (Fytas, 2007)
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1.03
Heat sources in underground mines:
There are several heating sources that will heat the air underground. For a typical deep mine (2000 m depth), the usual partition of the heat sources is as shown in the diagram of Figure 1-2.
Figure 1-2: Heat sources in deep underground mines (Hartman, 1982)
The temperature of the virgin rocks as they are freshly broken will depend on their depth and the geothermal gradient of the region. The geothermal gradient is given in terms of rock virgin temperature per length of depth as shown in Figure 1-3. The heat generated will also depend on the thermal properties of the rock. Another major heat source comes from the underground equipment, diesel vehicles and machinery that will discharge a large amount of heat due to their relatively low efficiencies (around 33%). If electrical equipment is used, the heat generated will be lower as their engines are more efficient. Mine water can also contribute to the heat sources. As the air comes in contact with the warm water, the air will gain some of the heat from convection and can also transfer some of its sensible heat into latent heat i.e. as the water will evaporate the dry bulb temperature will decrease without changing the actual specific heat of air. Auto compression is the transformation of potential energy into thermal energy. As the air is carried to lower underground levels, its specific heat increases, thus for a same amount of air there is a larger amount of heat than at surface. This occurrence will increase the temperature of air of approximately 1°C per 100 m of depth. The opposite will occur as the air is carried back up to surface. Other smaller sources of heat will include human metabolism and explosives. These heat sources can be a significant problem since the mine may require the use of cooling plants at significant capital and operating costs. In Canada, there are presently two deep mines that require air cooling during the warm periods, cooling systems will be explained more into depth in Chapter 8. The heat generated inside the mine will result in a much greater return air temperature than the fresh air. The deeper the mine, the greater will be the temperature difference between the return and fresh air.
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Figure 1-3: Geothermal gradients for different regions of the world (Hartman, 1982)
1.04
Exhaust air heat recovery
As mentioned above, during cold periods, return exhaust air is usually discharged to the atmosphere at much greater temperatures than the intake fresh air and consequently ambient air. Using a medium such as water, the heat can be extracted from the exhaust air and discharged at the desired location. In some mines, heat recovery could contribute to large energy savings. There are also other potentially recoverable heat sources at underground mine sites. They include warm mine water and heat from mine air compressor coolers. These can contribute to the heat recovery system and should always be taken into consideration to combine with the exhaust air heat recovery; however return air will usually represent the largest portion of the recoverable heat. Heat recovery systems should be more and more considered to offset some of the difficulties encountered with the increasing price of energy. There are presently at least two known exhaust air heat recovery systems and both projects have been known to be successful. Depending on the mine ventilation systems and installations, the most efficient heat recovery option from one mine site to another can differ, it is thus important to be aware and understand all the possibilities available for these types of projects in underground mines. 4
1.05
Heating in underground mines
In cold regions of the world, intake fresh air usually requires to be heated above 0°C. If air at sub-zero temperature enters the mine shaft, it can create severe ice build up along the rock walls of the intake shaft causing an increase in airflow resistance and can eventually fully block the airflow as shown in Figure 1-4. Air is usually heated with gas-fired heaters at surface as shown in Figure 1-5. The amount of heat required depends on the air flow rate and ambient temperature. The temperature is also increased to enhance the comfort of workers. Air is commonly heated with the use of propane or natural gas. Natural gas is used when a gas pipeline is located at proximity to the mine site. In remote areas, propane has to be carried and stored in large tanks. Usually when a mine uses propane, its heating bill is much greater than for those using natural gas. In Canadian underground mines, the heating cost for a small shallow mine can be over one million dollar per year. With the increasing price of hydrocarbons fuel, heating cost becomes more and more significant. At the Laronde mine located in Cadillac, Quebec, during the cold period, ventilation air is heated on surface at 1.5 °C and as the air goes to the lower levels it requires to be cooled, otherwise the lower levels are too hot for the workers. This mine is the first one in Canada facing such a controversial condition; heating the intake fresh air and then requiring cooling this same air at lower levels.
Figure 1-4: Ice block accumulations in intake raise (Gagnon, 2011)
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Figure 1-5: Intake air burners (Gagnon, 2011)
1.06
Thermodynamics of underground mine air for cold periods
The following will explain in detail the thermodynamic process of the air as it flows from the atmosphere to the underground workings to be finally discharged into the atmosphere. It should be noted that the thermodynamic process will be qualified with enthalpy as opposed to air dry bulb temperature since it does not reflect the actual change in energy; the humidity increase must be taken into consideration as it is often a large portion of the actual total energy increase in underground mines. The explanation is demonstrated with the schematic of Figure 1-6 and it correlates with the graph of Figure 1-7 demonstrating the change in enthalpy as mine air flows inside the mine.
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Figure 1-6: Mine ventilation system heat gain
Figure 1-7: Air enthalpy as it flows through underground mine workings during cold periods
As the ventilation air enters the intake building the gas fired heaters discharge a large quantity of heat. The fan will also increase the air temperature. Secondly as the air flows through the intake raise, it gains heat from rock walls and auto compression. The amount of heat discharged from the rock walls will depend on its temperature. With time, the rock wall temperature will decrease. If its temperature is greater than the air temperature, heat will be transferred from the strata to the air. In the opposite case, heat will be transferred from the air to the strata and thus air cooling will occur. Then as the air reaches the underground levels, the heat from mining activity is discharged; virgin rock walls and diesel equipment will generate the major part of the heat. The air enthalpy will achieve its highest peak; that is for the deepest level. Then as foul air will flow back to the exhaust raise, de-compression of air will occur and the air enthalpy will decrease linearly due to this effect. The heat transfer from rock walls to air is dependant again on the air 7
and rock wall temperatures. As air will have a greater temperature than at the intake raise there is a higher tendency to have air cooling occurring. The total heat gain as shown in Figure 1-7 will be much greater during the winter than summer as the intake air will have a lower temperature and thus heat transferred from the rocks walls and virgin rocks will be greater due to the higher temperature difference. A cross-flow tube-fin heat exchanger can be implemented within the existing upcast shaft surface building in order to extract the heat. The main consideration with a tube-fin heat exchanger is that the dirty exhaust mine air will cause heavy fouling on the air-side of the HE (Heat Exchanger). The dust accumulation will reduce the heat transfer rate of the HE and can eventually block the airway. One solution is the use of an automated washing system that frequently cleans the heat exchanger (Emond, 2009). Another way to avoid fouling on the airside would be to use a direct contact heat exchanger such as spray chambers or towers. These direct contact heat exchangers are already being used for air cooling and heat rejection of underground and surface refrigeration plants in deep mines. One disadvantage is that the water acts as an air cleaner, as a result contamination of water will continuously increase. Water contamination must be limited in order to avoid fouling on the water-side of the heat exchanger at the refrigeration plant. It is important to note that the heat exchanger will create a greater pressure drop across the ventilation system requiring the fan to deliver more power; this increase in fan power must be included in the operating costs. It is also important to ensure that the gas content within the air will not react with the heat exchanger material. It is the case of potash mines in which the nature of potash dust is very corrosive. Uncommon material must then be used to build the HEs which significantly increases the project’s capital cost (Hall, Mchaina, & Hardcastle, 1990). Exhaust mine air usually has high relative humidity from the evaporation of mine water. Due to the latter, as air cools down across the exchanger, it reaches the saturation point and humidity condenses; this process is called air cooling and dehumidifying as shown in the psychrometric chart of Figure 1-8. The vertical line indicates the heat load transferred from the water vapour condensation (latent heat) and the horizontal line represents the heat load transferred through convection (sensible heat). The latent heat is very important to consider in the energy calculations since it can sometimes contribute to more than 50% of the total heat transferred when relative humidity is close to 100%. The condensation will create a film of water on the heat exchanger which could increase the overall thermal resistance of the system thus decreasing the heat transfer rate. Such film must be taken into consideration when designing the heat exchanger. To avoid freezing of water vapour condensation, the heat exchanger working fluid temperature must always be kept above 0°C. To evacuate all of the condensate, a water drainage system must be installed.
8
Figure 1-8: Psychrometric chart, cooling and dehumidifying process
Since the largest part of the heating cost of a mine site will be for intake fresh air heating; the energy savings would be greater if the recovered heat is discharged at this location. Heating the intake fresh air can be done using a crossflow fin and tube heat exchanger with a running liquid, a schematic and a picture of a crossflow tube and fin HE are shown in Figure 1-9. The heat exchanger should be installed on the surface ahead of the gas-fired heaters to maximize heat transfer rate between the working fluid and ambient air. Since the latter is relatively clean, no corrosion or dust accumulation problems should be encountered.
Figure 1-9: Cross flow tube and fin HE (Jun-Jie & Wen-Quan, 2005)
The heat recovered could also be used to heat surface buildings. In this case, the use of a heat pump would be required unless the heat is extracted from a high grade heat source. It can also be possible to combine a heat recovery system with intake fresh air and surface building heating. When the intake fresh air would not require the full heating load recovered, the remaining load could be discharged in the surface building.
9
The air temperature at exhaust does not fluctuate much during the cold period of the year therefore the heat load recovered should remain relatively constant (Lafontaine, 2008) unless there is a large change in mining activities such as a complete stop in production for a long period of time.
1.07
Closed loop glycol circuit
This system involves two tube and fin HEs in crossflow arrangement, one located in the exhaust building to recover the heat, the second one located in the intake building to discharge the heat. An insulated piping system is in place to carry the water and ethylene glycol mixture around the loop. A schematic of the closed-loop glycol circuit is shown in Figure 1-10.
Figure 1-10: Heat recovery system schematic closed loop glycol circuit. Note: All of these components are located on surface
Ethylene glycol mixture is used so that the fluid does not freeze when flowing across the intake HE. As mentioned earlier, at exhaust the glycol mix should remain above the freezing point. The efficiency of the HE of the intake shaft and the amount of heat recovered can determine the minimum temperature that the ambient air can achieve in order to avoid that the glycol temperature does fall below the freezing point. Otherwise a bypass valve can be used. For the usual Canadian underground mine conditions, closed loop glycol circuit design is the most efficient and feasible. A software has been built in which a closed-loop glycol system is simulated to obtain the energy savings per year as well as the approximate capital cost of the system. The closed-loop glycol system will be discussed more into detail in Chapter 4 and 5. 10
Chapter 5 covers technical considerations in the design and building of each of the components of the heat recovery system.
11
CHAPTER 2. EXISTING UNDERGROUND MINE HEAT RECOVERY PROJECTS AND STUDIES Summary This Chapter lists existing heat recovery projects and studies related to underground mines.
2.01
Introduction
Heat recovery projects in underground mines have existed for a long time. However it is expected that more and more mines will take advantage of the available heat sources of the mine site as heating costs are elevated. The following describes some studies and existing heat recovery projects. All of the existing heat recovery projects are located in Canada. Existing projects use the three following heat sources; compressor coolers, mine water and return air, these systems will be outlined. This Chapter will also summarize heat recovery projects and studies for abandoned mines.
2.02
Heat recovery from abandoned mines
Recovering wasted heat from underground mines has been realised using the water from flooded abandoned mines. It is presently carried out in the former coal mines in Springhill Nova Scotia (Jessop, Jack, Macdonald, & Spence, 1995). A Ph.D. student of Université Laval, Jasmin Raymond, worked on the feasibility of using water from abandoned flooded mines to heat buildings at the industrial park in Murdochville, Gaspe (Raymond, 2006). A study has been carried out at Virginia Tech to verify the feasibility of using warm air from abandoned mines to heat greenhouses (Marsh & Singh, 1994). In the study, it was discovered that having the mine air directly channelled to the greenhouses would diminish the growth of the plants due to its high relative humidity (almost 100%). Water vapour inside the greenhouse would condensate on the walls and reduce light transmission to the plants. On the other hand, mine air usually has higher content of CO2 which enhances the growth of the plants. After evaluating the proposal of having mine air directly in contact with the plants it was found that an air-to-air heat exchanger would be more effective. In this case, solely thermal energy is recovered whereas no mine air enters the greenhouse. Figure 2-1 shows the concept with explanation of the components. The design proposed was evaluated to be feasible but it has not yet implemented in practice. The large amount of heat carried by the exhaust air could make this project profitable depending on the needs of greenhouses in the region. In the Abitibi region in the town of Guyenne, several greenhouses are operating all-year round and are at the proximity of several underground mines.
12
Figure 2-1: Heat pump coupling with an inactive mine shaft (A) Evaporator side of heat pump which absorbs heat energy from the mine air. (B) Vent which is open during heat pump operation to exhaust mine air and closed during direct mine air cooling. (C) Vent in greenhouse side wall. This vent is closed during heating and opened for direct mine air cooling. (D) Condenser side of heat pump. A fan (not shown) moves air past condenser coils for greenhouse heating. (E) Perforated polyethylene tube used to distribute warm air the length of the greenhouse. (F) Vents to exhaust ventilation air during direct mine air cooling operation. Closed for heating. (Marsh & Singh, 1994)
2.03
Exhaust air heat recovery studies
A feasibility study of heat recovery in Canadian Potash mines has been performed by the University of British Columbia,Vancouver, BC (Hall, Mchaina, & Hardcastle, 1990). One of the systems evaluated was to use the exhaust air to heat the intake air during cold periods. The heat would be transferred using a medium like water or glycol and a pump would carry the liquid from the intake to the outlet. The major problem was that due to the corrosive nature of potash dust the coils at the exhaust would have to be made of a corrosive resistant material such as plastic. This design was proved to be unsuccessful due to several operational problems. The feasibility of controlled recirculation of air was studied at the Rocanville division of Potash Corporation of Saskatchewan (Hall, Mchaina, & Hardcastle, 1990). Potash being a relatively soft rock; its excavation uses a minimum of blasting and has low utilization of diesel units which results in low contaminants and dust level of air. Air recirculation was experimented by opening an existing ventilation door between the intake and return airways. Exhaust air would then be bypassed through the opening and returned to the intake. The level of contaminants and dust and the temperature and flow of air were monitored during different working shifts. The quality of air in the mine was found to remain stable. The possibility of re-circulating air makes it possible to reduce the flow of air driven into the intake thus reducing the fan power. Also, since there is a lower flow of air, a smaller amount of heat is required to warm the intake air. The results of the experiments showed that recirculation of air was feasible in the case of the Rocanville division. A scientific article mentions the possibility of using heat pipes between mine fresh air and return air (Joy, 1978). A conventional heat pipe consists of a sealed pipe with an inside wick and a working fluid as shown in Figure 2-2. As the working fluid absorbs thermal energy it evaporates and migrates inside the cavity to subsequently flow towards the lower temperature end of the 13
pipe. When the working fluid reaches the cold end it condenses and gets absorbed by the wick and then flows towards the hot end of the pipe. The working fluid chosen depends on the operating temperatures of the application. Heat pipes are efficient to transfer heat and require no maintenance. Unfortunately the use of a heat pipe is only possible if the inlet and exhaust are very close to each other. If a mine has such a layout the heat pipe would most probably be the best option for heat recovery.
Figure 2-2: Conventional heat pipe
2.04
Existing heat recovery projects
2.04.1 Heat recovery system from mine water The Macassa mine located in Kirkland, Ontario, used heat from mine water and air compressor coolers to heat the intake air. A portion of the mine water was bypassed and gained heat by cooling hot glycol from the compressors cooling system. The totality of the warmer mine water then transferred its heat to another glycol loop which was connected to a tube and fin HE which would finally transfer its heat to mine intake air. The cold mine water was then sent to the disposal pond (Ruiter, 1992). In Finland the Pyhäsalmi zinc-copper mine uses mine water heat recovery to heat its 150 m3/s ventilation fresh air in winter (Pyhäsalmi, 2010). 2.04.2 Heat recovery from mine air compressors The following mines recovered heat from the mine air compressors, the maximum power recovered is included; Kidd Creek (5.86 MW), Strathcona (2.93 MW) and Lockerby mine (1.46 MW) all located in Northern Ontario (Sylvestre, 1999). The heat recovery system of the Kidd Creek mine is similar to the one at the Macassa mine. 2.04.3 Creighton mine heat recovery system The first known mine exhaust air heat recovery system was implemented in 1955 at Inco’s Creighton mine in Sudbury Ontario. The system used a direct contact HE at exhaust and a tube and fin HE at intake. The hot water heated from the exhaust air was carried in the coils at intake to heat the intake air. The system would recover in average a total power of 1.5 MW. The system ceased to operate in 1970 due to difficulties of maintaining the proper operating conditions (Sylvestre, 1999). 2.04.4 Strathcona mine heat recovery system The Stratchona mine located in northern Ontario implemented a heat recovery system in 1968. The system would recover heat from the exhaust air and the compressor cooling water. The system would recover a maximum power of 8.8 MW from exhaust air and 2.93 MW from 14
compressors coolers. After eight winters of operation, the exhaust coils had to be replaced (McCallum, 1969). The type of coils was brass tubes and copper fins, the cease of its operation is suspected to be the result of lack in coating. The coating to protect the coils is very important to extend the life of the coils as it was applied in the two following existing projects. 2.04.5 Kiena mine exhaust air heat recovery system In 1987, the Kiena mine located in Dubuisson, Quebec, implemented a similar design; closedloop glycol circuit transferring heat from return air and discharging it in the intake fresh air with the use of closed-loop glycol system. The total project capital cost was 760 000$ CAN. In 1988, the total annual savings were estimated to be 137 000$ CAN (Emond, 1988). The system is considered to be successful and is still presently running (Dubois, 2009). 2.04.6 Williams mine heat recovery system A feasibility study of recovering heat from several sources has been carried out by V.B. Cook Co. Limited for the Williams mine situated in Hemlo area of Northwestern Ontario (Smith & Arthur, 1996). The heat recovery is performed on exhaust air, mine water discharge and mine compressor’s inter-coolers and after-coolers. The feasibility study concluded that $500,000 per year gross energy savings could be achieved with a capital cost of $1,700,000. The project was approved based on a 3.8 years payback period and implemented in 1995. The system is still presently running (Pelletier, 2009). The design is a closed-loop glycol circuit which is explained more into detail in Chapters 4 and 5. The Williams mine system has a different configuration for the summer and winter. The heat recovery system will be described in detail as it is an exemplary successful heat recovery project that utilizes all available heat sources. The difference between the winter and summer configuration is as follows: Winter: The glycol runs through the different heat sources; ventilation exhaust air, mine water discharge and compressor inter-coolers and after coolers to be finally discharged at the mine ventilation intake. Since the mine water and ventilation exhaust air are low-grade heat sources they are positioned immediately downstream of the heat rejection points in order to maximize the temperature differential available for heat recovery from these low temperature sources. Summer: The flow direction is the same and the intake heating coils are bypassed. The compressors heat is rejected at the mine water and exhaust air coils since they become the lowest temperature available in the system. The different parts of the system are described below: Air intake The coils in the intake have their own fans so when heating is not required the air intake flow can bypass the coils reducing the pressure drop in the summer. The fans were part of the original ventilation design so they were not included as additional costs. A temperature controlled valve has been installed between the supply and return headers of the intake coils. It is used to ensure that the exhaust temperature of the pipes stays above 0°C to avoid freezing of the condensate on the pipes. The temperature controlled valves are programmed to ensure a minimum downstream 15
temperature of 1°C at the exhaust coils. The amount of flow to be bypassed depends on the total energy balance of the system. During the summer, the valve is fully opened to reduce pressure drop across the system. Exhaust The coils recovering the heat are installed in a steel framed structure attached to the discharge of the fan diffusers. In this case, the coils are similar to the ones used in the intake but in a larger number due to the larger volume of air at the exhaust. A spray cleaning automated system has been installed. The cleaning system is equipped with a soap tank, hot water tank and is programmed to clean the coils 3 times a day. The exhaust air is at 100% relative humidity and water condensates on the coils in large amounts. This water and the wash water are collected in a sump and pumped to the sedimentation pond. Compressors A pressure reducing valve is installed between the compressor glycol supply and return headers to maintain a predetermined pressure differential across the compressors. The compressor cooling demand should always be less than the total system capacity. In order to control the cooling rate, a controlled amount of glycol can bypass the compressors to achieve the desired cooling. The glycol pump is located in the compressor room. The pump provides different amount of power depending on where the system is operating on the summer or winter configuration. This option of the pump reduces energy consumption in the summer due to the lower head encountered by the system. Mine Water Heat Exchanger The mine water discharge line is connected to a plate heat exchanger. Some adjustments had to be made so that the flow of water stays constant for the whole day. The pump was previously operating twenty hours a day. A glycol bypass valve is provided to allow the exchanger to be disassembled for servicing during the summer period. After one year of operation the system had met its expectations with savings of up to $500,000. It is important to note that in this application the site does not have access to natural gas and therefore more expensive propane fuel has to be used which significantly increases the cost of heating and makes it more profitable to recover the heat.
2.05
Conclusions
Existing heat recovery studies and projects demonstrate that the potential for heat recovery systems is great and several underground mine sites would benefit from it. The first two exhaust air heat recovery projects (Creighton and Strathcona) have been found to have operational problems and had to stop their operation before the end of the mine life. The two recent projects (Kiena and Williams) have been found to be more than successful and are still in operation today. These two operations earned savings of several millions of dollars and have largely benefited from their initial investment. It is therefore a proof that these systems can be reliable if properly designed. Heat recovery from compressors has also found to be successful and should always be evaluated.
16
CHAPTER 3. PROBLEMATIC AND RESEARCH OBJECTIVES Summary This Chapter presents the research objectives and the problematic of this thesis.
3.01
Introduction
Every underground mine must use proper ventilation systems in order to provide adequate quantity of fresh air according to mine regulations. The objective is to provide miners with safe and healthy atmosphere, by reducing health hazards and improving working conditions. Fresh air is indispensable to provide enough oxygen, to dilute and evacuate exhaust gases generated by diesel engines, as well as other contaminants such as DPM (Diesel Particulate Matters), respirable dust or noxious gases associated with chemical composition of ore and waste. Furthermore, in deep mines, where the heat emanating from virgin rocks becomes an issue, mine ventilation systems are also used to control the air temperature underground.
3.02 Evaluating the economics of exhaust air heat recovery system Many Canadian mines are situated in a climate where winter is very harsh with low temperatures reaching down to -15°C and lower. Sending very cold air to the underground levels can have serious consequences; Excessive cooling of workspaces (discomfort for miners affecting their health and productivity). Accumulation of ice within the shaft or freezing of service water. It is then indispensable for most mines to heat intake air during winter in order to reach temperatures above the freezing point. The Canadian mining industry has to face the ever increasing energy prices in the future. Natural gas is commonly used to heat ventilation intake fresh air. On the other hand where natural gas pipelines are not available, propane is used which will usually increase significantly the heating bill. Figure 3-1 shows the trend of the average Canadian natural gas and propane prices since 2001 (NRCAN, 2011).
17
Canadian average natural gas and propane price since 2001 $30
$25
price ($/GJ)
$20
Propane Natural gas
$15
$10
$5
$0 2001
2002
2003
2005
2006
2007
2009
2010
2011
year
Figure 3-1: Canadian average natural gas and propane price since 2001
As opposed to natural gas, propane prices have been significantly increasing in the last two decades. It is important to note that the natural gas price will often differ from one region to another, for example, presently natural gas price in the region of Sudbury, Ontario is 8$/GJ (Sabau, 2010) while in Abitibi Quebec it will come up to 13$/GJ (Girard, 2010). The fuel price greatly affects the energy costs of the mine. For example, for a mine in a climate such as Abitibi, Qc and a ventilation flow rate of 400 m3/s; the annual total heat demand will be of approximately 120,000 GJ. At 13$/GJ the heating cost is 1,560,000$ and at 8$/GJ it is of 960,000$. This cost is even higher for mines in the far North such as Raglan located in Northern Quebec or the Diavik Diamond mine located in the Northwest Territories. Since average underground mine temperatures are stable around 15 to 20°C, the foul exhaust air leaving the underground workings in winter has a much higher temperature than the intake air. This relatively warm air is discharged into the atmosphere and the heat it transports is lost. Underground mines in Canada have to minimize the effect of the rise in energy prices on their mining costs. Optimizing the energy systems of the mine will ensure a more competitive operation. In northern regions of the world, it is possible to optimize the efficiency of the heating system by recovering heat from different sources and discharge it to the intake fresh air. Several types of heat sources can be found on a mine site such as mine water and compressor coolers but the most common one with a significant potential is ventilation exhaust air. The known mine sites that had chosen to install an exhaust air heat recovery system or perform a feasibility study have decided to do so for the simple reason that the exhaust and intake ventilation shafts were located relatively close to each other. These mine sites were presented in the previous Chapter. It seems that the decision to study the feasibility of installing an exhaust air heat recovery system was somehow arbitrary and that many additional mines should have done so but have come to quick conclusions that it would not be feasible simply by looking at the distance between intake and exhaust shafts. It seems that there is a lack of resources and research in energy calculation 18
and feasibility study for exhaust air heat recovery system. The distance between intake and exhaust shafts will affect the economics of the heat recovery system but there are in fact many other factors that will influence it. The practical and economical issues associated with heat recovery from mine exhaust air have not been studied profoundly as indicated in the literature review (Chapters 1 and 2). It is then important to investigate these aspects in a more detailed manner and develop the tools that will assist mine operators with decisions over the applicability and benefits of such systems. It was therefore decided to develop a software tool that could be used to evaluate the feasibility of installing a heat recovery system. It would be suitable for most underground mine sites. The software tool uses some parameters of the mine operation and calculates instantaneously the return of the investment of the heat recovery system. It is composed of two major parts which are the energy and capital cost calculations. One of the main components of the heat recovery system is the tube and fin HE. Its design will determine the pressure drop induced on the ventilation air as well as the amount of heat transferred at exhaust or intake. It was decided that accurate results of the performance of the heat exchanger was necessary. In order to do so, a tube and fin heat exchanger design software tool has been built. This tool can be used for system optimization.
3.03
Novel cooling systems
Deep underground mines generate large amounts of heat in many ways such as high virgin rock temperature, diesel equipment and autocompression. The depth and altitude above or below sea level will greatly affect the mine working areas ambient temperatures as heat from virgin rock and autocompression are both dependent on them. Mining at critical depths can therefore lead to the exposure of miners to hot environments which can become a safety hazard. In Canada, regulations are in place to minimize the working load when workers are exposed to such environment. Hot working areas will therefore affect safety and productivity of the mine. In some cases, fresh air flow from the ventilation system is just not sufficient to maintain acceptable temperatures; air cooling is therefore required which will greatly affect the mining cost. As more orebodies with great potential are being discovered at increasing depths; more mines are exposed to heat issues. Due to the latter, the number of mines equipped with air cooling systems will increase in the future as deep mining becomes more and more interesting. There are several ways of cooling underground mine ventilation air and there has been a significant amount of research performed on this subject around the world. Unfortunately most of this research results cannot be directly applied to Canadian mines as the climate is much cooler than at usual underground mine sites equipped with cooling plants. So far there are only two known Canadian mines equipped with mechanical cooling plants; Kidd Creek and Laronde. Their capital and operating costs constitute a significant proportion of mining cost. It is therefore important that they are as efficient as possible. Innovative new ideas are described as they could lead to a complete change in the way that cooling plants are designed in cold climate regions. The large temperature difference between winter and summer can be used as an advantage to reduce significantly cooling energy costs. Several innovative designs that exploit cold winter weather as an asset to cool mine intake air have been found to be very successful in the past and should always be taken into consideration. 19
3.04
Research objectives
The main objectives of the research are as follows: Research existing heat recovery systems or under study projects in underground mines. Investigate the available designs of these systems. Choose and study the most efficient and feasible design. Evaluate the cost of these designs. Develop a software tool to evaluate the feasibility of the heat recovery system design in mines. List recommendations and novel design proposals for mechanical cooling systems in underground mines. Research on the theory of adiabatic compression and how to reduce its adverse effects. Develop new ideas for mine cooling other than mechanical or natural cooling.
20
CHAPTER 4. FEASIBILITY STUDY SOFTWARE; ENERGY CALCULATIONS Summary \ This Chapter describes the energy calculations performed in the feasibility software of the closed-loop glycol circuit return air heat recovery system, the assumptions involved are explained. The theory on the heat transfer occurring within the system is as well described.
4.01
Introduction
The feasibility study software of the heat recovery system calculates annual energy savings and operating costs as well as an approximate capital cost of the system to finally obtain a payback period of the installation of the closed-loop glycol circuit. The closed-loop glycol circuit is briefly described in section 1.06 of Chapter 1, more detailed information on the design is included in Chapter 5. This Chapter contains the explanation of the energy cost savings and operating costs calculations. It was chosen to develop such a software in order to determine a payback period of implementing a heat recovery system at any given mine site. The software uses several input data from the mine site that are relatively easy to find. The variables required are as follow;
Distance between intake and exhaust Volumetric flow rate of air at exhaust and intake Dry and wet bulb temperature at exhaust Altitude of exhaust surface installations Minimum intake air temperature Outlet temperature of glycol at intake Average temperature of each month Fuel price Burners efficiency Price of electricity Fan operating efficiency Fan electric motor efficiency Maximum glycol velocity within piping system Exhaust HE efficiency under dry conditions
The software is developed around a spreadsheet, each of the variables and outputs are clearly identified. Several functions using VBA programming language have been built within the software. The spreadsheet form enables the user to better track the results obtained from the calculations. It also facilitates the user to modify any part of the software very quickly.
21
Nomenclature
A C cp
2
Area (m ) -1
Flow stream heat capacity rate (kW °C ) Specific heat of fluid at constant pressure -1 -1 (kJ kg K )
di
Pipe inside diameter (m)
f h
Friction factor, dimensionless Head loss (kPa)
T
Temperature (°C)
Tdb
Dry bulb temperature (°C)
Twb -1 -Wet bulb temperature (°C) (kJkg °C Mean velocity across cross-sectional area (m um -1 s ) -1
UA
Overall thermal conductance (W °C )
W
Humidity ratio (kghumidity/kgdry air)
Z
Altitude (m)
k K
Thermal conductivity (kW m °C )
h fg
Specific enthalpy of phase change (kJ kg )
Heat exchanger efficiency, dimensionless
L
Length (m)
fan
Fan efficiency, dimensionless
-1
-1
Pressure loss coefficient, dimensionless
Greek symbols -1
m
Mass flow rate (kg s )
mot(kg s-1Fan electric motor efficiency, dimensionless
ndays
Density (kg m )
Pb
Number of calendar days within a given a month Number of transfer units, dimensionless Nusselt number, dimensionless Barometric pressure (kPa)
Difference 2 -1 Kinematic viscosity (m s ) Fraction of mixture, dimensionless
Psens
Power recovered from sensible heat (kW)
Plat
Power recovered from latent heat (kW)
c
Cold fluid
Pr
Prandtl number, dimensionless
f
Film
3 -1
h
Hot fluid
-1
NTU Nu
-1
-3
Subscript
Q
Volumetric flow rate (m s )
R
Thermal resistance (°C W )
i
Inlet
Re S
Reynolds number, dimensionless Cost ($CAN)
o
Outlet
4.02
Calculation of heat capacity rate of air at exhaust
The heat capacity rate of exhaust air is given by Equation 4-1.
Cair c p, air air, dry Q air
Eq. 4-1
The barometric pressure is first estimated from the altitude of the mine site. Z
Pb 101.3 *1019075
Eq. 4-2
The density of dry air is then calculated from Equation 4-3. air, dry
Pb 0.287Tair, dry
Eq. 4-3
Assuming that the volumetric flow rate data is measured upstream of the heat exchanger it is valid to use the dry bulb temperature at exhaust entered by the user. 22
c p, air is assumed to be constant at 1.005 kJ/kg °C. This value should be valid within the range of
the average dry bulb temperature upstream and downstream of the tube and fin heat exchangers.
4.03 Predicting air conditions at HE outlet (humidity and dry bulb temperature) The volumetric flow rate, dry and wet bulb temperatures at the exhaust are the most important parameters since they will determine the total rate of heat that can be recovered. It will be assumed that these parameters are constant. The exhaust air conditions could differ from one day to another but not significantly. Thus an average exhaust temperature and humidity ratio should be given. According to (Lafontaine, 2008), at Laronde mine during winter months, the exhaust air dry bulb temperature at surface is around 18°C ±2°C. This fluctuation is most likely dependent on the change in productivity; the more tonnage, the more there will be freshly broken virgin rocks and equipment running and thus the more heat will be generated inside the mine. The ambient air temperature will not affect the exhaust air temperature as it is warmed at a constant temperature during the cold months of the year. The approach used to predict the air conditions downstream of the exhaust HE will first be described graphically on a psychrometric chart to ease the understanding of the concept. The calculations in detail will be shown following the explanation. A heat exchanger wall temperature has to be approximated before obtaining the outputs. From this approximated wall temperature, a calculated wall temperature is obtained and the user can easily iterate until the approximated and calculated values are similar. First, from the dry and wet bulb temperatures at exhaust, a point is drawn on the psychrometric chart. Then for the same humidity ratio (W) located on the Y-axis of Figure 4-1 another point is drawn on the psychrometric chart at a dry bulb temperature equal to the wall temperature. The different cases to which the software adapts and calculates the amount of energy that can be extracted from exhaust air are described in the following.
23
4.03.1 First case: Dry cooling of exhaust air
Figure 4-1: First case: Dry cooling of exhaust air
For this case shown in Figure 4-1, the assumed wall temperature is not smaller than the dew point temperature at the given humidity ratio upstream of the HE, thus no condensation occurs, solely sensible heat is transferred to the working fluid. The total power recovered, effectively the sensible heat transferred, is calculated as follows: Ptot Psens Cair (Th,i Tc,i ) Eq. 4-4 The calculation of the efficiency of the HE is described in section 4.04. 4.03.2 Second case: Cooling and dehumidifying of air The second case is shown in Figure 4-2. The assumed wall temperature is smaller than the dew point temperature; therefore as the air comes in contact with the HE walls, condensation occurs.
Figure 4-2: Cooling and dehumidifying of exhaust air
The air very close to the wall would be saturated; accordingly to (McQuiston, Parker, & Spitler, 2005) it is fair to approximate that the air conditions tend to move towards this saturated point in a linear manner. In this instance, an imaginary point is placed on the psychrometric chart on the saturation line at the wall temperature, as shown in Figure 4-2.
24
A straight line will then be drawn from the upstream HE air conditions to the imaginary wall temperature dew point. Then using the efficiency of the HE it is possible to find the outlet dry bulb temperature from Equation 4-5: Th,o Th,i
Cmin (Th,i Tc,i ) Ch
Eq. 4-5
Assuming that this temperature is the outlet dry bulb temperature, a point on the line as shown in Figure 4-3 should be located on the line connecting the two points. Two cases are then possible:
First case; the point on the line is lower than the saturation line at the predicted dry bulb temperature downstream of the HE as shown in Figure 4-3.
Figure 4-3: Outlet conditions predicted below saturation line
Second case; the point on the line is above the saturation line as shown in Figure 4-4. The second case will usually happen when the exhaust air conditions are close to the saturation line. The point will have to be moved down vertically on the saturation line curve since air conditions do not exist when the point is located on the straight line. This case can only happen if the straight line connecting the two points crosses over the saturation line.
Figure 4-4: Outlet conditions predicted above saturation line
25
The humidity ratio and dry bulb temperature at the blue points identified in Figure 4-3 and Figure 4-4 are compiled to determine the total power extracted from the return air. It is found using Equation 4-6, 4-7 and 4-8. Psens Cair (Th,i Tc,i )
Eq. 4-6
Plat h fg m dry,air (Win Wout )
Eq. 4-7
Ptot Plat Psens
Eq. 4-8
What was previously explained in the psychrometric chart is calculated as follow within the software. First, the saturation line curve function is developed. Using the inlet dry and wet bulb temperature at exhaust, the humidity ratio can be determined using the following equations: 17 .27Tdry Pdry 0.6105 exp Tdry 237 .3
17.27Twet Pwet 0.6105 exp Twet 237.3 Pu Pwet 0.000644Pb (Tdry Twet ) W 0.622
Pu Pb Pu
Eq. 4-9 Eq. 4-10 Eq. 4-11 Eq. 4-12
Using the dry bulb temperature and the humidity ratio, a point can be located on the psychrometric chart. Then the equation of the saturation curve is determined as follow: At saturation, Twet=Tdry and thus: Eq. 4-13
Pu Pwet Pdry
And the saturation line function is determined from Equation 4-14. W 0.622
Pdry Pb Pdry
17.27Tdry 0.6105 exp Tdry 237.3 0.622 17.27Tdry Pb 0.6105 exp Tdry 237.3
Eq. 4-14
Note: W is the value of the Y-axis and Tdry the value of the X-axis, Pb will remain constant for a given mine site.
26
To determine if condensation occurs, the assumed wall temperature is inserted into Equation 414. Then the calculated humidity ratio at exhaust is compared with the humidity ratio found using the wall temperature as described previously. If the humidity ratio found with wall temperature is equal or greater than the exhaust humidity ratio, the first case applies, no condensation occurs. On the other hand, if the humidity ratio found using wall temperature is lower than the humidity ratio at exhaust, condensation occurs and the equation of the line connecting the two points must be found: The linear equation between the two red points of Figure 4-2 is determined as follow: The slope of the line: a
Wi , exh Wwall
Eq. 4-15
Ti , dry Twall
The value of W when T is equal to 0 is given as: Eq. 4-16
b aTwall Wwall
The function of the line is found from Equation 4-17. W aTdb b
Eq. 4-17
From the assumed efficiency, the outlet dry bulb temperature is determined. The temperature is used in equation 4-17 to obtain the humidity ratio at outlet. In the case that it is higher than the saturation line as explained in Figure 4-4 the humidity ratio is modified accordingly. Then to find the latent power recovered, Equation 4-7 is used. The heat of vaporization is determined from Equation 4-18 (ASHRAE, 2009). The temperature is taken at the mean air dry bulb temperature across the HE. h fg 2501 1.805Tdb Eq. 4-18
4.04
Calculation of HE efficiency
At first, in order to determine the HE efficiency, the NTU-method was used. The main issue with this method was the lack of empirical solutions to determine the heat transfer coefficient on the air-side. This method is described in Chapter 5. Nonetheless, the HE efficiency will greatly differ when condensation occurs and very few correlations were found under these conditions. Vapour condensation will occur in most cases as the relative humidity of exhaust air in underground mines is usually close to saturation. Due to the latter, data was collected from Industrial Heat Transfer Inc. (IHT Inc.) for a detailed HE design for the Laronde mine ventilation installations as shown in Table 4-1. This data was used to determine the approximate efficiency of the tube and fin HE for different relative humidity. It is important to note that the efficiency takes into account solely the heat transferred under dry conditions therefore as the level of condensation increases, the efficiency will decrease as the heat transferred due to vapour condensation will increase significantly the fluid temperature decreasing the temperature difference between the two fluids and thus decreasing the HE efficiency. Although the efficiency decreases, it is important to note that the total energy recovered will be larger as the vapour condensation will increase the total heat transferred to the glycol mixture. 27
Relative Humidity % 0 10 20 30 40 50 60 70 80 90 100
Power recovered kW
Mass flow rate of condensate kg/s
417.4 417.6 417.9 418.5 418.8 427.9 449.0 477.4 511.4 551.9 591.4
0 0 0 0 0 0 0.068 0.126 0.191 0.248 0.315
Table 4-1: IHT Inc. data HE data for condensation
From this data the efficiency of the HE was determined using Equation 4-18. The results are shown in Table 4-2.
Ptot m cond h fg
Eq. 4-18
Cmin (Th,i Tc,i )
Relative Humidity % 0 10 20 30 40 50 60 70 80 90 100
Efficiency HE
of
0.675 0.675 0.675 0.675 0.675 0.69 0.59 0.52 0.45 0.39 0.35
Table 4-2: HE efficiency with respect to relative humidity
This data is valid for the following HE geometry and fluid conditions:
The HE is composed of several sets of small HE with the same geometry, 20 in total. It is assumed that the air and glycol flow is separated into 20 and thus the HE can be evaluated for one set of the smaller coil. The coil is composed of 8 rows of tubes with 46 tubes on each row. The flow is separated into 92 circuits having each 4 passes. The tube inner diameter is of 0.577 in (0.0001687 m2 cross sectional area). The tubes are made out of copper equipped with inside wire turbulators to initiate turbulent flow. The fins are made out of aluminum with a fin spacing of 120 fins per ft.
The rest of the dimensions are shown on the IHT Inc. data sheet in APPENDIX A. The glycol velocity within each pass is of 1.05 m/s (3.45 ft/s). The air face velocity speed is 3.2 m/s (10.52 ft/s). The inlet glycol temperature is of 1.5°C (34.7°F) and the inlet dry bulb air temperature is 18°C (64.4°F). The efficiency of the HE is dependant of the flow arrangement, HE geometry, 28
Reynolds and Prandlt numbers. The mean fluids temperatures are also dependent since they will determine the fluids properties that are required to determine the Reynolds and Prandlt number. Therefore, a difference in exhaust air temperature could influence the HE efficiency but assuming that the range of exhaust air can vary from a minimum of 5°C to a maximum of 30°C, the change in fluid properties should only affect the efficiency in very small proportions. Therefore for a given geometry, the air face velocity will be the only dependent variable of the efficiency as it will vary the Reynolds number. It is thus fair to approximate that the efficiencies in Table 4-2 are valid when the air face velocity is close to 3.2 m/s. A higher air velocity will reduce the efficiency of the heat exchange since the air will be in contact with the HE walls and fins for a shorter period of time as opposed to a lower velocity flow where the heat will have more time to dissipate. It is although important to note that a higher velocity will always result in a greater heat transfer rate as the heat capacity rate of the fluid is increased. It is possible to obtain a greater efficiency by modifying the geometry of the HE. It would although result in a greater pressure drop on both fluids side and an increase in capital cost. Due to the latter, if face velocity is 3.2 m/s or lower, the efficiencies in Table 4-2 are not necessarily the optimal ones but can still be used to obtain a fair approach of the desired results. Within the software, the user has to enter a HE efficiency under dry conditions, it is recommended to first use the result in Table 4-2 of 0.675. The HE efficiency calculation will then be solely dependent to the relative humidity of air. In the case where face velocity cannot be decreased to 3.2 m/s, the HE manufacturer should be able to determine the efficiency that can be achieved for a given pressure drop. The theory of tube and fin HEs is explained more into details in Chapter 7. From Table 4-2, the correlation shown in Figure 4-5 has been developed:
Figure 4-5: HE efficiency with respect to relative humidity
29
The correlation is used to approximate the HE efficiency where “x” is the relative humidity in percentage. It is assumed that this correlation is valid for any HE efficiency under dry conditions, thus for a higher or lower efficiency than in this case, the curve is assumed to be translated along the Y-axis of Figure 4-5. In the case of the HE designed for the Laronde mine, the efficiency under dry conditions was of 67.5% although it is said the efficiency can be up to 95% for multipass cross flow tube and fin HE (Shah & Sekulic, 2003). The user can change the efficiency of the exhaust HE to optimize the payback period but it is important to remember that the capital cost and pressure drop on both fluid sides is based on the 67.5% efficiency and should therefore be reviewed if a greater efficiency of the system is entered by the user.
4.05
Calculation of mean wall temperature
The mean wall temperature is required in order to determine if condensation should occur or not. As for the efficiency, in order to obtain the mean wall temperature, the heat transfer coefficient on the air side must be determined. Due to the uncertainty of the calculation of the heat transfer coefficient, a suitable correlation shown in Equation 4-18 has been developed with the air and glycol mean temperatures as variables. The mean wall temperature calculation is described in Chapter 7. The correlation was found from trials and errors using data from Table 4-1 and energy calculations described previously. Twall
Tair,m Tgly,m
Tgly,m
3.5
Eq. 4-18
From this calculated wall temperature, the user must enter the value of the assumed wall temperature and iterate until the two values are similar. This procedure has to be done since a wall temperature has to be assumed first in order to obtain the mean temperatures of the two fluids.
4.06
Calculation of the fluids mean temperatures
The heat capacity rate Cglycol is assumed to have a value 1.5 times greater than Cair. This assumption has been made based on the design study case performed by IHT Inc. at the Laronde mine. For design optimization, this value can be varied. From Cglycol and the total power recovered, it is possible to determine the outlet glycol temperature using the Equation 4-19. Tc,o Tc,i
Ptot Cglycol
Eq. 4-19
From the assumption that Cglycol is 1.5 times larger than Cair;
Cmin 0.5 Cmax
the mean temperature of
the fluid can thus be approximated from the following equations (Shah & Sekulic, 2003): Th, m
Tc, m
Th,i Th,o 2 Tc,o Tc,i
2
Eq. 4-20 Eq. 4-21 30
4.07
Glycol temperature above water freezing point
In order to ensure that the glycol temperature remains above the water freezing point, Equation 4-22 is used.
Tc,i
q C min
T h ,i
Eq. 4-22
For any ambient temperature below the one found in Equation 4-22 the glycol temperature will achieve a value below the desired glycol outlet temperature and will cause the condensate to freeze at exhaust. One way to avoid this issue is to have a glycol bypass valve at intake so that some of the hot glycol mixes with cold glycol to maintain the glycol outlet temperature above 0°C. The glycol bypass valve is described in section 5.11.
4.08
Calculation of nominal pipe size diameter
In order to transfer the ethylene glycol mixture from one HE to another, a piping system must be put in place in order to form the loop. The following will show how the software calculates the required nominal pipe size diameter of the piping that connects the two HEs. The volumetric flow rate is determined from Equation 4-23, the density of glycol is taken at its mean temperature:
Q glycol
m glycol
glycol
Eq. 4-23
From the volumetric flow rate, the pipe diameter of the main piping system can be determined. The diameter of the piping system is calculated so that the flow velocity does not exceed a given flow velocity entered by the user. The pipes used are Schedule 40 nominal pipe size (NPS) as shown in Table 4-3. The maximum flow velocity can be changed to optimize economic tradeoffs between the operating and capital cost of the piping system. Equation 4-24 is used to obtain the pipe diameter. 4Q gly D u m
0.5
Eq. 4-24
The software then returns the closest maximum NPS internal diameter of Table 4-3. The available diameters are from 2’ to 24’. Note that the pipe size diameter of 22" is excluded since it is usually not available at the piping manufacturer.
31
NPS (in)
Diameter (in) External 2 4 6 8 10 12 14 16 18 20 24
Internal
2.38 4.5 6.63 8.63 10.75 12.75 14 16 18 20 24
2.07 4.03 6.07 7.98 10.02 11.94 13.13 15 16.88 18.81 22.63
Table 4-3: Nominal pipe size
The new fluid velocity within the piping system will be determined from the returned value of the maximum internal pipe size diameter using Equation 4-25. um
4Q
Eq. 4-25
d i 2
4.09
Pressure drop across piping system
From the distance between intake and exhaust, an approximate pipe length will be determined. The assumption is that the pipe length is 15% greater than the distance separating the exhaust from the intake on both fluid sides. The number of bends is entered by the user. From (CRANE, 1982), the K factor (pressure loss coefficient) for each of the bends will be 30ft where ft is the friction factor at fully-turbulent flow. The friction factor is assumed to be of 0.06, it is chosen from the typical roughness of pipe material (Binder, 1973) carbon steel in this case. The friction factor is chosen in a conservative matter to take into account the fouling and corrosion of the pipe that could occur with time. From these inputs and an assumed large Reynolds number, a friction factor can be determined from the friction Fanno flow graph shown in Figure 4-6. The friction Fanno flow graph is used to determine a friction factor of a pipe for a given Reynolds number and relative roughness of the pipe. For each bend; the K factor is 1.8. The total head loss of the main piping system excluding the piping accessories is calculated using Equation 4-26: h pip (
glycolu m fL K) di 2
2
Eq. 4-26
32
Figure 4-6: Friction fanno flows (Shah & Sekulic, 2003)
From the HE design, a manifold system is required when entering and exiting the HEs. The losses in the manifold system are calculated as follows. The running pipe is divided as a tee into several branches. The number of branches is dependent on the number of coils required and is described in section 5.02. As the flow is divided into a branch as shown in Figure 4-7, it runs through about 3 m of piping and one elbow to connect to the tube and fin HE with the use of a flange.
Figure 4-7: Tee of manifolds
The K factor across the flow through the branch is of 60ft and across the header is of 20ft. The K factor of the elbow is of 30ft as in the main pipe system. The pressure drop due to the ball valve fully-opened is of 3ft. The valves should always remain fully-opened unless a set of coil requires maintenance (CRANE, 1982). The pipe diameter is of 4”. The friction factor at turbulent flow is as well assumed to be at 0.06. There are in total four manifolds, inlet and outlet of the HEs located at both intake and exhaust. The mean velocity through one branch is found from Equation 4-27, it is used to approximate the friction and shock losses in the manifold system. u m ,b
4Q nbd i2,b
Eq. 4-27
33
The total head loss across the manifolds is then found using Equation 4-28. fL glycolum,b 2 hmani nb 60 f t 30 f t 3 f t 20 f t 2 di ,b
Eq. 4-28
These pressure drops will be calculated separately for intake and exhaust since they can have a different number of coils required. Assuming that the losses within the manifold at inlet and outlet are the same, the pressure drop calculated for intake and exhaust requires to be multiplied by two. Since the coils are arranged in parallel, the required useful pumping power is determined by multiplying the total pressure drop in both HE by the total glycol volumetric flow rate. The HE at intake will have a lower efficiency than at exhaust. By comparing Equation 4-29 for exhaust and intake, the numerator should be the same in both cases assuming that they are no heat losses with the surroundings; on the other hand, the denominator should be greater due to the larger fluids inlet temperature difference at intake which will result in a lower HE efficiency for the same geometry at dry conditions.
Ch (Th,i Th, o ) Actual Power Maximum possible power Cmin (Th,i Tc,i )
Eq. 4-29
The temperature difference between the two fluids at the intake shaft is greater since the HE is designed to obtain an air outlet temperature slightly higher than the freezing point. Due to the latter, the design ambient temperature is relatively low which should result in a larger inlet fluids temperature difference. Decreasing the efficiency should reduce the pressure drop on both fluid sides. For example, in the Laronde design case, the HE at intake solely has two glycol passes as opposed to four passes at exhaust which reduces the pressure drop on the glycol side. Many other geometrical parameters of the HE can be modified such as the fin spacing and tube diameter. The pressure drop across one set of coil on the glycol side is given by the HE manufacturer; 67 kPa at exhaust and 30 kPa at intake. The check valve has a K factor of 100ft and the butterfly valve of 35ft. These values were averaged as they slightly change for different pipe diameters. It is assumed the design is composed of six butterfly valves and one check valve. 4.09.1 Calculation of pumping costs The amount of useful power required for the pump will then be calculated by adding the pressure drop of the two heat exchangers, the main piping system, the manifolds and the accessories to then multiply the total by the glycol volumetric flow rate. h p,tot hHEs h pip hmani hacc Eq. 4-30
Pusef QGlycol hp,tot
Eq. 4-31
Then the actual power required can be estimated by assuming an average pump and electric motor efficiency. The electric motor efficiency is assumed to be 0.95, and the average operating pump efficiency at 0.75. Pact
Pusef
mot pump
Eq. 4-32
34
Then for a given electricity price, an approximate operating pumping cost per year is obtained. The operating cost is only calculated for the months in which the system has a lower air temperature than the intake minimum temperature required i.e. for the months when the system is in operation. The electricity price is given in $/kWh thus the energy cost for each of the operating months is calculated from Equation 4-33 for a given power in kW. S pump ndays
SkWh sec Pusef 86400 3600 sec day
Eq. 4-33
When the system is not required to operate at full load, the glycol flow rate can be reduced with the use of a variable frequency drive. This option is not currently included in the software but it could reduce the system’s operating costs. The other option could be to incorporate a heat pump system and heat the surface buildings with the additional heat.
4.10
Calculation of additional fan power cost
The pressure loss on the air-side of the HE is quite an important factor in the economics of the project. It is presently assumed that the HEs are left in place throughout the whole year therefore the energy cost due to the additional fan power required is continuous. The pressure drop was given by the manufacturer: 250 Pa at exhaust and 225 Pa at intake. These values are valid for the HE geometry described in APPENDIX A and the air face velocity of 3.2 m/s at exhaust and 3.4 m/s at intake. If the air velocities are much different, a pressure drop should be re-evaluated by the HE manufacturer. In the case where the pressure drop across the HE would be large, the HE should be designed in such a way that it can be easily removed and installed during the nonoperating months. The other option would be that the airflow bypasses the HEs which would still generate losses due to shock. The additional useful fan power required is determined from Equation 4-34. Puse, fan Qair,i Pi Qair,o Po Eq. 4-34 The fan and motor efficiency are taken from the user inputs. The actual power is calculated from Equation 4-35: Pact, fan
Puse, fan
fanmot
Eq. 4-35
Then from the electricity price, the additional fan operating cost per year is estimated. The additional energy consumption is calculated for 365 days per year. The energy cost for a useful power given kW is calculated from Equation 4-35. S fan ndays
S kWh sec Pact, fan 86400 3600 sec day
Eq. 4-35
To calculate the net energy cost balance per year the Equation 4-36 is used: Sbal Ssav S pump S fan Eq. 4-36
35
4.11
Calculation of energy savings per year
For every month, an average outside temperature is given. It is assumed that this temperature is constant throughout the whole month. As it is assumed that the insulation of the piping in between the two HEs is relatively efficient, it was decided to assume that the heat loss with the surroundings in the piping system is equal to the heat gain from the pump, from the first law of thermodynamics; the power recovered at exhaust is therefore equal to the power transferred at intake. From the intake air volumetric flow rate, the maximum temperature difference that the intake air can achieve from the heat recovery system is determined using Equation 4-37.
T
min, air c p , air Ptot
Eq. 4-37
The user has to enter the desired minimum required air intake temperature i.e. the air temperature during cold periods usually achieved with the gas burners. This temperature is required to calculate the energy savings and will usually be slightly greater than the freezing point. In order to calculate the amount of energy saved, the following different cases have to be taken into consideration:
1st case: The outside air temperature is greater than the minimum required intake air temperature, thus the system is not running and there are no energy savings.
2nd case: The outside air temperature is lower than the minimum required intake air temperature. The temperature difference between the outside air and the minimum required fresh air temperature is lower than the maximum possible temperature difference that the heat recovery system can deliver. In this situation, the heat recovery system will not be running at full load and no heat is required from the gas burners. In this case, the energy savings are calculated from Equation 4-38 and 4-39.
T Tmin Tamb
Esav min, air C p, airT ndays/ month 86400 sec
Eq. 4-38 Eq. 4-39
3rd case: The outside air temperature is lower than the minimum required fresh air temperature. The temperature difference between the outside air and the minimum required intake air temperature is greater than the maximum possible temperature difference that the heat recovery system can deliver. In this case, the system would be running at full load and the energy savings will remain the same for this range of outside temperatures. Furthermore, the gas burners have to provide additional heating to achieve the minimum intake air temperature. In this situation, the energy savings are calculated from Equation 4-40. Esav Ptot ndays 86400 sec Eq. 4-40
36
For each separate month, the energy is calculated from the average given value. The total energy savings are then added for each month and then divided by the burners efficiency to obtain the total annual energy savings. The burners efficiency is a variable entered by the user, the first fair approximation can be of 92% (Fytas, 2008). To obtain more accurate results of the energy savings; climatic data can be compiled for the given mine site region and the number of hours per year within a given temperature range can be approximated. For example, calculating the average number of hours per year from data compiled over the last 10 years when outside temperature was between -35°C and -31°C. Then outside air temperature would be approximated at -33°C and the total energy saved would be calculated for the number of hours instead for one full month. All the temperature range below the minimum required intake air temperature would be compiled. This could be done for a detailed feasibility study. From the fuel cost, it is possible to determine the cost savings per year. The fuel cost is in $/GJ, the cost savings can be calculated by converting the energy into GJ and multiply it by the fuel cost. Ssav S fuel Esav Eq. 4-41
4.12
Calculation of maximum pipe heat loss to surroundings
The pipe heat losses to surroundings are not included within the energy calculation, in order to ensure that the losses with the surroundings are not significant compared to the total amount of power recovered, the maximum pipe heat loss to surroundings is calculated. In the case that the user detects that it could affect greatly the energy savings, an improvement in the pipe insulation material should be done. The present chosen insulation material is Foamglas® 1.5” thickness, the thermal conductivity of the material is; 0.038 W m-1 K-1. The thermal conductance per area will thus become: 1.05 W m-2 K-1. The maximum value will be taken for the lowest monthly ambient temperature and glycol minimum temperature. The area used is the total inside pipe area of the main piping system. The maximum pipe heat loss to surroundings is found from Equation 4-42. Ploss k ins Ain, pip,h (Tamb Tgly,h ) Eq. 4-42 Note that this is the power loss if pipes are exposed to ambient air, the software assumes that underground piping is used and thus the heat losses will be significantly lower. As mentioned in section 4.10, the software presently assumes that the pipe heat loss to surroundings is equal to the heat transferred by the pump work. Due to the latter, the amount of power recovered is equal to the amount of power discharged in the intake.
4.13
Calculation examples for heat losses
The following will show some examples of calculations to predict the heat losses to surroundings for different design cases. For the glycol and pipe data, the Laronde mine design study case is taken as an example. Note that the heat loss calculation examples are found assuming that the 37
wall temperature is constantly at the fluid temperature which is a conservative approach i.e. the actual heat loss should be slightly lower than the calculated value. 1st Example:
Underground piping with PVC pipes 14” NPS Schedule 40, no insulation Length of the piping system; 230 m on each side Glycol mix flow velocity: 3.46 m/s Dynamic viscosity of glycol mixture: Wall thickness; 0.438” (0.0111 m) PVC thermal conductivity: 0.19 W/ m K, for a wall thickness of 0.438”; k: 17.12W/ m2 K
The depth at which pipe is buried is 3 m; from the soil temperature data in APPENDIX A, the temperature surrounding the pipe is assumed to be 5°C. The temperature of the glycol mix on the hot side is: 12.95°C ΔT=7.95°C The total inside area of the pipe is of 241 m2 on each side. Ploss k pvc Apip, wT =32.8 kW Thus for the total heat recovered of 10.57 MW, 0.0328 MW is not a significant amount of heat loss and therefore PVC piping with no insulation should be enough. For the same design case except using Carbon steel pipes SCH 40 (k; 2443.7 W/m 2 K), the heat loss would be: 4.7 MW which will greatly affect the performance of the system. In this design case, it was assumed that the cold glycol piping side temperature is 3°C and thus by assuming that the soil temperature is 5°C, the soil would transfer geothermal heat to the glycol mixture and the system would become more efficient. The calculations will be performed assuming that the soil temperature always remains at 5°C. Carbon steel pipes would be used on the cold side to enhance the heat transfer between the soil and the ethylene glycol mixture.
Carbon steel thermal conductivity: 54 W/m K Carbon steel pipe Schedule 40 thickness: 0.87” k: 2443.7 W/m2 K ΔT=2°C Apip,w:241 m2
From Equation 4-42, Pgain =1.17 MW To determine the glycol mixture temperature increase, Equation 4-43 is used.
38
T
Pgain C gly
1.11°C
Eq. 4-43
Where Cgly= 1060 kW/K The glycol temperature at intake HE inlet would become 4.11°C which would increase significantly the energy savings of the system.
4.14
Calculation of mass flow rate of condensate
The mass flow rate of condensate is shown to determine the capacity of the drainage system when condensation occurs at exhaust, it is found from Equation 4-44.
m cond
Plat h fg
Eq. 4-44
4.15
Ethylene glycol mixture thermophysical properties
The density and specific heat capacity rate of the ethylene glycol mixture is required to complete the energy calculations. Some correlations were found from (M. Conde Engineering, 2002). The correlation takes into account the fraction content of glycol in the solution; it can thus be used for different mixtures. The correlation is as follows:
Px A1 A2 A3
273.15 273.15 273.15 A4 A5 T T T
2
For density (kg/m3): A1: 658.45, A2: -54.815, A3: 664.71, A4: 232.73, A5:-322.62 For specific heat capacity rate (kJ/kg K): A1: 5.364, A2: 0.7886, A3: -2.590, A4: -2.732, A5:1.437
4.16
Conclusions
In order to study the feasibility of exhaust air heat recovery, energy calculations are performed. From constant exhaust air conditions, the maximum possible power recovered is calculated. The energy calculations showed the importance of latent heat content within exhaust air. From the maximum possible power recovered, the net energy savings are estimated from average monthly temperatures. It is important to remember that ambient air temperature can differ from one year to another which could sometimes affect the savings results obtained. From the mine installations and capacity of the heat recovery system, the total operating costs are determined. The pressure drop on the air side of the HE must be minimized as much as possible to reduce operating costs of the system. 39
The total amount of heat from exhaust air can often be under-estimated due to non-consideration of latent heat from saturated exhaust air. Shallow mines should not eliminate the possibility of exhaust air heat recovery systems as significant amounts of heat can still be extracted from the geothermal heat source of lower levels. For future work, a solver could be put in place to calculate pipe heat losses to the surroundings from the chosen insulation material. Recovering heat from exhaust air should be evaluated for any operation using the software. It is obvious that the price of fuel will greatly affect the economics of the projects; any operation should take the time to use the software and estimate the benefits of recovering heat from exhaust air especially in the case that the mine uses expensive fuel.
40
CHAPTER 5. CAPITAL COST AND DESIGN CONSIDERATIONS Summary This Chapter describes the detailed capital cost calculation of each of the components of the closed-loop glycol circuit. It also outlines the design considerations and recommendations of the system. The economics of the project using different case studies is as well discussed.
5.01
Introduction
As mentioned earlier, the chosen design to study the feasibility of recovering exhaust air is the closed-loop glycol circuit. The design requires a relatively large amount of components to be functional and safe. The feasibility study software calculates the cost of each of the components that the system necessitates. The capital costs of the components were calculated from the inputs entered by the user and the various assumptions made are described in the following. For simplicity, some design considerations are included with the assumptions and calculations. Some design alternatives from the ones included in the software are also mentioned. Note that as a reference for some calculation examples, the Laronde design case has been taken as a model. It is important to note that the Laronde mine is one of the largest underground mines in Canada hence the cost and size of the installations will usually be much greater than in smaller shallow mines. The capital costs calculations are divided into following components:
Heat Exchangers and its installations Underground piping installation Pumps and electric motor Piping accessories Manifolds Automated washing system
In the energy calculations, several variables from the existing or future mine sites are required. Some of these variables will also be used in the capital cost calculations; some others are required solely for the capital costs calculations and are as follows:
Actual face area of the exhaust and intake ventilation installations Distance between the main power supply and intake ventilation installations (choose shortest) Distance between exhaust, main power and water supply Nature of soil (common earth, loam and sandy clay, sand and gravel and hard clay) Numbers of elbows per path Labour rate ($/hr) 41
The application of these variables within the calculations will be explained throughout the Chapter. The costs of the components within the system were mostly determined using RSMeans© book collection. These books have a wide variety of materials for any types of projects. There are several books which are separated into different fields that were used for this project, they are as follow:
RSMeans mechanical 2010; RSMeans electrical 2010 RSMeans construction 2008 RSMeans site work 2010 RSMeans assemblies 2010
Two other books were used to determine the pump and electric motor cost as well as the man hours for welding activities;
Mine and Mill equipment cost guide 2004 Estimator's piping man-hour manual 1999
Most of the costs were found in these books and tabulated within the software. For example, the cost and labour hours for the couplings of the piping system of a given nominal pipe size is entered for diameters from 2 to 24” as shown in Table 5-1. The program then executes a function and assigns the proper value within the table. NPS
Material
Labour
in
$
hrs
2
16.4
0.16
4
31
0.32
6
53
0.48
8
83.5
0.571
10
149
0.686
12
167
0.75
14
192
1
16
250
1.2
18
289
1.333
20
395
1.5
24
505
1.846
Table 5-1: Couplings material cost and labour hours with respect to NPS
For the labour cost, the hours are multiplied by a constant labour rate of the workers. The labour rate must be entered by the user. The cost tables are included in APPENDIX B, the page at which you can find the cost in the referenced book is included within the title of the Table.
42
5.02
Heat exchangers and its installations
The heat exchangers are used to exchange the heat between the air and the glycol mixture. They are tube and fin heat exchangers and their theory will be explained more into details in Chapter 7. This section will explain the capital cost calculations and design considerations of the installation of the coils and the HE building extension. For usual mining applications, the HE will be of relatively large size. Therefore it will be composed of several sets of coils i.e. one HE is composed of many coils in parallel of smaller size. This design requirement is due to several reasons; ease the maintenance, manufacturing limitations and of course handling purposes since a too large coil cannot be transported and would also be very difficult to install. The heat exchangers are required to occupy a relatively large face area in order to enhance heat transfer but mainly to reduce air pressure drop across the coils. If the heat recovery system is installed on an operating mine site, the ventilation surface installations will most likely have to be modified. Nomenclature
A C
Area (m )
um
Mean velocity across cross-sectional area (m s )
Flow stream heat capacity rate (kW °C )
Vs
Total system’s volume (m )
di
Pipe outside diameter (m)
Vt
Expansion tank’s volume (m )
do
Pipe inside diameter (m)
W
Width (m)
h
Head loss (kPa)
W pipe
Weight of pipe
2
-1
k K
Thermal conductivity (kW m °C )
hg L
-1
-1
-1
3
3
Greek symbols
Linear coefficient of thermal expansion (m/ m K)
Specific enthalpy of phase change (kJ kg )
Heat exchanger efficiency, dimensionless
Length (m)
fan
Fan efficiency, dimensionless
Pressure loss coefficient, dimensionless -1
m
Mass flow rate (kg s )
mot(kg s-1Fan electric motor efficiency, dimensionless
ndays
Number of days within a given a month
Density (kg m )
Pb
Barometric pressure (kPa)
Difference
Psens
Power recovered from sensible heat (kW)
Plat
Power recovered from latent heat (kW)
h
Pr
Prandtl number, dimensionless
c
Kinematic viscosity (m s ) Specific volume of glycol/water at higher 3 temperature (m /kg) Specific volume of glycol/water at lower 3 temperature (m /kg)
-1
-3
2 -1
Q
Volumetric flow rate (m s )
Fraction of mixture
Re S
Reynolds number, dimensionless Cost ($CAN) Pipe thickness (m)
c f
Subscript Cold fluid Film
Temperature (°C)
h
Hot fluid
Tdb
Dry bulb temperature (°C)
i
Inlet
Twb
Wet bulb temperature (°C)
o
Outlet
t
3 -1
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5.03
Cost of the tube and fin heat exchangers
For the Laronde design, the total price of the HE at intake and exhaust was approximated to be at a total of 814 685 USD from the Industrial Heat Transfer Inc. (IHT Inc.) quote. It is assumed that 55% of the cost represents the HE at exhaust and 45% for the HE at intake. The exhaust’s greater volumetric flow rate and efficiency justifies the superior cost. From this assumption: Cost of the HE at exhaust: 448,100$US Cost of the HE at intake: 366,700$US At exhaust, IHT provided that the design would be composed of 20 separated identical coils placed in parallel. The cost of a single coil would be of: 22 405$. The capital cost calculation will depend on the number of coils required for the given operation. The volumetric flow rate of air at exhaust for the Laronde mine is of 1,300,000 cfm (610 m3/s) for a flow rate of 65,000 cfm (30.7 m3/s) per coil. The number of coils is determined from Equation 5-1. N coils
Q Int 1 30.7
Eq. 5-1
Where “Int” returns the integer value of the result. At intake, IHT determined that the design would be of 16 separated identical coils placed in parallel. Therefore the cost of a single coil at exhaust is: 22 920$. The volumetric flow rate of air at intake at the Laronde mine is of 1,190,000 cfm (560 m3/s), for a flow rate 74,375 cfm (35.1 m3/s) per coil. The calculation for the number of coils will be as in Equation 5-1 except for a different flow rate. The number of coils is then multiplied by the unit price of both intake and exhaust. In order to have the specifications of the coils at intake and exhaust, see APPENDIX A. It is important to mention that the coils will require Heresite® coating to minimize corrosion problems. It could be possible that after several years of operation a new layer of coating might have to be applied. The coating application should depend on the air conditions at exhaust and should definitely be discussed with the supplier as the life of the coils will greatly depend on it.
5.04
Manifolds
To divide the running fluid at each coil, a manifold system is required. The flow must be divided at the outlet and inlet of the heat exchanger. A total of four manifolds must be installed, two at intake and two at exhaust. The geometry of each manifold is as follows; the pipe is 4” diameter to carry and discharge the fluid to each coil. The flow is separated with the use of a weldolet which is welded onto the main piping system. A flanged flexible hose connects the coil with the manifolds piping branches. It is assumed that each branch requires approximately 3 m of piping with one 90° elbow which is connected with couplings. Downstream of the weldolet, a ball valve is installed to isolate the coil 44
in the case of a leak or for maintenance. The ball valve is connected to the piping system with the use of threads. The schematic of the manifold system is shown in Figure 5-1.
Figure 5-1: Manifold schematic
For one branch, the cost and labour time of each of the components is shown in Table 5-2. The cost is solely dependent on the number of coils. The manifold system shown in Figure 5-1 should be placed vertically along the coil sets to minimize the length of each branch. The type and orientation of the connection between the manifold and the coils should be decided with the HE manufacturer. From the design of the connection on the coils, the type of manifold system design can be chosen. Work or/and material
Material $
Labour hrs
Hole cutting in main pipe Weld-o-let
0.6 58
2.667
3 pipe cut 4” dia.
3 x 0.205
2 pipe grooving
2 x 0.186
2 couplings 4” dia.
2 x 31
2 x 0.32
Elbow
41.50
0.640
Flange 5 ft (3 m) of 4” dia. Ball valve Flexible hose flanged Total
35
1.6
48.25
1.46
720
0.421
580
1.667
1486
9.22
Table 5-2: Manifold system components
Instead of using a weldolet, an IPS hole cut could be used as shown in Figure 5-2. The labour cost would decrease but the capital cost would be much greater.
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Figure 5-2: IPS hole cut (Victaulic, 2009)
5.05
Ventilation Building extension cost
As mentioned earlier, the ventilation installations will require some modifications to decrease the air velocity across the coils. The usual surface ventilation installations both at intake and exhaust will most likely require the installation of a building extension. In order to calculate the required face area of the surface ventilation buildings, Equation 5-2 is used: Areq A1, coil N coils Eq. 5-2 In order to calculate the cost of the building extension, it is assumed that installations have an elbow as shown in Figure 5-3 i.e. that the airflow is parallel to the ground. In order that the flow diffuses properly within the building extension, a pyramid rectangular shape as in Figure 5-3 should be built to avoid major shock losses. It will be assumed that the required building length is proportional to a usual mine ventilation diffuser. The actual face area (present diffuser installed) and required face area will both assumed to be of a squared shape to calculate the length of the building. The optimal angle required of a usual round diffuser is from 8 to 11° (Fytas, 2007), 11° angle will be chosen. The equation to calculate the length is as follow; Areq Aact L
2 tan 11
Eq. 5-3
Figure 5-3: Surface building extension
46
It will be assumed that the maximum height of the steel buildings is of 24’ (7.315 m), thus when the calculated required face area is larger than a 24’ cube, solely the building width should increase.
Figure 5-4: Surface building extension building top and side view
The actual size of the building are then found using the following equations: W
Areq
Eq. 5-4
24'
The length of the foundation of the walls is determined from the width, and length of the new building and the actual face area (assuming it as a square of length B). Lwall, f
L cos
Eq. 5-5
Where arctan Where B
(W B) / 2 L
Aact
Eq. 5-6 Eq. 5-7
The area of the side walls is found using Equation 5-8. L ( H B ) L cos Awalls 2 B 2 cos
Eq. 5-8
The area of the roof is found from Equation 5-9.
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L L Aroof B (W B) cos cos
Eq. 5-9
The floor area is found from Equation 5-10. A floor LB L(W B)
Eq. 5-10
At each extremity of the HE Fins (On top and on the sides), insulation should be put in place to avoid direct contact with outside air. Low vapour permeability insulation should be used since the environment should be highly exposed to humidity. Some insulation could also be installed in the existing exhaust installations to reduce heat losses with surroundings. In the case that there are several diffusers located close to each other as in Figure 5-5; the new exhaust building should contain all the diffusers. If they are relatively far from each other, the parallel openings should each have its own building with its respective set of coils.
Figure 5-5: Exhaust ventilation collar at surface (Gagnon, 2011)
It should be important to note that the building extension will decrease the velocity of air and increase its static pressure as in an evase (diffuser). The additional power savings from the building extension could therefore be calculated using Equation 5-11 if the efficiency of these types of evase (pyramid rectangular prism shape) are known.
Q3 PR c 2
1 1 2 2 Ain Aout
Eq. 5-11
The efficiency ( c ) of a usual round cross-section diffuser can be found from the geometrical parameters and Figure 5-6. The efficiency would most likely be less as round shapes are smoother than square shapes.
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Figure 5-6: Diffuser efficiency (Hartman, 1982)
5.05.1 Foundations for coil supports and walls At the extremity of the building, the foundations will support the load of the coils. The foundations cost is assumed that they are built within the soil. The load is calculated from the following assumptions;
Building height of 24’ (7.3 m) There are four coils on top of each having a weight of 3660 lbs (1660 kg) each The width of the coil set is of 18’ (5.5 m) The weight of the glycol mix within the HE is 820 lbs (372 kg) (Calculated in section 3.08.9) and Steel support weighs 400 lbs (181 kg) per coil
The total weight for one coil is then: 4880 lbs (2214 kg). With four coils placed on top of each other, there is a total weight of 19520 lbs (8854 kg) over the length of 18’ (5.5 m); thus a load 1100 lbs (500 kg) per linear foot. The remaining foundations will solely support the roof and walls and therefore require supporting a lighter load. The foundations are although assumed to be the same to facilitate cost calculations. The cost is dependent on the total length of the foundations which is found from Equation 5-12 and the cost per linear foot is in Table 1 of APPENDIX B. L fds 2Lwalls, f W Eq. 5-12
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5.06
Building
From the width and height of the building, the function assigns the calculated dimensions to the nearest maximum value of Table 2 in APPENDIX B; the type of building chosen for costing is pre-engineered steel buildings. The cost is calculated with respect to the floor area. The floor area is assumed to be: A WL which is not the actual floor area that was calculated from Equation 3-10. The building shape is assumed to be as in Figure 5-7.
Figure 5-7: Assumed building shape for cost estimation, top view
The actual geometry of the building is a rectangular pyramid occupying a smaller volume as in Figure 5-3. The data provided from RSMeans are pre-engineered steel buildings and their cost will be lower than the actual building shape (pyramid). It is supposed that the greater complexity of the project is offset by the larger volume assumption in terms of cost. 5.06.1 Slab on grade for the HE building. The slab should not have to support heavy loads. The type of slab on grade chosen for costing is 4” thick, non industrial and non-reinforced. The cost is found in Table 3 of APPENDIX B, with respect to the actual floor area calculated with Equation 5-10. 5.06.2 Insulation of building Insulation of the building could be required to reduce as much as possible heat losses to surroundings. Calculations should be performed to ensure that the heat losses are not too significant. The chosen type of insulation for the building is; 1.5” thickness, R5, vynil/scrim/foil. The cost is calculated with respect to the total area of the roof and walls which is found using Equation 3-8 and 3-9. The cost data is found in Table 4 of APPENDIX B. 5.06.3 Coils support To determine the cost of the coil supports, an experienced metal worker was contacted (Lacasse, 2009). From the size and weight of the coil the following cost was approximated: Material: 500$ per coil Labour: 3 hours per coil 5.06.4 Coils arrangement In order to minimize friction losses across the HEs, a design should be implemented to remove or bypass the HE during the non-operating months. The bypass will induce a change in flow 50
direction which consequently will increase shock losses. A design proposal is shown in Figure 5-8. It consists of having large doors on the sides of the HE building that would open during nonoperating months. The doors would be hinged at the extremity of the sides of coils and would block the coils when they are in the opened position.
Figure 5-8: Airflow bypassed at the exhaust building, top view
A design that would completely remove the coils from their operating position would be much more efficient since no losses would be encountered in the non-operating period. The proposal presented is to have coil supports with train wheels and a rail as in Figure 5-9. The foundations with rails would be extended linearly to the outside of the face area of the building. The coils would be moved from the outlet area of the building extension during the non-operating periods. The main piping system would first have to be disconnected and the coils could then be rolled to the outside.
Figure 5-9: Train wheels for coils support
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The mine ventilation system will usually have to overcome a lower pressure drop in the winter season as the natural ventilation effect will be greater due to a larger temperature difference between downcast and upcast shaft than in summer months. In the case that the main ventilation fans pressure is not large enough to overcome the additional pressure drop from the coils, a booster fan would be required. The moveable coils could eliminate the requirement of a booster fan as the pressure drop is only induced during the cold periods and the natural ventilation increase could overcome this additional pressure drop from the coils.
5.07
Main piping system
In order to carry the fluid around the loop, a piping system requires to be installed. There are two installation options; using pipe supports or digging a trench to bury the pipes below surface which is usually called underground piping. Both options are described in the following section. Design considerations and capital cost calculations of both systems are presented. The main piping system is considered to be composed of pipes, couplings and elbows. The remaining components of the system will be included in the piping accessories section. The chosen piping material is carbon steel. The pipes will be connected with couplings. The pipes usually come in lengths of 21 ft. The total number of pipes required is first determined by dividing the total pipe length in feet and dividing it by 21, this value then returns its integer and adds 1 as in Equation 5-13. L pip n pip Int 1 21
Eq. 5-13
The number of elbows has to be entered by the user. At first, a fair approximation can be of 16 elbows on each side for a total of 32. It will be assumed that there are one third additional couplings required to the number of elbows required. The number of couplings required for the main piping system is calculated from Equation 5-14. L pip 1 ncps Int 1 Int nelb 1 21 3
Eq. 5-14
The cost is calculated with respect to the NPS, from Equations 5-15 and 5-16. The cost data is found in Table 8 of APPENDIX B. Slab, pip Slab L pip H pip n pip H 2, gr ncps H cps nelb H elb Eq. 5-15 Smat, pip L pipS pip ncps Scps nelb Selb
Eq. 5-16
5.07.1 Underground piping The piping system is located below ground level at a certain depth, in the case of a pipe or coupling failure, it would be difficult to locate and repair the damaged part. Due to the latter, the system has to be reliable as leaks or failures could create serious problems such as soil contamination. The quality of the couplings has to be good enough to avoid any failure 52
throughout the whole life of the mine. At a given depth, depending on the region and winter conditions, the soil should remain at a constant temperature (Rieger, 1921). If the depth is below the frost level, the temperature difference between the soil and glycol mixture should not be too large and the system might not require insulation. In the case that the soil temperature exceeds the glycol mixture, heat will be transferred from the soil to the fluid which would increase the efficiency of the system by using additional geothermal heat. When the trench is dug, a layer of small rocks should fill the bottom of the trench to minimize the piping movement, this procedure is called bedding. Underground piping was chosen to calculate the main piping system capital cost within the software. The procedure and assumptions to calculate the capital cost is described in the following. Trench digging First, a trench must be dug; the dimensions of the trench are as follow:
NPS in 2 to 10 12 to 18 20 to 24
Trench width: 4” larger than the pipe diameter to accommodate the insulation material of 1.5” thickness with an additional clearance of 0.5” on each side of each tube. The cold and hot side of the piping system is insulated and installed within the same trench and thus the width should be multiplied by 2. The depth of the trench is assumed from the NPS diameter of the piping system as shown in Table 5-3. Depth ft 3 4 4
Table 5-3: Assumed depth of trench with respect to the NPS diameter
Equation 5-17 is used to determine the total volume to be excavated. 2(d o 4" ) V 0.037 depth L 12
Eq. 5-17
Note: The 0.037 factor is to convert cubic foot into cubic yards For all digging work, a ½ yd3 excavator is assumed to be used. The trench excavation cost is calculated with respect to the total volume to be excavated and the soil nature. The cost data is shown in Table 5 of APPENDIX B. The different soil natures to choose from are: Common earth, Loam and sandy clay, sand and gravel and dense hard clay. Assuming one eight hours shift per day, the number of days required for the backhoe rental is estimated to be: H V N days Int exc +1 8
Eq. 5-18
53
The total rental cost is then found from daily rental cost of a ½ yd3 Backhoe which is of 500$ per day (RSMeans Co., 2008). Bedding Bedding is a layer deposited below the underground piping system inside the trench. It is required to minimize the movement of the piping system. The bedding type is crushed stone ¾” to ½”. The depth of the utility bedding is assumed from the NPS diameter of the piping system as shown in Table 5-4. NPS in 2 to 10 12 to 18 20 to 24
Depth in 3 6 12
Table 5-4: Depth of bedding with respect to NPS
The cost of bedding is calculated with respect to the total volume occupied determined from Equation 5-19. The cost data is found in Table 6 of APPENDIX B. V 0.037 WL depth Eq. 5-19 Backfill and compaction After the trench is dug and the bedding inserted, the earth that was previously removed to dig the trench will be shoved back and compacted in the trench to bury the piping system. The compaction will usually be performed with a vibrating plate. The labour time required to backfill and compact the earth is found from Table 7 of APPENDIX B, it is with respect to the total volume of the trench found previously from Equation 5-17. 5.07.2 Pipe supports: The piping system is located above ground at a certain height of the surface. In the case of a pipe or coupling failure it is much easier to identify where the failure has occurred and repair it as opposed to the underground piping system. The pipes are exposed to cold outside air therefore thick insulation must be used. In the case that the insulation is damaged or not properly installed, the heat loss would be significant when outside air reaches low temperatures. The pipe supports require foundations as shown in Figure 5-10.
54
Figure 5-10 : Pipe support schematic
Figure 5-11: Adjustable saddle with stanchion
First, the number of pipe supports required for the project must be known. For the analysis, the type of support chosen is the adjustable saddle as shown in Figure 5-12. The saddle is connected to a stanchion which is positioned on the pipe support foundations. The maximum load is dependent on the NPS which was determined from the maximum velocity within the main running pipe size as described in section 4.08. The Anvil© international’s saddle specification sheet provides the maximum load for a given NPS, the specification sheet is included in APPENDIX A. From the piping system’s weight and maximum load support, the maximum span between two supports is determined as well as the total number of supports required for the project. For carbon steel pipe, the weight per length of the pipe is calculated as follow:
W pipe 15.91t (d o t )
Eq. 5-20 (Engineering toolbox)
Then to calculate the weight of the glycol mixture per meters within the pipe; Equation 5-21 is used;
d 2 Wgly mix i 4
Eq. 5-21
Where the density of the glycol mixture is approximately 1100 kg/m3. The weight of the couplings should also be taken into consideration but it is presently unknown. The maximum span becomes; Lsp
Max.Load Wgly W pip
Eq. 5-22
The total number of supports required is the total pipe length divided by the maximum support span.
55
L nsup Int tot 1 Lsp
Eq. 5-23
The weight per length and maximum span for different NPS is shown in Table 5-5. NPS in
Weight kg/m
Maximum Span m
2
7.8
220
4 6 8 10 12 14 16 18 20 24
25.1 48.8 78.0 116.2 159.2 189.8 241.5 315.1 380.3 539.9
68.7 35.4 22.1 14.9 10.8 12.7 9.96 9.65 7.99 6.13
Table 5-5: Maximum span of pipe supports
If the insulation used is a zero water permeability material, the stanchion can be located very close to the surface; in this case, during winter, snow will cover the pipes and should act as an additional insulation. Moreover, the snow will eliminate the convection effect of the wind over the pipes. In the case that the insulation used has non-zero water vapour permeability, the snow could enhance the transmission of water within the insulation material which would decrease significantly the insulation properties of the material. Due to the latter, the length of the stanchion should be slightly above the level of the snow. The cost data of the supports is found in Table 9 of APPENDIX B. The installation and unit price calculation of the pipe support is found from Equation 5-24. S sup,f S sup,t nsup S sup,1 H sup Rlab 2
Eq. 5-24
The installation and material cost of the foundations requires to be added to the cost. 5.07.3 Pipe insulation Pipe insulation is required to limit the heat losses from the pipe to the surrounding atmosphere or soil. For pipe support design, both fluid sides (hot and cold) will always require insulation. For underground piping, the soil temperature will determine if the piping system requires insulation or not. It is dependent on the depth of the trench and the region, the deeper the trench, the greater the soil temperature will be. The temperature difference between the fluid and the soil can then determine if it requires insulation or not and if so the required thermal conductivity of the material. Calculation examples of heat losses for pipes are shown in section 4.13. On the cold glycol side, if the soil temperature and glycol temperature difference is low, insulation may not be required. It could also be possible to dig the trench deep enough to enhance the performance of the system and have the soil transferring additional geothermal heat to the fluid.
56
One of the main considerations in insulation is the accumulation of moisture in permeable material. For very low temperatures, it is recommended to apply two coats of vapour seal mastic reinforced with open weave glass or other fabric. The insulation should be sealed off every 15 or 20 ft to limit water penetration if the vapour seal gets damaged. As the piping system is exposed all year-round, a constant vapour drives exist under humid outside air conditions and moisture will inevitably accumulate in the insulation permeable material even if all precautions are taken (ASHRAE, 2009). Due to the latter, for permeable insulation, periodic replacement has to be performed. Foamglas® is a new insulation material that has zero water-vapour permeability. It does not require any coating, thus installation cost is reduced. Its life is said to be approximately 20 to 30 years with no maintenance required. It can also be used for underground piping (Foamglas, 2009). The data sheet of the material is found in APPENDIX A. It will be the assumed material used for the closed-loop glycol circuit. For underground piping, the thickness of the insulation used will be of 1.5” which is the minimal available size. From the NPS and Table 10 of APPENDIX B the labour and material cost can be found using Equations 5-25 and 5-26. Slab,ins Slab L pip H ins nelb H elb,ins Eq. 5-25
Smat,ins L pipSins nelb Selb,ins
Eq. 5-26
Note that the couplings and piping accessories insulation is not included within the cost. The totality of the piping system exposed must be properly insulated. Neglecting any small area exposed such as valves or couplings can greatly affect the performance of the system.
5.08
Pumps and electric motor
The pump is required to carry the fluid around the closed-loop system. The calculations performed to determine the required flow rate and head of the pump are shown in CHAPTER 4. The pump is driven by an electric motor. It is required to place the pump and electric motor units under stable foundations to ensure that they will keep their position throughout the whole life of the system. The electric motor requires a control center. The following will explain more into detail each of these components and also how the approximate capital cost has been determined. 5.08.1 Pump The pump is chosen from the total pressure drop across the system and the fluid flow rate. The data for the capital cost of the pump is determined from Table 11 in APPENDIX B. From the calculated pressure drop and flow rate; one or several pumps in series will be selected. First, from the volumetric flow rate of glycol, the range of available pump rated head size is determined. The program will choose the pump whose maximum flow rate in between the range of the rated flow rate of the two pumps, for example; if the flow is of 3000 gpm, this flow being between the 2000 and 5000 gpm pumps in Table 11, the chosen pump is the rated 5000 gpm. Subsequently, the available rated head determines the required pump and assigns its cost. A sample of the VBA code is shown in the following:
57
Select case Q (“Volumetric flow rate”) Case 2000 To 5000 If ∆P (“head”) < 100 Then Unit price = 13700$ ElseIf ∆P >= 100 And ∆P < 150 Then Unit price = 21400$ In the case that the required head is greater than the pump maximum rated head, for a given flow rate, several pumps will be assumed to be used in series. It will be assumed that the glycol flow rate will not exceed 10,000 gpm (631 l/s) as it is the maximum pump rated flow rate in Table 11 of APPENDIX B. Table 5-6 shows the types of pumps that are assumed to be used in series if required. Flow rate gpm
head ft
1000
200
2000
200
5000
100
10000
100
Table 5-6: Rated head chosen with respect to flow rate for pumps in series
The required head is divided by the rated head pump in Table 11 of APPENDIX B and it then returns its integer and adds 1 to it as in Equation 5-27. This number will be the number of pumps used in series. hsys N pump Int 1 h pump
Eq. 5-27
In the case that more than one pump is used, the unit price of the pump, electric motor, MCC (Motor control center) and motor feeder is multiplied by that number. From the chosen pump, the required electric motor is determined again using Table 11 of APPENDIX B. The capital cost of the electric motor is found from Table 12 of APPENDIX B. From the chosen electric motor, the MCC and motor feeder cost can be approximated. 5.08.2 MCC and motor feeder The MCC is necessary to start and stop the motor in a safely matter. It will also protect the motor from overloads and faults. Note that the cost of the MCC will be much greater in the case where a variable frequency drive (VFD) is used, it should be important to assess a feasibility study of the installation of a VFD. The VFD will vary the speed of the pump to decrease the flow rate of glycol in the case when the system does not need to run at full load decreasing the energy consumption. 58
The cost of the MCC components is determined from the power of the motor and Table 13 of APPENDIX B. The components included in the cost are as follow: Copper wire, Steel intermediate conduit, Magnetic FVNR, Safety switch fused, Safety switch non-fused, Flexible metallic conduit, Connectors, Coupling to conduit and Fuse cartridge non-renewable. The motor feeder is the electric cable delivering the power to the motor. Its cost will depend on the power of the electric motor as well as the distance between the pump and the main power supply. The cost per feet is found in Table 14 of APPENDIX B. In order to approximate the length of the feeder, the user must enter the two following inputs: “distance between main power supply and intake building” and “distance between main power supply and exhaust building”, the minimum value of these two inputs is chosen. There are some components missing in the capital cost such as the foundations and labour for pump and electric motor installation. Also, the cost of the electric motor and pump were taken from data of 2004, inflation was not included. These additional costs should not be too significant to the total capital cost. The cost calculations were performed conservatively due to the small array of material cost available for pumps and electric motors. Also the pumps included in (CostMine, 2004) are used to pump dirty water out of the mine; these pumps are more expensive as they require running in harsher conditions. Depending on the required head and flow rate of the system, the pump used can have a much lower capacity than what was chosen for the cost calculations. Especially in the case of pumps used in series, a larger pump is sometimes available which could in some cases reduce the capital cost. It is recommended to be the first component to be revised as it can largely differ from its actual cost.
5.09
Piping accessories
A piping system must be composed of several accessories which are necessary for functioning and safety of the system. They are usually not required in large numbers and the amount required is usually independent of the piping length. Each component will be briefly described, some of them may not be necessary and others could be missing. It is therefore very important for the designer to review properly each component and have the confirmation of a specialist that there are no important accessories missing and that the system should run properly. Most of the accessories required were found using (ASHRAE, 2004). The cost calculation for all accessories is dependent on the NPS. Some of the costs for larger pipe diameters were not included in (RSMeans Co., 2010) and have been assumed using a linear relationship with the smaller diameters accessories. The approximated costs are highlighted in the Tables of APPENDIX B and their linear relationship is included below the table. Some of the accessories are not included in the costing but they are still mentioned. Note that the costs of elbows and couplings are included in section 0 and the manifolds in section 5.04. These systems are therefore not included in this section. It is important to note that for each accessory, the required couplings to connect the piping system are included within the cost unless otherwise specified.
59
5.09.1 Strainer A strainer is used to remove particles within the fluid; although the system is closed loop, there is still contamination from the pump and erosion or corrosion of the piping system. The undesired particles can be removed using a strainer tee type as shown in Figure 5-12. The strainer must be removed and cleaned at a given interval depending on the accumulation of particles within the glycol. Its cost is found in Table 15 of APPENDIX B.
Figure 5-12: Strainer tee type
5.09.2 Air bleed lines It is important to bleed off the air accumulated within the piping; the actual cost of the system is not included in the feasibility study since the values were not found. The cost should not be too significant but the system must be implemented within the design. 5.09.3 Reducer Reducers are to connect two different components with different pipe diameters. The reducer is often required to connect the piping system with the pump since it will usually have a smaller pipe diameter than the main piping system. No reducers are included in the costing but the data can still be found in Table 24 of APPENDIX B. 5.09.4 Tee Tees are used in the case where the flow requires to be divided. Two types of tees are used; reducing tees which have one end with a smaller diameter and constant diameter tees where the three connections all have the same diameter. It is predicted that the total number of tees required for the system is four; one for the expansion tank, one for the filling tank (reducing tees) and two for the bypass valve at intake (constant diameter tees). The cost of the two different types is found in Table 16 and 17 of APPENDIX B. 5.09.5 Check valve This component is in place to ensure that the flow doesn’t reverse. Only one is required and it should be positioned downstream of the pump. The reversing flow can damage the pump. Its cost is found in Table 18 of APPENDIX B. 5.09.6 Butterfly valves Butterfly valves can be used to manually close the flow in the case of a leak or failure but will usually be used for the bypass valve which is described in section 5.11. It will be assumed that there are a total of 6 butterfly valves required. Its cost is found in Table 19 of APPENDIX B.
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5.09.7 Gate valve Gate valves are used to manually stop the flow. If the maximum pressure across the system is not too elevated, less expensive butterfly valves can be used. Within the software, it will be assumed that solely butterfly valves are used. It should be verified that the butterfly valve can withstand the maximum pressure within the system otherwise gate valves shall be used. The cost of gate valves is found in Table 21 of APPENDIX B. 5.09.8 Flange Flanges can be used instead of couplings to connect different accessories of the piping system. It will usually be used to connect the pipe to the pump. Since most of the accessories are assumed to have grooved joints, the flanges will solely be required for the pump and there will be a total number of two required for the system. Its cost is found in Table 20 of APPENDIX B. 5.09.9 Expansion tank The expansion tanks are used to compensate for the change in volume of closed-loop systems. There are several types of expansion tanks; open-air, closed-air and diaphragm tanks. Open air expansion tanks will be used for the design. It is important to note that the open-air expansion tank has to be located at greater height than the rest of the glycol system. In the case of the heat recovery system, it will most likely have to be located at a greater height than the intake or exhaust coils. The pipe connection should be upstream or downstream of the pump. To size the open-air expansion tank, Equation 5-28 is used (ASHRAE, 2004) : Vt Vs h 1 3T c
Assumptions: 60% ethylene glycol /water mixture Lower temperature: 0°C ( c =0.00090334 m3/kg) Higher temperature: 15°C ( h =0.00091075 m3/kg) : 11.7 x 10-6 m/ m K for steel For the assumed conditions: Vt 0.00768 Vs
Eq. 5-28
Eq. 5-29
As a general rule: in a closed loop system, there can only be one single expansion tank. The expansion tank can be compared analogically to a ground in an electrical system. The calculation of the volume occupied by the main piping system is calculated as follow: d 2 V pip o 2 1.15L Eq. 5-30 4 The volume occupied by the manifolds is determined assuming that each branch has a length of about 5 m with an inside pipe diameter of 4.03” (4” NPS) which is shown in Equation 5-31. d 2 Vmani 5 2 N coils o Eq. 5-31 4
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The volume occupied by each coil is: At intake; tube inner diameter; 0.577”, length of one tube; 222”, number of tubes; 276; 0.26 m3 At exhaust; tube inner diameter; 0.577”, length of tube; 216”, number of tubes; 368; 0.34 m3 The number of coils is then multiplied by the volume. The volumes are then added together to obtain the total volume of the system to then obtain the required volume of the expansion tank from Equation 5-29 that was derived from Equation 5-30 and 5-31. The total volume of the system will also be used to determine the cost of ethylene glycol. From Table 23 of APPENDIX B and the required volume of the tank, the cost of the expansion tank is taken from the maximum nearest volume available. In the case that the capacity required is greater than the largest available tank cost (400 gal.), Equation 5-32 and 5-33 have been derived from data of Table 23 assuming that the unit price and labour time have a linear relationship with the required volume of the expansion tank. For material cost: S mat,tan 10.2Vtan 156 Eq. 5-32 For labour time: H lab,tan 0.012Vtan 1.424 Eq. 5-33 5.09.10 Expansion joints Expansion joints are used to offset the total longitudinal thermal expansion of the piping system. The expansion of the piping system could be transmitted to the pump which could deform the casing and ultimately cause a failure. If pipes are located underground and insulated, the temperature of the pipes should remain relatively stable and therefore no expansion joints should be required since the soil temperature difference between the cold and warm months should be relatively low. In the case that pipes are located on surface; much greater temperature difference should be encountered. Especially if the system would stop operating for some reason during the cold months, if the non-operating period is long enough, the pipe could achieve the ambient temperature therefore the temperature range of the piping system should be considered to be from -40°C to 30°C. The following will explain the procedure in order to determine the total possible thermal expansion of the piping system. Example of pipe expansion calculations using the Laronde mine case; Using Table 5-7 and assuming the temperature range to be from 30°C to -40°C the expansion difference between the two temperatures is from 0.12 to -0.75 in per 100ft for a difference of 0.87. The total distance must be divided by 100 ft and multiplied by 0.87 inch to determine the required expansion joint of the system. The piping distance for one side is 230 m (754 ft) and thus the expansion is: 7.54* 0.87 in/100ft =6.6” The expansion joints shown in (RSMeans Co., 2010) have a capability of expanding 10”; one joint on each fluid side should therefore be enough to overcome the maximum expansion and contraction of the system. 62
Temp. Expansion °F In./100 ft -75 -1.00 -50 -0.84 -25 -0.68 0 -0.49 25 -0.32 50 -0.14 70 0.00 100 0.23 Table 5-7: Total linear thermal expansion for carbon steel pipes (Weldbend)
The cost of the expansion joints are found in Table 22 of APPENDIX B but are not included in the software.
5.10
Ethylene glycol
The ethylene glycol cost will depend on the percentage of the mixture required and the total volume occupied by the mixture. For usual Canadian weather, 60% ethylene glycol will be sufficient since its freezing temperature should be approximately -45°C. The lower the temperature is, the higher content of ethylene glycol is required within the mixture. For large quantities the cost is 6.90$/gallons (RSMeans Co., 2010). The cost is found from Equation 5-34 1 m3= 264.17 gallons
S gly X glyVtot S gly / gal
Eq. 5-34
The cost will also include an additional 10% of the volume required. To calculate the total volume of glycol mixture in the system, see section 5.09.9.
5.11
Bypass valve
The bypass valve is located at exhaust and is used to ensure that the glycol flow remains at a temperature above the freezing point at exhaust to avoid that the condensate freezes on the coils. The design of the system is as shown in Figure 5-13.
63
Figure 5-13: Bypass valve
The temperature sensor controls the pneumatic actuated valves to ensure that the temperature downstream of the recovery coils will always remain above 1°C. The butterfly valves control the resistance of the flow path and thus control the flow rate since for parallel flow: R1 (Q1) 2= R2 (Q2) 2 Eq. 5-35 The price of the pneumatic actuators has been guessed and therefore requires to be reviewed. Unit price; 3400$ Labour time; 8 hrs The cost of the temperature sensors and (PLC) programmable logic controls of the valve must also be included. It should also be noted that manual by-pass valves can be installed; it would enable the system to still run if the pneumatic system would fail. Back-up manual by-pass valves are presently being used at the Kiena mine (Dubois, 2009). The cost of the manual valves is not included within the software.
5.12
Automated Washing System
The automated washing system is used to clean the coils at exhaust in order to reduce fouling effects on the air-side of the HE. The system is implemented at the two mine sites that have the heat recovery system but both are presently non-functional due to mechanical problems. The Kiena mine heat recovery system operator mentioned that the washing system will get repaired and that it is well-worth running due to the reduction of heat transfer after several months of accumulated fouling. On the other hand, the Williams mine have chosen not to repair the automated washing system and to instead send some workers once a year to clean the coils with water compressors (Shaddock, 2010). The price of the automated washing system can be included or not within the cost calculations as is it an option for the user. The automated washing system design schematic is shown in Figure 5-14.
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Figure 5-14: Automated washing system schematic
The system is equipped with several components but the main concept is to have several nozzles located on the air side downstream of the coils. It is assumed that the water pressure delivered is sufficient to properly clean the coils otherwise; soap and/or water heating could be included within the design. To cover a greater area of cleaning, the nozzles are connected with a flexible hose as shown in Figure 5-15.
Figure 5-15: flexible hose for automated washing system
The pressure of the water creates a whip movement which increases the surface cleaned by a single nozzle. 5.12.1 Cost and geometry of the flexible hose and nozzles It is assumed that the geometry of the flexible nozzle is as follow: The flat spray nozzle Angle of 37.5° on both sides, Distance between the spray and the coil: 0.3 m.
65
The flat spray covers a length of ; 2 0.3 tan 37.5 = 0.46 m. The flexible hose: Length of 1.8 m Whip angle of 20° The whip of the flexible hose covers a length of; 2 1.8 tan 20 =1.31 m. The flexible hose length required is 2 m per nozzles; its cost is in Table 28 of APPENDIX B.
The total area covered by one nozzle: 0.60 m2 The face area of one coil: 9.6 m2.
Therefore one coil necessitates approximately 16 nozzles. To determine the number of nozzles required, the number of coils will be multiplied by 16. The material cost of the nozzles is of 81.25$ for 25 (Loctite, 2009). The nozzles are connected with a stainless steel ring clamp and an O-ring which are easy to install. It is assumed that each nozzle requires labour time of 0.06 hrs. They are pre-installed to the plastic tube. For each nozzle, a clamp is required to connect the flexible hose to the stainless steel tube. The material cost and labour time for one clamp is shown in Table 26 of APPENDIX B. It is not necessarily the required clamp but will be used for pricing purposes. 5.12.2 Piping system of spray nozzles branches The height of the coils is 24’ (7.3 m), 6 nozzles should be placed vertically in the same piping branch. There is approximately 1 m of piping required for each nozzle. Butt welds are performed to connect all joints. 2” stainless steel (SS) pipes will be used to minimize corrosion of pipes in contact with dirty foul air. The cost of SS pipes is in Table 27 of APPENDIX B. For each nozzle it is assumed that 1.2 butt welds are required. From (Page, 1999); a butt weld requires 0.4 hours of labour time for a 2” pipe diameter. The man hours include; cutting, bevelling, fitting, teck welding, manual single pass or backing ring, machine set-up and submerged welding. 5.12.3 Pump and electric motor The pump is used to deliver water to the nozzles with a sufficient pressure to remove accumulations on the coils. The sizing of the pump has been determined from the following assumptions. A common garden hose will usually deliver a pressure of 40 psi, having a pressure of 60 psi (150 ft) is assumed to be sufficient to properly clean the coils. The flow rate of 4 gpm is approximately the flow rate of a shower head which should be enough when covering a length of of 0.6 m for flat spray nozzle. It will be assumed that a 200 gpm pump will be used; 50 nozzles can spray water at the same time. 66
The cost of a pump of 200 gpm and 150 ft of head is found in Table 11 of APPENDIX B. The pump requires a 100 hp motor. For the motor, the cost is found from Table 12 of APPENDIX B. The cost of the MCC component for the motor is found from Table 13 of APPENDIX B. 5.12.4 Valves Within the SS piping system, several groups of nozzles are isolated with motor actuated threeway valves. The valves are installed to reduce the required capacity of the pump. During the spraying operation, only one valve is opened while the others remain closed. The pump only delivers water to that group of nozzles while the other ones remain inactive. As the pump has terminated cleaning the section of coils, the valve closes and another one opens. The nozzles are placed in a parallel arrangement; the volumetric flow rate provided by the pump is divided equally for each separate nozzle and the delivered pressure is the same for all nozzles. The sequencing and frequency of the nozzles spraying operation will depend on the level of fouling and should be adjusted according to the conditions of exhaust air. The flow velocity within a 2” diameter pipe for a volumetric flow rate of 200 gpm is of 2.6 m/s; no erosion problems should be encountered. For each 50 nozzles one valve is required, the number of valves required is calculated from Equation 5-36. N N val Int noz 1 50
Eq. 5-36
The cost of the electric motor actuated valves is in Table 29 of APPENDIX B. The valves will be controlled by a PLC, its cost is not included in the software. Assuming that the motors used are 1 hp motors, the motor feeder cost is determined from the distance between the exhaust ventilation installations and the main power supply of the mine site. The cost of the motor feeder is found in Table 30 of APPENDIX B. 5.12.5 Trench for Piping from main water supply to exhaust The trench is assumed to be 2’ deep and 2’ wide. The cost for trenching a 2’ x 2’ trench is in Table 25 of APPENDIX B. The bedding at the bottom will be assumed to be 2” deep and the cost is calculated using Table 6 of APPENDIX B. 5.12.6 Piping from main water supply to exhaust The piping cost procedure will be the same as in the underground piping section except that solely 2” diameter piping will be used, the piping length is required only on one side. It will be assumed that the system is composed of 10 elbows.
5.13
Other systems not included
Drainage system The condensation of water vapour on the coils will create undesirable flooding around or inside the surface ventilation installations. The water will most likely be dirty and therefore should be 67
sent to a contained pond to minimize the environmental impact. The drainage system cost will depend on the distance between the exhaust and the contained pond. Filling tank In the case of leaking or servicing of the components of the system, it could be required to empty and re-fill the piping system. In this case, a filling tank with a pump should be put in place.
5.14
Economy of the project
When calculating the capital cost, the material and labour costs are added separately together. As mentioned earlier the labour cost is determined from a constant labour rate entered by the user. The material cost is affected by several factors which are added in that same order; the inflation since 2010, the location factor on the material cost (location factors from Canadian cities are found in APPENDIX A), the overhead and profit on material (it was suggested by RSMeans Co. to use 10%). Then the labour cost is added to the material cost to include the engineering fees and then the contingency to finally obtain the total approximate cost of the project. The total project capital cost is then divided by the net annual energy cost savings to find the payback period. In order to better understand the distribution cost of the main components, different case studies have been evaluated. The cost components have been separated into the following; HE; includes the coils, its support, its foundations and the building extension to reduce air velocity. Pipe: main piping system, accessories and ethylene glycol Pump: Pumps and motors Wash; Automated washing system cost The distribution of the cost is shown in percentage of the total cost of the project (including solely material and labour cost). The different case studies evaluated are as follows: Case 1 (Fig. 5-16) Airflow: 200m3/s, distance: 200m Case 2 (Fig. 5-17) Airflow: 200m3/s, distance: 500m Case 3 (Fig. 5-18) Airflow: 200m3/s, distance: 1000m Case 4 (Fig. 5-19)Airflow: 400m3/s, distance: 200m Case 5 (Fig. 5-20) Airflow: 400m3/s, distance: 500m Case 6 (Fig. 5-21) Airflow: 400m3/s, distance: 1000m Case 7(Fig. 5-22) Airflow: 600m3/s, distance: 200m Case 8(Fig. 5-23) Airflow: 600m3/s, distance: 500m Case 9(Fig. 5-24) Airflow: 600m3/s, distance: 1000m
68
Figure 5-16: Cost distribution of components case 1
Figure 5-17: Cost distribution of components case 2
Figure 5-18: Cost distribution of components case 3
Figure 5-19: Cost distribution of components case 4
Figure 5-20: Cost distribution of components case 5
Figure 5-21: Cost distribution of components case 6
Figure 5-22: Cost distribution of components case 7
Figure 5-23: Cost distribution of components case 8
69
Figure 5-24: Cost distribution of components case 9
It is possible to observe that the HE and piping system consist of the major cost of the system. For relatively short distances between intake and exhaust (200 m), the HE cost will be greater than the piping cost and the opposite for a distance of 500 m or greater. This study shows that the distance between intake and exhaust will affect significantly the project’s capital cost.
5.15
Case studies
Several scenarios using the feasibility study software tool will be evaluated in this section. The results presented are gross annual energy savings, net annual energy savings and payback period. The gross annual energy savings are the total savings in heating fuel. The net annual energy savings are the total savings in heating fuel minus the total operational costs which includes the cost of air pressure loss across the coils and the pumping of glycol-water mixture across the loop. The payback period is the capital cost divided by the net annual energy savings. The following parameters will remain constant: Exhaust ventilation air temperature at: 13°C, 100% humidity The exhaust air temperature will be mostly dependent on the depth of the mine as geothermal heat from strata rock will be transferred. One of the deepest mines in Canada (Laronde) can reach exhaust air temperature up to 18°C during the winter period (Lafontaine, 2008). Intake air temperature set point: 1.5°C (after heating) Intake fresh air heating is mainly performed to avoid ice accumulation within the intake shaft. Commonly the set point will be close to 0°C but can however be set to a higher level to increase the comfort of miners or in the case that ambient air is entering by other means such as the main shaft collar. Electricity cost: 0.08$/kWh. The electricity cost is used to determine the operating costs of the system. For most Canadian mines it will often be around this value except for mines where no power lines are available and diesel generators are used where the cost will be significantly higher. Labour rate: 75 $/hr, Location factor: 111.4 (Sudbury), Contingency: 10%, Overhead and profit on material: 10%, Engineering fees: 10%
70
The following parameters will be varied: Price of fuel: 8$/GJ, 13$/GJ & 20$/GJ The price of fuel will differ greatly from one region to another. In some regions such as Sudbury Ontario, natural gas is relatively cheap, the rate is approximately 8$/GJ (Sabau, 2010). On the other hand, in Abitibi Quebec, the price of natural gas could go up to 13$/GJ (Girard, 2010). In remote locations where no natural gas pipelines are available, propane must be carried and stored at the mine site. Propane is usually a much more expensive option, in 2009, the average Canadian propane price was over 20$/GJ (NRCAN, 2010). Temperature of region: Three different Canadian locations were considered in the case studies: Fort Simpson in Northwest Territories, Rouyn-Noranda in Quebec & Smithers in British Columbia. Average weather data of the past 5 years have been used to forecast monthly temperatures in each of these regions. Fort Simpson, NWT is the coldest and Smithers, BC has the mildest climate of the three. Mine ventilation intake and exhaust airflow: 200 m3/s, 400 m3/s & 600 m3/s. Distance between intake and exhaust shafts: 200 m, 1000 m & 2000 m Figures 5-16 to 5-24 show the summary results of the feasibility study software tool for each of the following nine case studies. The estimated capital cost by the software is included in the title of each Figure. Gross and net energy savings are also shown on the bar charts. Case 1 (Fig. 5-25) Airflow: 200m3/s, distance: 200m, fuel price: 8$/GJ, 13$/GJ & 20$/GJ Case 2 (Fig. 5-26) Airflow: 200m3/s, distance: 1000m, fuel price: 8$/GJ, 13$/GJ & 20$/GJ Case 3 (Fig. 5-27) Airflow: 200m3/s, distance: 2000m, fuel price: 8$/GJ, 13$/GJ & 20$/GJ Case 4 (Fig. 5-28)Airflow: 400m3/s, distance: 200m, fuel price: 8$/GJ, 13$/GJ & 20$/GJ Case 5 (Fig. 5-29) Airflow: 400m3/s, distance: 1000m, fuel price: 8$/GJ, 13$/GJ & 20$/GJ Case 6 (Fig. 5-30) Airflow: 400m3/s, distance: 2000m, fuel price: 8$/GJ, 13$/GJ & 20$/GJ Case 7(Fig. 5-31) Airflow: 600m3/s, distance: 200m, fuel price: 8$/GJ, 13$/GJ & 20$/GJ Case 8(Fig. 5-32) Airflow: 600m3/s, distance: 1000m, fuel price: 8$/GJ, 13$/GJ & 20$/GJ Case 9(Fig. 5-33) Airflow: 600m3/s, distance: 2000m, fuel price: 8$/GJ, 13$/GJ & 20$/GJ Cost data of return air heat recovery system (Q: 200m3/s, dist.: 200 m, capital cost: 1,196,000$)
Cost data of return air heat recovery system (Q: 200m3/s, dist.: 1000 m, capital cost: 2,700,000$) $2,500,000
$2,500,000 14 $2,000,000
12
39
25
15
15
14
$2,000,000
12 10
10 $1,500,000
$1,500,000
6
yrs
$1,000,000
$1,000,000
6
yrs
8
8
4
4 $500,000
$500,000
2
2 $0
0 BC
Qc 8 $/GJ
Payback period (yrs)
NWT
BC
Qc NWT 13 $/GJ
Gross energy savings ($/yr)
BC
Qc NWT 20 $/GJ
Net energy savings ($/yr)
Figure 5-25: Heat recovery system economics, case 1
$0
0 BC
Qc 8 $/GJ
Payback period (yrs)
NWT
BC
Qc NWT 13 $/GJ
Gross energy savings ($/yr)
BC
Qc NWT 20 $/GJ
Net energy savings ($/yr)
Figure 5-26: Heat recovery system economics, case 2
71
Cost data of return air heat recovery system (Q: 200m3/s, dist.: 2000 m, capital cost: 4,700,000$)
87
52
30
29
21
15
$2,500,000 14
$2,000,000
12
14 $2,000,000
12
10 $1,500,000
10 $1,500,000
6
8 $1,000,000
6
4
4
$500,000
$500,000 2
$0
2
0 BC
Qc 8 $/GJ
NWT
Payback period (yrs)
BC
Qc NWT 13 $/GJ
Gross energy savings ($/yr)
BC
$0
Qc NWT 20 $/GJ
0 BC
Net energy savings ($/yr)
NWT
BC
Qc NWT 13 $/GJ
Gross energy savings ($/yr)
BC
Qc NWT 20 $/GJ
Net energy savings ($/yr)
Figure 5-28: Heat recovery system economics, case 4
Cost data of return air heat recovery system (Q: 400m3/s, dist.: 1000 m, capital cost: 4,100,000$)
33
Qc 8 $/GJ
Payback period (yrs)
Figure 5-27: Heat recovery system economics, case 3
Cost data of return air heat recovery system (Q: 400m3/s, dist.: 2000 m, capital cost: 6,784,000$)
20
$2,500,000 14
$2,000,000
12
75
43
23
22
16
14
$2,000,000
12
10 $1,500,000
10 $1,500,000
$1,000,000
6
8 yrs
8 $1,000,000
6
4
yrs
$2,500,000
yrs
$1,000,000
yrs
8
4
$500,000
$500,000 2
$0
2
0 BC
Qc 8 $/GJ
NWT
Payback period (yrs)
BC
Qc NWT 13 $/GJ
Gross energy savings ($/yr)
BC
$0
Qc NWT 20 $/GJ
0 BC
Net energy savings ($/yr)
Qc 8 $/GJ
Payback period (yrs)
Figure 5-29: Heat recovery system economics, case 5
NWT
BC
Qc NWT 13 $/GJ
Gross energy savings ($/yr)
BC
Qc NWT 20 $/GJ
Net energy savings ($/yr)
Figure 5-30: Heat recovery system economics, case 6
Cost data of return air heat recovery system (Q: 600m3/s, dist.: 200 m, capital cost: 2,846,000$)
Cost data of return air heat recovery system (Q: 600m3/s, dist.: 1000 m, capital cost: 5,478,000$)
$2,500,000
$2,500,000 14
$2,000,000
12
36
21
14
$2,000,000
12
10 $1,500,000
10 $1,500,000
6
8 yrs
8 $1,000,000
$1,000,000
6
4 $500,000
yrs
$2,500,000
Cost data of return air heat recovery system (Q: 400m3/s, dist.: 200 m, capital cost: 2,000,000$)
4 $500,000
2 $0
0 BC
Qc 8 $/GJ
NWT
Payback period (yrs)
BC
Qc NWT 13 $/GJ
Gross energy savings ($/yr)
BC
Qc NWT 20 $/GJ
2 $0
0 BC
Qc 8 $/GJ
Payback period (yrs)
Net energy savings ($/yr)
Figure 5-31: Heat recovery system economics, case 7
NWT
BC
Qc NWT 13 $/GJ
Gross energy savings ($/yr)
BC
Qc NWT 20 $/GJ
Net energy savings ($/yr)
Figure 5-32: Heat recovery system economics, case 8
Cost data of return air heat recovery system (Q: 600m3/s, dist.: 2000 m, capital cost: 8,800,000$)
$2,500,000
93 43
23
21
14 $2,000,000
12 10
$1,500,000
6
yrs
8 $1,000,000
4 $500,000 2 $0
0 BC
Qc 8 $/GJ
Payback period (yrs)
NWT
BC
Qc NWT 13 $/GJ
Gross energy savings ($/yr)
BC
Qc NWT 20 $/GJ
Net energy savings ($/yr)
Figure 5-24: Heat recovery system economics, case 9
72
Through the analysis of the results, it is possible to observe that long distances between intake and exhaust shafts and low fuel price will often result in a long payback period of the project. The greater volumetric flow rates will result in a higher capital cost of the system as the pipes and HE will have to be of larger size. Due to the latter for low fuel cost and long distance between shafts; the payback period will be longer for greater flow rates. On the other hand, for high fuel cost and short distance between shafts, the payback period will be shorter than for lower flow rates as energy savings will be more significant. It is very important to mention that small variations within any of the input parameters can significantly affect the results. Before arriving to any conclusions, the software shall always be used for a given operation. The results previously shown should be only used as a rough guide and a more extensive study of the actual parameters should be performed. Relatively long distances between intake and exhaust shafts (1000 and 2000 m) were chosen to demonstrate that this parameter should not be a decisive point in choosing to perform a study for an exhaust air heat recovery system. From various interviews with people from the industry, it seems that the decision to study the feasibility of installing an exhaust air heat recovery system was somehow arbitrary and that many additional mines should have done so but have come to quick conclusions that it would not be feasible simply by looking at the distance between intake and exhaust shafts. The capital cost is still a rough estimate and several components of the cost should be reviewed. It is expected that total capital costs obtained using the software will be higher than its actual cost as several conservative assumptions were made within the calculations even though some components are not included in the costing.
5.16
Conclusions
Most of the design of the return air heat recovery system has been covered in this section. The main components of the system are as follows: heat exchangers and their installations, main piping system, pumps and electric motor, piping accessories, manifolds and automated washing system. The system has been studied carefully but the whole design should still be reviewed by specialists of closed-loop piping systems. Closed-loop systems can be dangerous if not designed properly; the system should be reviewed prior to the start of the construction to ensure a safe operation. It is expected that total capital cost obtained using the software will be higher than its actual cost as several conservative assumptions were used within the calculations even though some components are not included in the costing. Several costs should be re-evaluated, especially for the pumping system. The cost calculation software tool can also be modified to obtain the costing of similar types of projects such as dewatering systems and refrigeration plants. The developed software tool is subject to improvements and recommendations and any users are encouraged to do so. If a heat recovery system project is implemented, the tool could be improved by comparing the real cost from the results of the software. It could also be used in the development stage of a mining operation where it could influence the positioning of the shafts. 73
CHAPTER 6. ALTERNATIVE DESIGNS OF THE HEAT RECOVERY SYSTEM Summary This Chapter presents designs of heat recovery systems other than the closed-loop glycol circuit. Heat sources other than exhaust air available at underground mine sites are listed. The possibility of heating surface building is as well discussed.
6.01
Introduction
The closed-loop glycol circuit is the simplest design of a heat recovery system and it is presently being used at the Kiena and Williams mines to recover heat from exhaust air. Several interesting designs that could substitute the closed-loop glycol circuit are presented in this Chapter. Some designs have been studied and they were found to be not feasible in most cases. They are although still presented to avoid future unnecessary research or in case that the reader finds them suitable for a specific operation. Other heat sources than return air have the potential to be recovered. These alternative heat sources are outlined. The heat recovery system could be also used for space heating of the surface buildings instead of heating mine ventilation fresh air; some proposed ideas are presented. Several different options of the heat recovery system will also be discussed in this Chapter. Technical information and calculations are included as well.
6.02
Recovering heat from the depths of the mine
As mines get deeper, refrigeration may be required. There are presently three underground Canadian mines that require cooling; Laronde in Abitibi, Qc, Kidd Creek in Timmins, Ont. and Creighton in Sudbury, Ont. Underground mine air cooling is described into detail in Chapter 8. The idea of recovering heat from the depths was developed for the present situation at the Laronde mine: due to ice accumulation problems air needs to be warmed using natural gas burners at the surface to a temperature of 1°C. During cold periods, as cold fresh air is carried underground, it picks up enough heat to require mechanical refrigeration to improve the working conditions at lower levels. In other words, the air is first heated to be cooled later which results in elevated energy costs. The plant operates at a power of 400 tons of refrigeration (1405 kW) during winter and at 1000 tons of refrigeration (3517 kW) during summer (Quirion, 2009). One of the advantages of removing heat at the lower levels is that the heat recovery system would act as a refrigeration plant and therefore the use of the mechanical refrigeration plant during cold periods would most likely not be required. Moreover, due to auto-compression, the air enthalpy is much greater which enhances heat transfer. Unfortunately, the main disadvantage of extracting heat at deep levels is the use of expensive insulation combined with vapour barrier for chilled water pipes. The pipes must also be high pressure resistant due to the increased water pressure with depth. Carrying the fluid through long distances would also generate a large amount of friction losses which would have to be overcome with a pump and consequently 74
increase operating and capital costs. Due to the latter, implementing such a system would generate a high capital cost making it very difficult to obtain a short payback period. The system would most likely not be feasible if it were used solely for a heat recovery application (no cooling is required at the mine site). In the cases similar to the Laronde mine (cooling required during winter period), if the cooling load required becomes extensive, the system might be interesting to look at. The schematic of the design is shown in Figure 6-1.
Figure 6-1: Schematic of heat recovery system from the depths of the mine
6.03
Heat pump, evaporator at exhaust and condenser at intake
This proposed design consists of having the evaporator as the tube and fin HE at intake and condenser as the tube and fin HE at exhaust. The heat pump is necessary to recover heat from a lower to a higher grade heat source as in a usual geothermal heat pump; the earth ground at which the heat is recovered has a lower temperature than the ambient temperature of the house which requires heating. The heat pump can be defined as a reverse cooling cycle. The cooling cycle is described in Chapter 8. The heat recovery system does not require the use of a heat pump as it recovers heat from a higher to a lower grade heat source. The use of external mechanical work (compressor) is therefore not required and inefficient. Also, downstream of the exhaust, low density vapour must be carried to the intake which would require the use of expensive large diameter pipes to reduce friction losses. Due to the latter, this system is considered to be inefficient and not feasible.
6.04 Spray chambers at exhaust, tube-fin HE at intake, plate heat exchangers to transfer the heat 75
The following would be the most interesting alternative design to use. The system could extract heat at exhaust with the use of a direct-contact HE. After the heating process of water at exhaust, the water would transmit its heat with the use of a plate heat exchanger to ethylene glycol mixture which would then be carried at a tube-fin HE to heat the intake air. The schematic of the design is shown in Figure 6-2.
Figure 6-2: Diagram of heat recovery system with plate heat exchanger
By varying the flow rate of the glycol, it is possible to keep the water temperature above the freezing point and still obtain the same heating load; no bypass valves would be required as in the closed-loop glycol circuit. This design requires having a minimum of two pumps, one for the water circuit and the other one for the glycol. 6.04.1 Direct contact HE The direct contact HE would eliminate the fouling problem encountered with the tube-fin HE. It would also act as an air cleaner as particulates within the air will tend to mix with water making the heat recovery system more environmental friendly than it already is. Of course, as the air is continuously cleaned, the water continuously gets dirty; the use of a filtration system would be required. Another great advantage of the direct contact heat exchanger (spray chambers) is the negligible pressure drop on the air-side as opposed to the tube-fin HE. Direct contact HE will usually come in two different types; towers or spray chambers. Towers are mostly used to cool water, they are usually known as cooling towers. A natural or forced draft of air comes in contact with the water which evaporates it and transfers its latent heat to the outside air. The amount of water evaporated will determine the cooling load. The specific heat of water vaporisation is 2257 kJ/kg, thus for each kg of water evaporated, the water will transfer 2257 kJ of latent heat to the 76
air. Some of the heat will also be transferred through sensible heat depending on the water and air temperature. For the heat recovery system, the tower will be used to heat the water instead of cooling it and it should therefore be called a heating tower. Towers are interesting but will most likely still encounter fouling problems on the fill packing of the tower which is used to increase the exposed surface area of the water (Bourret, 2009). Due to the latter, spray chambers should be used instead. From the usual geometry of the surface ventilation installations, the spray chambers should have a cross flow arrangement as in Figure 6-3. Horizontal spray chambers are essentially cross flow heat exchangers in which water is sprayed upwards or downward and the air flow is horizontal.
Figure 6-3: Crossflow horizontal spray chambers: low water loading
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Figure 6-4: Schematic of two-stage cross flow horizontal spray high water loading
To increase the efficiency, it is possible to have two-stage spray chambers; it implies that the water is re-circulated twice through the air as in Figure 6-4. Note that in Figure 6-3, the horizontal spray chamber is defined as low water loading and in Figure 6-4 as high water loading. Low water loading will be more efficient as the water will be in contact with the air for a longer period of time. Unlike normal heat exchangers, direct contact heat exchanger’s performance is not characterized by efficiency but by its factor of merit. Typical factors of merit are suggested by (Whillier, 1977) and (Bluhm, 1981). The following equations will describe on how to calculate the outlet water temperature from an assumed factor of merit. *
E FR Where: F : Factor of merit
R * : Tower capacity factor If R=1, a E Tw,in Tw, out Where w Water efficiency Tw,in Ta ,in
Eq. 4-1
Eq. 4-2
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a
Ta,in Ta,out Ta,in Tw,in
Air efficiency
Eq. 4-3
Tw,in Ta,in m R a w C p, w (McPherson, 1993) w S w,in S a,in
Eq. 4-5
ma
Where: S w,in : Sigma heat of saturated air at same temperature as the inlet water (J/kg dry air) S a,in : Sigma heat of inlet air at exhaust (J/kg dry air) Tw,in : Inlet water temperature
Ta ,in : Inlet air temperature C p, w :
mw :
Specific heat capacity rate of water at constant pressure
Mass flow rate of water
: Mass flow rate of air Lw (2502.5 2.386Twet )
ma
Sin LwW 1005Twet Where: Twet : Wet bulb temperature of inlet air (°C) W : Moisture content of dry air (kgmoist/kgdry air) In order to have a better idea of the physical size and cost of the spray chambers, an example from a 15 MW cooling plant using horizontal spray chambers to cool the air will be described (Bluhm, Funnell, & Smit, 2001). A heat recovery system with a 15 MW heating capacity can be assumed to be of similar size. The air speed is 5 m/s, the face area of the building enclosing the spray chambers must be sized according to the exhaust volumetric flow. The 15 MW cooling plant spray chamber requires a 7 m high building with a plan area of 540 m2. Mist eliminators are required at the outlet of the spray chambers building to ensure that all the water droplets remain in the system. The spray chambers capital cost was found to be of 810,000$US back in 2001. 6.04.2 Filtration system A filtration system was proposed by (Howes, 2010) to have a water tank equipped two outlets. One outlet is connected to the main water tank; the other one is connected to an enclosed partition within the tank filled with sand. The schematic of the system is shown in Figure 6-5.
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Figure 6-5: Schematic of filtration system
The two outlets are coupled to a motor actuated two way valve. When the system does not require filtration, valve 1 opens and valve 2 closes. For the filtration operation, valve 1 closes and valve 2 opens; the water within the main tank overflows in the enclosed partition. As the water flows through the filtering sand, undesired particulates are removed. 6.04.3 Plate Heat Exchanger Plate heat exchangers will generate a high turbulence flow which will decrease the fouling effect on the plates (Shah & Sekulic, 2003). They are also easy for maintenance and cleaning. These features are very important since the contaminated water could result in heavy fouling. A plate heat exchanger picture and schematic is found in Figure 6-6.
Figure 6-6: Plate heat exchanger (Made-in-China.com, 2010) (IQS inc.)
Underground mine refrigeration plants using spray chambers will usually have a PHE as the evaporator and condenser heat exchangers between the refrigerant and the water. The plate heat exchangers pressure drop calculations are described in the following:
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To determine the pressure drop across the PHE, from (Shah & Sekulic, 2003) Equation 6-6 is used: P
1.5G p n p 2
4 fLG 2 2 De
Eq. 6-6
The first part is the pressure drop associated with the inlet and outlet manifolds and ports. The second part is the pressure drop within the core (plate passages). Where:
Gp
m ( / 4) D 2p
m G A0 A0 N p w b
Re
GDe
Where b : Distance between chevron plates D p : Diameter of ports De : Hydraulic diameter between chevron plates
f 0.8 Re 0.25 : friction factor L : Length of chevron plates
m : Mass flow rate of fluid N p : Number of fluid passes
n p : Number of ports w:
Width of chevron plates : Fluid density
To determine the outlet fluid temperatures, the HE effectiveness is determined by the manufacturer. A quote of a plate heat exchanger has been provided by Thermofin©, located in Candiac, Qc, Canada. The quote was requested for a 9 MW heat recovery system. The designer determined that the system would require 4 plate heat exchanger units, 2.25 MW and 100 kg/s on each fluid side per PHE. The cost of one PHE is of 40 000$CAN. The effectiveness of the PHE is of 0.78. The technical information of the PHE quote is found in APPENDIX A, note that the document is in French.
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6.05 Heat pump from a refrigeration plant; Direct-contact HE at exhaust, glycol tube and fin HE at intake The design is similar to the previous one except that the PHE is replaced by a surface refrigeration plant. The schematic of the system is shown in Figure 6-7. It is recommended to first read section 9.01 in order to understand better the design.
Direct contact HE at exhaust
Figure 6-7: Heat recovery system with the use of a refrigeration plant
A similar proposal was first evaluated at the Kidd Creek mine located in Timmins, Ont. Canada to use their refrigeration plant on the surface to cool the return air and warm fresh intake air in cold periods (Howes & Hortin, 2005). The system would use towers at intake and exhaust. At the intake, to avoid water freezing, a natural gas burner would be placed upstream of the heating towers to warm air at a temperature of -2.0/2.0 °C wet/dry bulb. The heated flow of air would then flow through the heating tower to achieve an approximate air temperature of 27.3°C at full load. The hot air would then mix with fresh outside air to maintain a temperature of 1.0°C at the intake. Unfortunately, the project was found to have a low rate of return and was rejected. The system was evaluated not to require additional heating if ambient air was over -28°C wet bulb not considering the pre-heating of the air upstream of the cooling tower. The difference between the design proposed by (Howes & Hortin, 2005) from the one in Figure 6-7 is the use of a tubefin heat exchanger with glycol at intake to replace the cooling tower. It would disable the necessity of pre-heating the air upstream of the HE. Also, the heat transfer between fresh air and ethylene glycol would be greater due to the increased temperature difference between the two fluids. The system is interesting since the refrigeration plant would be used to heat and cool the air. The design can be implemented in a way that when the cooling is no more required, the cycle is reversed with the use of valves to the heating use. In cooling mode, it should be verified if it is feasible to reject the heat at exhaust. The forced draft induced by the fans could reject the heat more efficiently than with the cooling towers. However it is important to consider that the pressure drop to carry the water to the exhaust can increase significantly the operating costs of the cooling system. Also, the exhaust air is usually saturated and therefore solely sensible heat transfer will occur which will eliminate the benefits of the latent heat rejection. Due to the latter, in some cases it could be more efficient to use cooling towers located at proximity to the intake instead of rejecting the heat back at the exhaust. 82
The advantage of the refrigeration plant is that the fluid temperature at intake can be greater than with the closed-loop glycol circuit reducing the required flow rate of ethylene glycol at intake. On the other hand, since the fluid temperature at exhaust must always remain above the freezing point, the refrigeration plant will not increase the efficiency of heat recovery. Due to the latter, the only possibility of justifying the capital cost of the refrigeration plant for the unique application of heat recovery (i.e. no cooling required) would be as follows. As the heat will be recovered, the hot glycol temperature would be increased to a certain extent. The greater the glycol temperature is, the lower the flow rate of glycol is required. In other words,
increasing the ΔT component in the m C p T equation will enable the possibility to decrease the
m (mass flow rate) and thus reduce the capital and operating cost of the piping and pumping
system. Therefore if the costs are reduced to an extent where it would offset the capital and operating cost of the refrigeration plant, the system could be feasible. However it is very important to perform all energy calculations especially for the efficiency of the refrigeration plant to transfer the compressor work into heat at the intake. A disadvantage of the design is that the heat recovered is limited to the power of the refrigeration plant. In the case that the possible power recovered at exhaust is greater than the refrigeration plant capacity, it could be more feasible to use a PHE as in the design previously presented.
6.06
Re-circulation of return air
Re-circulating exhaust warm air into the intake fresh air will work as a heat recovery system. Controlled re-circulation of air is usually performed to reduce ventilation costs as it increases the airflow for the same energy input. It can also be used to reduce heating costs during winter. However, exhaust air re-circulation is prohibited for most mining regulations as it can create excessive dust concentrations and high gas levels. In Canada, regulations regarding re-circulation differ for different provinces. Controlled re-circulations can be adequate if several safety precautions are taken such as automatic cleaning systems, fire control and air quality monitoring stations. A study case performed in Canadian Potash mines is described in (Hall, Mchaina, & Hardcastle, 1990).
6.07
Heat sources other than exhaust mine air
Other heat sources than mine return air can be found on a mine site, the heat sources can be used independently or they can be combined within a system with other heat sources. The two first heat sources presented are the most common to be found. As every mine site is different, there can be some available potential heat sources that are not mentioned in this work, it is therefore important to analyze all possible options when performing an energy assessment. 6.07.1 Mine water heat recovery Some mines pump large quantities of water to the surface in order to avoid flooding. If a sufficient amount of water is pumped, it could be feasible to recover some of the heat contained in the water. The total amount of heat that can be recovered will depend on the water flow rate and its temperature. Mine water is usually very dirty and acidic; it is thus important to use a plate heat exchanger for heat transfer as it limits fouling and is easy for maintenance. If mine water is to be carried inside the coils at intake, the pipes would require expensive maintenance and 83
necessitate great wall thickness due to heavy corrosion from the acidity which would result in a reduction of heat transfer. The proposed design is as follows: water transfers heat to an ethylene glycol mixture with the use of a PHE, the glycol is then carried at the intake air to discharge the heat with the use of a tube and fin HE. The system must be equipped with sensors and controlled valves to bypass the intake coils at low temperature in order to ensure that the water never reaches the freezing point 6.07.2 Heat recovery of mine air compressors In underground mines, air is compressed with the use of volumetric piston cylinders. As ambient air is compressed, its relative humidity increases until it reaches saturation and then condensation occurs. The presence of water in the compressed air is the source of several problems such as corrosion of the piping system and reduction in the compression cycle efficiency (Lrimie, Lrimie, & Tulbure, 1996). In order to remove humidity, air is cooled so that water vapour condenses prior to the compression cycle. The heat from the air compressor cooler is usually rejected with the use of cooling towers. During the cold periods, instead of discharging the heat at cooling towers, it could be discharged to the intake fresh air. Heating coils at intake could be installed to reject heat during cool periods. During the warmer months, heat cannot be discharged at intake as it is undesirable to heat the intake air during that period. If the heat recovery system is combined with other heat sources such as the exhaust air, it is possible to use other heat sources as a heat sink during the warmer months provided that they are of a lower grade than the heat from the compressor coolers. For example, exhaust air or mine water will usually have a lower temperature than compressor coolers and these could be therefore used to cool the compressors. 6.07.3 Geothermal ground heat pump Geothermal heat is usually of low grade and thus requires to be extracted with the use of a heat pump. Common residential or commercial geothermal heat pump systems could be adapted for underground mine sites. The different types of heat pumps are listed as follows (Raymond, Therrien, & Gosselin, 2010): Ground water heat pump: Groundwater is pumped out of the soil, heat is recovered with the use of a heat pump and the water is then discharged in a pond on surface. Surface water heat pump: o Open-loop: Water is pumped from a lake or a pond, heat is extracted from a heat pump and water is discharged back to a lake. o Closed-loop: Indirect contact heat exchanger is located in a lake or pond and transfers heat to a medium such as glycol which is carried around the loop, heat is then extracted from another heat exchanger connected to a heat pump. Ground-coupled heat pump: it is a closed-loop system in which the ground transfers heat to a medium and then transfer heat to the heat pump. 6.07.4 Recovering heat from tailings Tailings are the material residues after the mineral extraction process. They occupy a large volume and can sometimes be located at proximity to the mining operation. As tailings are in contact with ambient air, oxidation occurs and thermal energy is released which results in large amounts of heat released to the atmosphere. At the Doyon and Rum Jungle mines, temperatures 84
above 65°C and 55°C were measured in boreholes installed in waste dumps (Raymond, Therrien, & Gosselin, 2010). If temperatures are so high, the use of a heat pump would not be required, a closed-loop water circuit would be sufficient to heat fresh air and it could be even enough for surface buildings. The elimination of the heat pump reduces significantly operating and capital costs. It should be important to first assess the total amount of thermal energy that is possible to recover from the tailings. Heat exchangers should be positioned in a way to recover the maximum amount of heat. There would be some challenges regarding the type of HE material to use as tailings are usually very acidic.
6.08
Heating the surface buildings
It is first important to understand the difference between heating the intake mine fresh air and the surface buildings. The surface buildings will require heating for a much longer period of the year since they are heated to about 22°C as opposed to 1°C for intake fresh air. When performing an energy assessment, all of the available heat sources at the mine site should first be evaluated (nature, temperature, flow rate, position). The next question would then be; is it more feasible to discharge the heat to the surface buildings or to the ventilation fresh air? As the fresh air heating energy calculations are already known (Chapter 2), they can be compared with the surface building heating energy savings which can be evaluated from the following variables:
Building heat demand: it is dependent on the outside temperature; the colder it is, the greater the heat losses will be. Thus for each different temperature, the demand can be compiled throughout the whole year by predicting the ambient temperatures of the region. Actual heating cost at different heating demands: it is dependent on the efficiency of the present heating system and the heat source used; electricity, natural gas or propane. The portion of the heat recovered that can be discharged to the building for different heat demands. The efficiency of the heat pump.
From this information, the energy savings can be approximated and compared with the energy savings from the fresh air heating. Combining a heat recovery system from exhaust air with surface buildings heating could improve the efficiency of the system as it would be utilized for a longer period of the year instead of using it just for heating the mine air in winter. It is important to note that all scenarios should be evaluated. There could be some cases where one heat source should be discharged to the buildings and another one to the fresh air. The study will also determine if it is beneficial to combine the fresh air heating with the building heating. The distances between the heat sources, surface building and intake air shaft will greatly influence the best design to choose from. The following section describes a proposal to discharge the heat to the surface buildings and will help to better understand on how to evaluate the efficiency of the system.
85
In the case that the heat source has an elevated temperature, it can be discharged directly into the building with the use of a tube and fin HE. If the system is evaluated for an existing building, the present heating system should be adjusted to accommodate the additional heat source. For example, if the building is using a hot water heating system with natural gas burners, the water could be pre-heated upstream of the natural gas burners. Generally, the heat source will not have a grade high enough to be discharged directly into the building; a heat pump would be required. The heat would be transferred to the evaporator which would then transfer the heat into the building with the condenser. The condenser can transfer heat directly to the air or it can heat water which is then carried throughout the rooms of the building. The heat pump cycle requires work from the compressor which will usually reduce the overall efficiency of the system. For more information on the heat pump cycle, see Chapter 8. If the building heating system is combined with the fresh air heating, the system could be designed as follows: the same temperature sensor for the intake bypass valve, described in section 5.11, would be used to determine if the building heating system should be operating or not. The intake bypass valve system of the closed-loop glycol circuit is described in Chapter 3. If the outlet glycol temperature at the fresh air would achieve a temperature greater than its minimum (usually 1.5°C), it would indicate that the system does not require running at full load. For example, when the outlet glycol temperature of the fresh air HE reaches 2°C (instead of the required 1.5°C), the heat demand at intake is less than the total heat recovered. In other words, there is additional heat that could be used for other purposes than fresh air heating. Therefore to maximize the efficiency of the system, the remaining heat could be discharged to the buildings. The diagrams of different cases at which the combined system would operate are shown in Figure 6-8,Figure 6-9 and Figure 6-10.
Figure 6-8: Fresh air heat demand greater or equal than total heat recovered, no building heating
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Figure 6-9: Fresh air heat demand lower than total heat recovered, building heating from remaining heat recovered
Figure 6-10: No fresh air heat demand, portion of the total heat recovered for building heating, building heat demand fully satisfied
The proposed design is as follows: at the closest location to the main piping system and surface buildings, a bypass would be put in place with a shell and tube HE to transfer the heat to the evaporator as shown in Figure 6-11.
Figure 6-11: Bypass for building heating
As hot glycol flows across the shell and tube HE, it would cool down as it transfers heat to the refrigerant within the evaporator. The cold glycol would then mix back with hot glycol to heat the intake fresh air. The refrigerant vapour would flow to the condenser to transfer its heat to the building. As less heat would be required for the fresh air, more heat would be discharged to the building. The pneumatic actuated valve shown in Figure 6-11, would control the amount of heat discharged to the building. Note that when the pneumatic valve opens, there would be a greater pressure drop across the glycol circuit; the glycol flow rate would therefore decrease. A variable frequency drive motor for the pump would have to be used in order to maintain a constant flow rate of glycol. 87
As mentioned earlier, the building would require heating for a longer part of the year than the fresh air. Whenever fresh air does not require any heating and the building heating is functional as in Figure 6-10, undesired friction losses would be encountered as the glycol does not require to be carried all the way to the fresh air coils. One way to avoid this would be to fully bypass the intake HE as shown in Figure 6-12 with the use of valves. The use a VFD drive would be required to save pumping operating costs.
Figure 6-12: Bypass for building heating for no intake air heating
6.09
Conclusions
The alternative designs of heat recovery systems in underground mines are presented as follow: Spray chambers at exhaust, tube-fin HE at intake, plate heat exchangers to transfer the heat o Direct contact HE o Filtration system o Plate heat exchangers
Heat pump with the use of a refrigeration plant; direct-contact HE at exhaust, glycol tube and fin HE at intake Heat pump, evaporator at exhaust and condenser at intake Recovering heat from the depths of the mine Re-circulation of exhaust air Heat sources other than exhaust mine air o Heat recovery from mine air compressors o Mine water heat recovery o Geothermal ground heat pump o Recovering heat from tailings Space heating of the surface buildings
Alternative designs to the closed-loop glycol circuit are very important to take into consideration as they can be more suitable for a given operation. With possible new legislations to reduce greenhouse gas emissions, spray chambers at exhaust could be interesting. Innovative practices 88
for safer and effective air re-circulation could be a subject of further research in mine ventilation. Additional heat sources are also very important to consider as they can sometimes contribute to large energy savings at lower costs. The building heating should also be assessed as it could sometimes be more economical than heating the fresh air. It is also important to mention that there could be more additional heat sources and alternate designs than the ones listed previously; for example if a smelter plant is at proximity to the mine site there could be other possibilities of heating use or heat recovery.
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CHAPTER 7. TUBE AND FIN HE TECHNOLOGY; SOFTWARE FOR HEAT EXCHANGER DESIGN Summary This Chapter outlines the calculations of the tube and fin heat exchanger design performances. A design software tool is available. The performances of different designs of the HE are compared using the software tool.
7.01
Introduction
One of the main components of a heat recovery system is the tube and fin HE as it is used to recover and discharge the heat. The tube-fin HE performances are hard to predict and the existing theory is not valid in all cases. In order to understand better the design procedures of the tube-fin HE; a software has been developed. It is important to note that it is different and not included within the previous software described in Chapter 4 and 5. The software follows design procedures of tube and fin HE in (Shah & Sekulic, 2003). All the calculations and steps involved are explained. Some calculations are not described in detail as it was judged to be unnecessary for this thesis. In the case that more information is required; the user should consult (Shah & Sekulic, 2003). As explained in Chapter 5, the HE at exhaust and intake is separated into several smaller sets of HEs as they can be manufactured and shipped up to a maximum size. Therefore when using the software, the volume occupied by the HE should not be greater than the specifications of the HEs shown in APPENDIX A in order to obtain a realistic design. The volumetric flow rate of air and glycol would then have to be divided by the total number of HEs to input to the software. The heat transfer calculations will assume that solely dry cooling of air is occurring i.e. there is no condensation of water vapour on the coils disregarding the humidity content of air. Under humid conditions, very few correlations have been found in the literature and it was thus chosen not to include them in the software as they cover a very small range of tube and fin HE and would not be reliable in most cases. The heat transfer characteristics in humid conditions are described in Chapter 4.
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Nomenclature 2 Secondary surface area (fin) (m ) Af 2
A fr
HE frontal area (m )
Ad i
Cross-sectional inside tube area (m )
2
2
Ao
Minimum free flow area (m )
Ap
Primary surface area (tube) (m )
At
2
2
Total heat transfer surface area (m ) -1
Pu
Actual vapour pressure (kPa)
Pw Pr
Saturated vapour pressure at wet bulb temperature (kPa) Prandlt number, dimensionless
Q
Volumetric flow rate (m s )
rh
Hydraulic radius (m)
Re
Reynolds number, dimensionless
3 -1
-1
C
Heat capacity rate (kW K )
R
Thermal resistance (K W )
Cmin
td
Tube thickness (m)
T
Temperature (°C)
cp
Min heat capacity rate between the two fluids -1 (kW K ) Max heat capacity rate between the 2 fluids -1 (kW K ) -1 -1 Specific heat at const. pres (kJ kg K )
Tdb
Dry bulb temperature (°C)
DH
Hydraulic diameter (m)
Twb
Wet bulb temperature (°C)
di
Tube inside diameter (m)
Tlm
Log mean temperature difference
do
Tube outside diameter (m)
UA
Overall thermal conductance (W °C )
f
Friction factor, dimensionless
um
Air mean velocity (m s )
Fp
Fin pitch (m)
V
Volume occupied by HE (m )
G
W
gc
Mass flux (kg m s ) -2 Gravitational acceleration (9.8 m s )
Xl
Humidity ration (kgdry vapour/kgdry air) Longitudinal tube pitch (m)
h
Heat transfer coefficient (W m °C )
Xt
Transversal tube pitch (m)
j
Z
Kc
j Colburn factor, dimensionless -1 -1 Thermal conductivity (W m °C ) Entrance contraction-loss coefficient, ( )
Altitude (m) Greek symbol Difference
Ke
Exit loss coefficient, dimensionless
HE efficiency, dimensionless
L1
Width of HE (m)
p
L2
Length of HE (m)
Fin thickness (m)
L3
Height of HE (m)
f
Fin efficiency, dimensionless
Cmax
k
-2 -1
-2
-1
-1
-1
3
Single pass HE efficiency, dimensionless
Mass flow rate of water (kg s )
o
Overall fin efficiency, dimensionless
Nc
Number of circuits
Dynamic viscosity (Pa s)
Nf
Number of fins
Density (kg m )
NR
Number of tube rows
Nt / R
Number of tubes per row
Nt
Total number of tubes
Ratio of total surface area on one side of the heat exchanger to the total volume on both -1 side of the heat exchanger, (m ) Ratio of free-flow area to frontal area on one side of the heat exchanger, dimensionless Subscript
NTU
Number of Transfer Units Nusselt number, dimensionless Barometric pressure (kPa)
c i
Cold fluid Hot fluid Inlet
Saturated vapour pressure at dry bulb temperature (kPa)
m
Mean
m
Nu
Pb Pd
-1
h
-3
91
o
7.02
Outlet
Geometrical parameters calculations
Before performing the heat transfer calculations, the geometrical parameters of the HE must be known. The following will explain the variables required to be input by the user and the calculations to obtain all of the geometrical parameters. Due to the condensation and heavy fouling on the heat exchanger, the fins should be placed longitudinally so that the water and particulates evacuate more efficiently. The core area dimensions will thus be as shown in Figure 7-1.
Figure 7-1: Longitudinal fins heat exchanger
The following variables have first to be input within the heat exchanger design software
Tube arrangements: Inline
Staggered
Figure 7-2: Tube arrangements
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Longitudinal and transversal tube pitch: Xl, Xt:
Figure 7-3: Longitudinal and transversal tube pitch
Note: L3 is calculated from Xt and the number of tubes per row. The user can then verify if it corresponds to the desired height. Note: it is impossible to have a value of Xt or Xl smaller than the outer tube diameter.
L1: Width of the heat exchanger Number of fins per m: determined from fin spacing. Fin thickness: usually approximately 1 mm Tube wall thickness: Usually approximately 1 mm Tube inside diameter Number of tube rows Number of circuits: the number of divided flows as it enters the tube and fin HE.
The following geometrical outputs can then be calculated
Total number of tubes: N t N t / R N R
Number of fluid passes: N p
Eq. 7-1
Nt Eq. 7-2 Nc o Note: this number should return an integer; otherwise, some input variables should be changed to obtain a realistic value. Width: L2 N R X l Eq. 7-3 Note: Xl/2 is added at both ends.
Height: L3 o For inline arrangement: L3 X t N t / R o For staggered arrangement: L3 X t N t / R
Eq. 7-4 Xt 2
Eq. 7-5
Note: Takes into account the Xt/2 at both ends. 93
Note: As mentioned previously, the user should ensure that it corresponds to the available height of the heat exchanger casing.
Total number of fins: N f L1 N f / L
Eq. 7-6
Tube outside diameter: d o d i t d
Eq. 7-7
Frontal area: A fr L1 L3
Eq. 7-8
Primary heat transfer surface area:
Ap d o ( L1 N f L1 ) N t 2( L2 L 3
d o2 4
Nt )
Eq. 7-10
Secondary heat transfer area (fin surface): d 2 A f 2 L2 L3 o N t N f L1 2 L3N f L1 4
Eq. 7-11
The total heat transfer surface: At A f Ap
Minimum free flow area o For Inline arrangement: ) ( [(
)
Eq. 7-12
]
Eq. 7-13
o For staggered arrangement The following variables must first be calculated ( ) ( ) [( )
]
(
Eq. 7-14
)
Eq. 7-15
{ To then find the minimum free flow area [( ) ( ) (
Volume occupied by the heat exchanger: Hydraulic Diameter of the heat exchanger The following variables must first be calculated:
Eq. 7-16
)
]
Eq. 7-17 Eq. 7-18
Eq. 7-19 Eq. 7-20 To then find the hydraulic diameter
Eq. 7-21
94
7.03
Air conditions calculations
The fluid conditions are evaluated in order to calculate the air density, then the air mass flow rate and its mean velocity. The most important variables are the dry bulb temperature and volumetric flow rate. The rest could be omitted although it could help to obtain a more accurate value of the air density. The calculations to find the density and humidity ratio will not be described as they can be found in section 4.03. Inlet air variables at exhaust The following inlet air variables are first required to input in order to perform the calculations: Volume flow rate across frontal area Dry bulb temperature Wet bulb temperature Altitude Using the air inputs and some geometrical parameter of the heat exchanger it is possible to first determine the following:
Average Air speed across Heat Exchanger: u m
Mass flow rate of dry air: m Q
Q A fr
Eq. 7-22
7.04
Eq. 7-23
Inlet water variables
The running fluid is water, the following variables are required to implement within the design.
Inlet temperature Mass flow rate
These variables can be varied by the user in order to optimize the heat exchanger. In the heat recovery application, the exhaust air input variables would remain constant and it would therefore be important to vary the water mass flow rate and inlet water temperature in order to study the performance of the system. It is first important to assume an efficiency of the heat exchanger in order to have an approximation of the fluid outlet and mean temperatures. The efficiency must be entered by the user. For a single pass cross-flow heat exchanger, the efficiency should be between 50 and 75% (Shah & Sekulic, 2003). From the assumed efficiency the outlet temperatures can be calculated from the following equations:
C cp m Th,o Th,i
Eq. 7-24 Cmin (Th,i Tc,i ) Ch
Eq. 7-25
95
Tc,o Tc,i
Cmin (T Tc,i ) C c h ,i
Eq. 7-26
The cp’s of the two fluids were approximated to be at a temperature in between the inlet hot and kJ cold fluid temperatures. Thus the cp of air is evaluated to be 1.005 and that of water 4.18 kg K kJ . The specific heats of water and air are relatively constant at the operating temperatures kg K thus the assumed values do not have to change. In the case that the operating temperatures of the two fluids are not in between 0 and 25°C, the cp’s value might have to be changed within the software. From the outlet temperatures, the mean temperatures of the two fluids can thus be found from the following equations: For
Cmin 0.5 Cmax
Th,m
Th,i Th,o
Tc,o Tc,i
Tc,m For
Eq. 7-27
2
Eq. 7-28
2
Cmin 0.5 and Cmin is the hot fluid and Cmax cold fluid: Cmax
Th,m Tc,m Tlm
Tlm Tc,m
For
(Th,i Tc,m ) (Th,o Tc,m )
ln (Th,i Tc,m ) /(Th,o Tc,m )
Eq. 7-29
Tc,o Tc,i
Eq. 7-30 Eq. 7-31
2
Cmin 0.5 and Cmin is the cold fluid and Cmax is hot fluid: Cmax
Tc,m Th,m Tlm
Tlm Th,m
(Th,m Tc,o ) (Th,m Tc,i )
ln (Th,m Tc,o ) /(Th,m Tc,i )
Th,i Th,o 2
Eq. 7-32
Eq. 7-33 Eq. 7-34
96
7.05
Fluid properties and velocity calculations
After finding the mean temperatures the following fluid properties of the fluids are calculated; ρ, μ, k and Pr, using the pre-determined cp and the assumed efficiency. These properties are determined using data from water and air properties tables found in (ASHRAE, 2009), correlations have been developed to determine the fluid properties in function of the fluid temperature. These correlations developed are found in APPENDIX C, the temperature range at which they are valid is also included. The software does not take into effect the variable fluid properties as the temperature change is not of a large extent. It is thus assumed that it should not affect greatly the results. From the mean density and dynamic viscosity, the Reynolds number on the air side can be found. Depending on the correlation used, the Reynolds number may be required to be a function of the outer tube diameter (Equation 7-35) or the hydraulic diameter (Equation 7-36) u d Re d o m o Eq. 7-35
Re DH
u m DH
Eq. 7-36
The two correlations presented use the Reynolds number with respect to the tube outer diameter but the hydraulic diameter is still given in case it would be required for additional correlations added by the user. The working fluid within the pipes is assumed to be separated into the total number of circuits within the heat exchanger, the velocity within a single tube can be found using Equation 7-37. Q um Eq. 7-37 Ad i N c If the fluid velocity is too elevated, there could be intensive erosion of pipes which could lead to leaks which would then require the HE to be repaired. From (Thomas, 2008) the maximum fluid velocity for copper tubes should be of 1.37 m/s and 3.35 m/s for stainless steel tubes.
7.06
Heat transfer calculations
The main goal of the calculations performed within this software is to obtain a HE efficiency. In tube and fin HE, the most difficult variable to predict is the heat transfer coefficient on the air side. Although there have been several research project carried out on the tube and fin HE, there is still a lack of accuracy regarding the heat transfer coefficient results due to the large number of design possibilities. However the data calculated from the software is very useful in order to better understand the effect of the change in the HE design. Thus, even if the results found have errors, they should remain somehow proportional for any design. Understanding the behaviour of the HE with the change in data inputs was found to be very important to ensure the optimization of the design and not to take for granted the data received from the HE manufacturer. The 97
manufacturers have test cells running all the time to determine better the HE effectiveness (Thomas, 2008) and this is why their data was implemented within the heat recovery software. The HE design software should be used to finalize the heat recovery system and ensure that the HE has the best suitable trade-offs in terms of pressure drop and effectiveness for a given operation. 7.06.1 j factors correlations Using the Prandlt, Reynolds numbers and some geometrical parameters, the j factor can be calculated. The j factor calculation depends on the correlation used. The correlations will usually calculate a j factor or a Nusselt number. Two correlations will be used within the software, the Kayansayan and Wang correlations. The following will first describe the Kayansayan correlation. The j factor is defined by Equation 7-38. j 0.15 Re
0.28
A o A to
0.362
Eq. 7-38
Where 4 X Ao 1 l Ato d o
X t d o
d o d o N f 11 X t 2
Eq. 7-39
The Kayansayan correlation is valid for staggered arrangements. It uses the Finning factor which is defined by Equation 7-39 and takes into consideration most of the geometrical parameters of the tube and fin HE. The correlation is said to be valid within the range of Reynolds number with respect to the coil hydraulic diameter between 500 and 30,000. The Finning factor valid for the correlation should be between 11.2 and 23.5. The results from the experimental data was found to lie within a +/- 10% dispersion band around the mean line for 71.8% of the data. For more information on the correlation, see (Kayansayan, 1993). The Wang correlation uses less geometrical parameters; the correlation does not take into consideration the lateral and longitudinal distance between the tubes. The j factor is defined by Equation 7-40. The correlation is valid for a Reynolds number based on tube collar diameter between 800 and 7500. From the experimental data, it was found that Equation 7-40 can describe 97% of the results within a 10% error. For more information on the correlation, see (Wang & Chang, 1996).
j
0.394 Red o0.392
do
0.0449
N
0.0897
Fp d o
0.212
Eq. 7-40
Then using the following relation, the Nusselt number is found from Equation 7-37. j
Nu Pr 1/ 3 Re
Eq. 7-41
Using the Nusselt number, the heat transfer coefficient is found from Equation 7-38. Nu k h h Eq. 7-42 DH 98
From the heat transfer coefficient and the following equations, it is possible to calculate the fin efficiency. X X re l t
ro
do 2
2h m a k f e
0.5
f
1/2
* n , m e (r ) , n exp(0.13m
2
, r * re / ro ,
f
b 0.9706 0.17125 e r *n tanh
1.3863),
re ro
b 0.9107 0.0893 r * f
e
for r * 2 for r * 2
for 0.6 2.257 (r * ) 0.445
f a(m e ) -b for 0.6 2.257 (r * ) 0.445
Eq. 7-43 Eq. 7-44
From the fin efficiency it is then possible to calculate the overall efficiency from Equation 7-45: Af
(1 f ) At
o 1
Eq. 7-45
For more information on the fin efficiency calculations see (Shah & Sekulic, 2003) On the water-side, the Nusselt number can be calculated using the Dittus-Boelter correlation for turbulent flow in a smooth pipe. Nu=0.023Re0.8Pr0.4 Eq. 7-46 The heat transfer coefficient can be found using the following equation:
h
Nu k c do
Eq. 7-47
Then the resistance of both fluids and the wall are found from Equations 7-48, 7-49 and 7-50. The water side:
1 (hA) c
Eq. 7-48
The air side 1 Rh ( o hA) h The wall resistance
Eq. 7-49
Rc
99
Rw
ln(ro / ri ) 2kt L
Eq. 7-50
Note that in most cases, the wall resistance is not significant enough to affect the results. Thermal resistance should be included in the calculations if fins are wrapped in tension or mechanically expanded onto the tubes. In the case that they are attached by a mechanical fit, resistance can be neglected. This resistance is not included within the software calculations. There can also be some resistance due to fouling which can be added to the system. Note that in our case a water film is present due to vapour condensation and should therefore be taken into consideration. The software does not take it into account. The mean wall temperature can then be determined from Equation 7-51
Tm, w
Tc,m Th,m 1 1 R R R R h h c c
1
Eq. 7-51
From the resistance of the system, it is possible to determine the overall thermal conductance UA of the system.
1 Rh Rw Rc UA
Eq. 7-52
From the thermal conductance, the number of transfer units (NTU) can be found from the following equation:
NTU
UA Cmin
Eq. 7-53
Using the NTU and the C*, the efficiency can be determined assuming that the exchanger is a cross flow arrangement with the air-side unmixed and the tube side mixed. The conditions to determine if the fluid is mixed or unmixed are as follows:
The air-side of the fluid is always unmixed unless there are individually finned tubes The tube-side of the fluid is unmixed for 4 rows or more. The tube-side of the fluid is partially mixed for 2 or 3 rows. The tube side of the fluid is mixed for one row.
It has been reported by (Di Giovanni & Webb, 1989) that the mixed arrangement has the most conservative approach, thus it shall be used for partially mixed case. No partially mixed efficiency formula has been found in the literature. For Cmin mixed and Cmax unmixed
p 1 exp 1 exp( NTU C*)/ C *
Eq. 7-54 100
For Cmin unmixed and Cmax mixed 1 p (1 exp C * 1 exp( NTU ) C* For both fluids unmixed
Eq. 7-55
n
C * P ( NTU )
p 1 exp( NTU ) exp (1 C*) NTU 1 Pn ( y ) (n 1)!
n
j 1
n
(n 1 j ) n 1 y j! j 1
Eq. 7-56
Note that the formula for both fluids unmixed is not included in the calculations since the efficiencies found were much greater than the results given by the manufacturer. Solely Equation 7-55 will be used since the software does not take into account individually finned tubes. The efficiency found from Equation 7-55 is for a single pass arrangement on the tube side. In the case of a multi-pass HE, Equation 7-57 or 7-58 (Joadar & Jacobi, 2008) will determine the efficiency for overall counter or parallel flow. It is always more efficient to have overall counter flow passes. n
1 pC 1 1 p n (Overall counter flow) 1 pC C 1 p
1 (1 ( p (1 C * ))) n 1 C*
(Overall parallel flow)
Eq. 7-57
Eq. 7-58
From the efficiency found, iterations with the assumed efficiency must be performed until the two values are similar. The longitudinal conduction effect (λ) is not included in the calculations as it is generally small and negligible for cross flow heat exchangers (Shah & Sekulic, 2003).
7.07
HE efficiency design study
In APPENDIX A the data sheet of IHT Inc. demonstrates that the air face velocity is of 10.52 ft/s (3.2 m/s). For the Laronde mine installations, this would mean increasing the actual face area of the building 3 times its size which would result in significant capital cost. It thus important to determine the maximum face velocity achievable to optimize the trade-offs between efficiency, pressure drop and capital cost. The IHT Inc. design variables were first input in the software to determine the difference between the results given by IHT Inc. and the ones obtained with the software. The results obtained using the Wang correlation to determine the efficiency were found 101
to be much greater than the results given by IHT Inc, an error of 44% was calculated between the two values. The Kayansayan correlation results were also found to be greater, an error of 23% was found thus much less than with the Wang correlation and it was therefore chosen to perform the analysis. Note that the efficiencies found from the two correlations are both greater than the value given by IHT Inc. and it is therefore important to remember that the values are most probably not conservative as compared to the reality. As mentioned in Chapter 2, it is assumed that the manufacturers have the better results as they have a much greater experience and have experiment test cells running continuously. 7.07.1 Air face velocity effect on the efficiency A study will be performed on the effect the air face velocity on the efficiency of the HE. The air face velocity will be varied from about 3.2 m/s to 9.6 m/s. Several HE designs will be studied (6 in total) and the effectiveness for a single pass glycol flow will be taken into consideration. The air face velocity will be varied by changing the width (L1). The geometrical parameters of the designs can be found in APPENDIX C. Design no. 1 is as the IHT Inc. design. Figure 7-4 shows the change in efficiency with respect to the air face velocity for each of the design number one to five.
Figure 7-4: HE efficiency vs Face air speed
By looking at Figure 7-4 it is possible to note that the relationship can almost be considered to be linear. Also, it is possible to see that the greater the efficiency of a given geometry, the steeper the slope will be. Therefore, for low single pass HE efficiencies design, the air speed should not affect greatly its heat transfer performance. The discontinuity of the fin efficiency in design no. 2 is due to the change in calculation from Equation 7-43 to Equation 7-44 which affects the HTC. Obviously this discrepancy does not reflect the reality and shows that the calculations involved to obtain the efficiency cannot always be taken for granted. 102
7.07.2 Air face velocity effect on the pressure drop IHT Inc. designed the HE in a way that the face velocity of the fluid would be 10.52 ft/s (3.21 m/s). They did not mention what would have been the effect of increasing this velocity. The effect of increasing the face velocity will be studied in the following. The main issue with increasing the air face velocity is the increase in pressure drop of the system. The pressure drop of the HE is dependent on the square of the velocity; the equation is as follows; i G L 1 G'2 2 2 p f i 2 i 1 1 ' Kc 1 ' Ke 2 g c i rh m o 2 g c i o densitychangelosses entrance and exit losses friction losses
2
Eq. 7-58
Where
G' um leadingedge
Eq. 7-59
To simplify calculations, it will be assumed that the losses due to density change, entrance and exit are not significant. The entrance and exit losses should be included when finalizing the HE design as they can be significant in some cases. It is also fair to assume that the effect of the change in density is negligible for most cases as the temperature difference is relatively low. Equation 7-58 can thus be approximated to Equation 7-60.
G2 p 2 g c i
L 1 f i rh m
Eq. 7-60
The air face velocity is not directly proportional to the square of the face velocity (G) since the friction factor decreases with increasing Reynolds number. For example, for an air face velocity 2 times greater, the pressure drop would be approximately 3.4 times greater instead of 4. The friction factor is determined from correlations from experimental data. Presently, the program uses a correlation from (Wang & Chang, 1996), the correlation is as follows:
f
1.039 Red o0.418
t do
0.104
N
0.0935
Fp d o
0.197
Eq. 7-61
The correlation is said to be valid for a Reynolds number between 2000 and 7500 with respect to tube outside diameter. As in the HE efficiency, the pressure drop vs. the air speed was plotted for all of the 6 designs. It is shown in Figure 7-5.
103
Face air speed vs pressure drop 500 450 Pressure drop (Pa)
400 350
n5
300
n4
250
n2
200 150
n3
100
n1
50
n6
0 0
2
4
6
8
10
12
14
Air speed (m/s) Figure 7-5: Pressure drop vs Face air speed
The pressure drop is proportional to a little less than the square of the velocity. It is easier to observe with design no. 1; at 4 m/s the pressure drop is approximated to be 110 kPa, at 8 m/s it is 340 Pa, thus 3.3 times greater instead of 4 (square of two times the velocity). The difference is due to the change in friction factor which will be 3.3/4=0.825. The value of 0.825 would therefore be the ratio of the friction factor using Reynolds number at 4 and 8 m/s face speed. Thus if a pressure drop is known for a given geometry and face velocity, it is possible to approximate the new pressure drop for a different velocity using Equation 7-62.
f P2 2 f1
2
um2 P1 u m1
Eq. 7-62
Note that this equation was derived from the previous analysis and has not been found in the existing literature. It could not be valid in the case that the density changes effect is great enough to affect the values or if the entrance and exit losses in the two cases are different.
7.08
Conclusions
The calculations of tube and fin HE performances require a relatively large number of procedures. A software was developped to better understand the influence of the air speed and geometry on the pressure drop and HE efficiency. As mentioned earlier this software is different from the one described in Chapters 4 and 5. It should be used in the detailed feasibility study stage of the design. As the research on tube and fin HEs will further evolve, heat transfer coefficients and friction factor correlations should become more accurate in the future and the software could be used to accurately determine the pressure drop and efficiency for a large range of geometries and air conditions.
104
CHAPTER 8. MEANS OF REDUCING THE ADVERSE EFFECTS OF ADIABATIC COMPRESSION (EXCLUDING NATURAL AND MECHANICAL COOLING) Summary This Chapter explains the effect and theory on adiabatic compression in underground mines. Some means of reducing its heating effect other than by mechanical or natural cooling are described.
8.01
Introduction
The adiabatic compression can be described in several ways; it is often referred to as the conversion of potential energy into thermal energy. As air descents to the underground levels, its enthalpy increases due to the compression of the air columns. This phenomenon can be analogically compared to air going through a compressor; its pressure is increased and therefore its enthalpy increases. The work of the compressor can be described as follow:
Figure 8-1: Compressor adding positive work to air
W in Q( P2 P1 )
Eq. 8-1
Where Q is the volumetric flow rate of air The change in specific enthalpy will be:
h2 h1
W in
Eq. 8-2
m
Where m is the mass flow rate of air For an underground mine, the rate of work would become:
W in m gZ
Eq. 8-3
Where Z is the change in elevation and g the constant gravitational acceleration (9.81 m/s2) 105
m Q
Eq. 8-4
The increase in pressure for an underground mine is calculated as follows: gZ P2 P1 Eq. 8-5 This equation is valid assuming that the density change between the two different altitudes is not significant. Equation 8-6 is a correlation that calculates the barometric pressure for different altitudes (below or above sea level). As opposed to Equation 8-5, this correlation takes into account the change in air density.
Pb
Z 101.3 10 19075
Eq. 8-6
Assuming a mine with a shaft collar at sea level, the pressure increase with depth has been calculated with Equation 8-5 and 8-6. The results have been plotted in Figure 8-2, Equation 8-5 is shown in red and Equation 8-6 in blue. Change in pressure with negative Altitude 90 80 70
Exponential
ΔP (kPa)
60
Linear
50 40 30 20 10 0 0
1000
2000
3000
4000
5000
6000
Depth (m) Figure 8-2: Change in pressure with negative altitude
By observing Figure 8-2, the difference between the two equations becomes significant for depths of one thousand meter or more. As Equation 8-6 is the real pressure change, it should always be used for deeper levels. The change in specific enthalpy at different depths is: h2 h1
gZ 1000
Eq. 8-7
106
Using Equation 8-7 and the constant Cp of air at 1.005 kJ kg-1 K-1, the change in dry bulb temperature is 0.974°C/100 m. As the change in Cp with temperature is relatively constant for temperature range between 0 and 40°C, the relation should be accurate enough. The change in temperature due to adiabatic compression can also be calculated using the polytropic equation (Moreby, 2007). The polytropic equation shown in Equation 8-8 is used in adiabatic process and is valid for ideal gases. n
P1 T1 n 1 P2 T2
Eq. 8-8
Using Equation 8-8 and 8-6 and the following assumptions: Shaft collar at sea level (P1=101.3 kPa) T1=20°C n=1.4 (for air) The temperature change with depth is calculated from; T2-T1 The temperature increase with depth has been plotted in Figure 8-3 for the 0.974°C/100 m relation (red) and the polytropic equation (blue). Temperature increase with depth for intake temperature of 20°C
Dry bulb temperature change (°C)
60 50 40 Polytropic
30
Linear enthalpy 20 10 0 0
1000
2000
3000
4000
5000
6000
Depth (m) Figure 8-3: Temperature change with depth, polytropic equations and linear relationship
The linear relationship should be more realistic than the polytropic equation as it respects the first law of thermodynamics i.e. the change in enthalpy is equal to the work performed on the 107
fluid. From this analysis and Figure 8-3 it is possible to mention that the polytropic equations should not be used for deep levels. It should be noted that both relationships assume that the Cp is constant. In reality, the increase in dry bulb temperature will usually be much less than 0.974°C/100 m. As the air flows down the ventilation shaft, it gains humidity (Whillier, 1990), the increase in humidity will decrease the dry bulb temperature. However, the change in total enthalpy will still be as in Equation 8-7. Using air temperature and humidity, the change in enthalpy from adiabatic compression is calculated from Equation 8-9.
h2 h1 m C p (T2 T1 ) hlg (W2 W1 )
Eq. 8-9
Where hlg : Heat of evaporation W
: Humidity ratio
As the humidity content of air increases ( W2 W1 ), the change in dry bulb temperature ( T2 T1 ) decreases since the change in enthalpy ( h2 h1 ) remains constant with regards to the adiabatic compression effects.
8.02
Use of turbines instead of regulators
As air flows along the ventilation shaft, friction losses induce a pressure drop. Due to the latter, the actual absolute pressure change is slightly less than what is found with Equation 8-6. Intuitively, the pressure drop from friction losses would reduce the effect of adiabatic compression. Unfortunately friction losses are dissipated into heat and therefore increasing friction does not reduce the heat load within the mine. If it would do so, one recommendation would be to install a regulator at the intake instead of the exhaust of the level. In order to reduce the effect of adiabatic compression a turbine can be used to control the airflow. As regulators induce losses to control the airflow inside the mine, some of these losses could be recovered and transferred into work instead of dissipating them into heat. The turbine would be connected to a generator which would induce a different load depending on the required pressure drop to control the mine airflow. To reduce some of the effects of adiabatic compression, the turbine could be installed at the intake of the level. The schematic of the system is shown in Figure 8-4.
108
Figure 8-4: Turbine coupled with generator with variable load for flow regulation
This suggestion was first proposed by (Barenburg, 1976) but to use the system solely to reduce the effect of adiabatic compression. It was considered to be not practical as the relatively low efficiencies of turbines and generators would leave a very small portion of the actual power recovered. However, using them instead of regulators could make the system feasible as the concept of regulators already involves a waste in energy. Assuming an efficiency of generator at 90% and turbine at 70%, the resulting efficiency of the system would be of 63%. In this case, the system would reduce the heat load by 63% that is assuming that the friction losses within a regulator are fully dissipated into heat. A calculation example is performed to evaluate the benefits of the system; For a regulator: Resistance across the regulator: 1 N s2 m-8 Flow rate of air: 50 m3/s P RQ 2
Eq. 8-10
From Equation 8-10 the pressure drop across the regulator is 2.5 kPa. The total power loss across the regulator is 125 kW (ΔP x Q), assuming that the conditions are constant throughout the whole year, the energy loss is 1095 MWh/year, which for an electricity cost of 50$/MWh would result in an operating cost of 54,750$/year. Note that these losses would not be present if booster fans would instead control the mine ventilation air. Assuming air density at 1.3 kg/m3, mass flow rate of air at 65 kg/s, and Cp=1.005 kJ kg-1 K-1, the increase in dry bulb temperature would be (assuming no change in humidity ratio): T
125
1.9C
Eq. 8-11
mCp
109
On the other hand, the negative work would decompress the air and therefore reduce its temperature to the same amount; the actual change in temperature within the regulator would then be equal to zero. This respects the first law of thermodynamics as the process can be considered to be adiabatic and no external work is involved. Thus there is no change in air enthalpy or temperature assuming that there would not be any change in kinetic or potential energy. For a turbine: From previous assumptions of generator and turbine efficiencies, 63% of the negative work induced on the air is converted into electricity. Total air power loss: 125 kW Power loss from friction: 46 kW Power converted into electricity: 79 kW From Equation 8-9, the increase in dry bulb temperature due friction losses is 0.7°C. The temperature decrease due to decompression of air is of 1.9°C therefore the air temperature across the turbine would be decreased of 1.2°C. Nonetheless, 79 kW of electricity could be used to power a fan.
8.03
Brattice wall
The system consists of having the upcast and downcast air in parallel within the same shaft separated by a brattice wall with low thermal resistance material such as steel (Barenburg, 1976). The heat transfer would occur in the desired direction solely if the dry bulb temperature of the downcast is higher than the upcast air dry-bulb temperature; downcast air would therefore get cooler. However, brattice walls are not used in underground metal mines but only in coal mines. As fresh air temperature is relatively low, the dry bulb temperature will be greater in the intake than in the exhaust shaft. Other disadvantages of this design are as follow: Steel wall would be costly. No fans could be placed at the bottom of the mine increasing the temperature thus reducing the heat transfer.
8.04
Conclusion
The theory of adiabatic compression has been presented to have a better feel of the thermodynamics behind this effect. Most of the ideas proposed in this Chapter are considered to be relatively ambitious as they would require the installation of complex systems and would most likely encounter operational problems. However, the most efficient way to alleviate the adverse effect of adiabatic compression is the use of cooling plants which will be discussed in the next chapter.
110
CHAPTER 9. NOVEL COOLING SYSTEMS; APPLICATIONS TO CANADIAN MINES Summary This Chapter describes existing and novel cooling system designs with some recommendations for implementation in Northern climates. Natural cooling and heating systems using the advantage of the large temperature difference between winter and summer in Northern regions are outlined. The vapour compression cycle is as well explained.
9.01
Introduction
Due to the adiabatic compression of air, elevated virgin rock temperature at depths and heat created from underground diesel equipment, mechanical cooling of the intake air is presently used in two deep Canadian mines; Laronde, in Abitibi, Qc and Kidd Creek in Timmins Ont. However, the number of mines equipped with an air cooling system will increase in the future as mines become deeper. As opposed to Australia and South Africa, air cooling is fairly recent in Canadian mines. Recommendations and technologies for mine cooling systems in the warmer countries cannot be directly applied to the colder climate systems as the ambient air conditions are much different. The best approach to design the cooling systems in cold regions is still not straight forward. Recommendations and novel cooling plant designs are presented in this Chapter. The large temperature difference between winter and summer can be used as an advantage to reduce energy costs. Several innovative designs that exploit cold winter weather as an asset to cool mine intake air are explained. Also, the vapour compression cycle is described.
9.02
Vapour compression cycle
Mechanical refrigeration is mostly achieved using the vapour-compression cycle. The personnel of an underground mine where a refrigeration plant or heat pump is to be installed should clearly understand the vapour-compression cycle to ensure that the manufacturer provides the proper equipment for the required cooling or heating demand at the best efficiency. The basics of the vapour compression cycle are described in this section. Also, some recommendations and new technologies to increase the efficiency of the cycle are outlined. Basic calculations of the ideal cycle are included. The vapour compression cycle is behind most heat pump or refrigeration plant. The only difference between the two is the purpose; a heat pump is used for heating and a refrigeration plant for cooling but the concept remains the same for both. The following explains the vapour compression cycle using a common household refrigerator as an example.
111
The second law of thermodynamics states that the heat will always flow from a hot to a cold object and never the opposite. Therefore cooling food within a refrigerator would seem to be impossible where there is no connection to a cold source to remove heat. How can heat be extracted from a fridge as heat cannot flow from a cold source to a warmer source? One possibility is to utilize external work and a phase change fluid. The phase change fluid is called the refrigerant. By varying its pressure, the refrigerant has the ability to evaporate at relatively low temperatures and condensate at warm temperatures. The heat exchanger inside the fridge is called the evaporator. The heat from your food is transferred to the liquid refrigerant and evaporation occurs. Then as the refrigerant is evaporated, it goes through a compressor which increases its pressure and temperature. The vapour will then attain a higher condensation temperature. The heat exchanger that rejects heat to the surrounding air is called the condenser. It is the hot area usually located in the back of the refrigerator. Afterwards, the liquid refrigerant is set to a lower pressure with the use of an expansion valve to decrease its boiling temperature. Then the vapour returns to the evaporator. When varying the desired temperature within the fridge, the expansion valve and compressor settings are changed. 9.02.1 Ideal cycle To first understand better the calculations, the ideal cycle will be described. The ideal vapourcompression cycle is defined as follows: No pressure drop in heat exchangers or connecting pipes Saturated vapour is leaving the evaporator Saturated liquid is leaving the condenser The expansion process is isenthalpic (i.e. no change in enthalpy) The compression process is isentropic (i.e. no change in entropy). The ideal vapour compression process is shown in Figure 9-1 with the different processes on the ln (P) vs. specific enthalpy graph.
Figure 9-1 (a) Schematic of the vapour compression-cycle (b) Schematic of a ln (P) vs h diagram with the state points of a vapour-compression cycle (Radermacher & Hwang, 2005)
From Figure 9-1, points 1, 2 and 3 are in the saturated vapour zone and point 4 in the saturated liquid zone. If the temperature in your refrigerator is increased, line 1-5 will be translated higher due to a change in the expansion valve (line 4-5) and compressor (line 1-2) settings. In the ideal 112
cycle, the totality of the work from the compressor is rejected into heat at the condenser (line 23-4). The change in enthalpy from point 1 to 2 is approximately equal to the change in enthalpy from point 2 to 3. The heat transferred to the evaporator is fully discharged from the condenser to the ambient air therefore the change in enthalpy from point 3 to 4 is equal to the change in enthalpy from point 1 to 5. In order to design a refrigeration plant, the operating temperatures must first be determined. In the evaporator (inside the refrigerator), the air temperature always remains a little higher than the refrigerant temperature to maintain heat transfer. Therefore the temperature of evaporation (line 1-5) must be set to a lower level than the desired ambient air temperature. On the other hand, the condenser (line 2-3-4) must be set to a higher temperature than the maximum possible ambient temperature so that the heat exchange will always occur from the condenser to the ambient air and not the opposite. From the previous explanation, it should be understood that the heat from the food inside a refrigerator actually warms the surrounding ambient air. Determining the actual operating temperatures requires a more detailed analysis of the evaporator and condenser HEs which can be done using the NTU analysis as in Chapter 7. It is important to note that in a phase change refrigerant heat exchange, (McQuiston, Parker, & Spitler, 2005) Cmin/Cmax can be assumed to be 0. In this case, NTU ln(1 ) (Shah & Sekulic, 2003). This analysis will depend on the type of HE used but the procedure is similar to what was presented in Chapter 7. When the required refrigerant operating temperatures are known, point 1, 3, 4 and 5 can be located on an actual ln (P) vs. h graph of a given refrigerant. Figure 9-2 shows the determined points on the Ammonia (NH3) ln (P) vs h diagram. From the required cooling power of the plant, the refrigerant mass flow rate can be determined from Equation 9-1 as it is constant throughout the whole cycle.
m NH3
Pcool (h1 h5 )
Eq. 9-1
Line 1-2 is the compressor work, it is assumed to be isentropic so the line is drawn parallel to the S-lines of the diagram. Point 2 will be the intersection of this line with line 2-3-4. The required work from the compressor is calculated from Equation 9-2.
W m NH3 (h2 h1 )
Eq. 9-2
The amount of power rejected from the condenser is calculated with Equation 9-3.
q cond q evap W
Eq. 9-3
Finally point 4 to 5 can be determined to size the expansion valve.
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Figure 9-2: Ammonia vapour ln(P)-h diagram (ASHRAE, 2009)
9.02.2 Actual vapour compression cycle Under realistic operations, there are several effects omitted in the ideal cycle. They affect considerably the performance of the cycle. The ln (P) vs. enthalpy diagram of the actual vapour compression cycle is shown in Figure 9-3.
Figure 9-3: (a) Schematic of vapour compression cycle with state points to explain realistic operating conditions (b) Schematic of realistic vapour compression cycle on a ln (P)-h diagram. Superheat (1-1a), subcooling (4a-4) and pressure drop (slant in line 2-4 and 5-1a) are included (Radermacher & Hwang, 2005)
It is first possible to notice that lines 2-4 and 5-1 are not isobars; as the fluid flows in the evaporator and condenser, friction losses decrease its pressure. The greater the friction losses, the greater the slope of the lines will be. Friction losses also occur within the connecting pipes 114
between the evaporator and condenser. The compressor must overcome these losses which reduces the performance of the system. The manufacturer must design the system in a way that the friction losses are minimized while achieving the required heat transfer. At the evaporator, the refrigerant is superheated (line 1 to 1a). It is performed so that no liquid flows across the compressor as it could damage it. Also it ensures that the totality of the refrigerant is evaporated where it contributes to the cooling capacity. Line 2-2s represents the amount of irreversible internal friction within the compressor i.e. the increase in entropy within the system. Again, the compressor must overcome these additional losses. From 4 to 4a, the saturated liquid is subcooled to ensure that only liquid refrigerant enters the evaporator to maximize the cooling by increasing the liquid to vapour ratio (Radermacher & Hwang, 2005). In the ideal cycle, the properties of the refrigerant will not influence the performance of the system. However, in the realistic cycle, the efficiency and performance of the cycle will be significantly affected by the properties of the refrigerant. Furthermore, the properties of the refrigerant will have an effect on both the capital and operating costs. One way to increase the efficiency and performance is the use of refrigerant mixtures. The following describes the difference between the use of zeotropic mixtures and a pure refrigerant.
Figure 9-4: Temperature-enthalpy diagram of cycle with heat exchange (Radermacher & Hwang, 2005) (a) Pure R22 (b) R22/R114 (50/50 wt. %)
Figure 9-4 shows the vapour compression cycle on the T-h diagram where A-B is the evaporation process, and C-D the condensation process. Ac-Bc and Cc-Dc is the temperature profile of the heat sink and heat source within the evaporator and condenser respectively. Figure 9-4 (a) shows the cycle with the use of a pure R-22 refrigerant, the temperature of the refrigerant does not vary during the evaporation and condensation process. Figure 9-4 (b) demonstrates the cycle with a zeotropic refrigerant mixture (R22/R114, 50/50 wt. %). As the refrigerant condensates, its temperature decreases. This approach is called “matching” the temperature glide and reduces the amount of irreversible losses within the system (Radermacher & Hwang, 2005). There are several other advantages of matching the temperature glide but the main goal here is to instigate awareness that several new technologies have been developed to increase performances of refrigeration plants. As the mine refrigeration plants have high capacities, technologies such as the use of refrigerant mixtures and others must be looked into carefully as energy savings may be significant. The performance of the refrigeration cycle is described from the coefficient of 115
performance COP. It is defined as the total cooling capacity of the evaporator divided by the compressor work as shown in Equation 9-4. COP
h1 h5 h2 h1
Eq. 9-4
Where point 1, 2 and 5 are as in Figure 9-1
9.03 New design proposal: producing work from the refrigeration plant With new technologies presently being developed in the field of low grade heat engines, it could be possible to produce work from the rejected heat in the cooling cycle. In other words, can the condenser use heat to generate work instead of rejecting it to the exhaust air? In the case that such technology would be available; the cycle would become more efficient but most importantly it would reduce significantly the heat rejection problems in underground mines. It would most likely require that the heat source at the evaporator (fresh air) has a relatively elevated temperature. To determine if the application is possible, the following question should first be asked: using the latest available technology on low grade heat engines, what is the lowest air temperature at which useful work can be generated? Note that even if this temperature is greater than the required cooling level, it could still be feasible to implement the system. For example, if an airflow rate at 40°C must be cooled to 28°C and useful work can be generated for temperature of 35°C and over, power would still be generated from 40 to 35°C.
9.04 Questioning the use of surface air cooling for Canadian mines In warmer countries such as Australia, cooling is required for the deeper as well as the lower levels as ambient air is already very hot. During hot periods of the year, ambient air is itself the major heat source. Surface air cooling is usually performed to cool the lower levels and the air can sometimes be cooled once again in an underground cooling plant for the deeper levels. Having a refrigeration plant on surface is cheaper and more convenient than an underground cooling plant as heat rejection is not limited to the ventilation air flow rate and conditions. In underground cooling plants, as the exhaust air becomes very hot and humid, the capacity of rejecting the heat becomes limited. In surface plants, heat rejection can be performed with large capacities cooling towers. Heat rejection capability for underground plants in Canadian mines is relatively reasonable as air conditions are usually not as extreme as in warm regions of the world. In Canadian climate, as ambient air temperature is relatively low, the lower levels of the mine will usually not require cooling. Moreover, as ventilation air gains heat from strata, the surface cooling system has to offset this heat gain so that the cooling effect reaches the lower levels as desired. This geothermal heat can be very significant and result in elevated energy costs. It was mentioned that the Laronde mine surface cooling plant efficiency was decreased significantly after the transfer of the original underground cooling plant to the surface due to the heat transferred from the rock walls (Lafontaine, 2010). It was also mentioned in (Tuck & Paudel, 116
2010) that at the Stawell Gold mine in Victoria Australia (considered to have a mild surface climate), half of the cooling provided by the refrigeration surface plant is dissipated before it reaches the active headings. Due to the latter, surface cooling for deep underground mines in cold regions will not be the most efficient option for most operations. Using underground refrigeration plants or spot coolers are options that should be better suited for such environments. In all cases, prior to decide on the best option for cooling, a careful assessment on the geothermal heat gains of the cooled air should be performed.
9.05 New design proposal: Closed-loop glycol circuit for refrigeration plants on surface at sub-zero glycol temperatures The following design proposal can be applied to any underground mines across the world. A refrigeration system would be located on surface cooling ethylene glycol to sub-zero temperatures in a closed-loop glycol circuit. The lower the temperature is the less flow rate is required to be carried around the loop which reduces significantly capital and operating costs of the piping and pumping system. As sub-zero glycol would reach the deeper levels, a tube and fin HE would transfer the heat from air to glycol. As air is usually elevated in humidity, condensate would accumulate and freeze on the HE. The latter would remove a large amount of water vapour decreasing the wet-bulb temperature as desired. After a certain period of time the HE will be covered in ice. Then as it reaches a certain size, the system would stop running and the ice would cool the air until it is fully melted. As the ice would be fully melted, the system would run again until the ice block reaches its maximum size. Using a closed-loop circuit instead of bulk air coolers is advantageous since the fluid will not have to overcome gravity when pumped back to surface. However, highpressure resistant piping system has to be used which results in elevated capital cost. The main issue with this design is that the pipes would have to be insulated almost perfectly; otherwise there could be severe ice accumulation problems at undesired locations. Furthermore, the insulation cost would be higher as the material would have to be less conductive than usual chilled water pipes as the temperature difference between the two fluids would be greater. Finally, cooling at such low temperatures would require that the compressor within the refrigeration plant provides a larger amount of work than with chilled water pipes which will consequently decrease the efficiency of the plant. This proposal could be interesting to look at more into depth in future research.
9.06
Natural heating and cooling system (NHEA)
In some underground mines, where the caving methods are used, it is possible to find large amounts of broken rocks connecting the surface with underground levels. If air is carried through the mass of broken rocks, it is possible to use them as a large heat capacitor. During hot periods, warm air transfers its heat to the rocks, cooling the air. On the other hand, during colder months, cold air flowing through the rocks will capture heat from the rocks. It is therefore a natural heating and cooling system using the heat capacity of rocks also called the natural heat exchanger area (NHEA). It is efficient and can restrain the need of installing a costly refrigeration plant. Any underground mine located in cold regions having an important mass of 117
broken rocks connecting the sub-surface levels to the surface should definitely look into the possibility of using such a system. The schematic of this system is shown in Figure 9-5.
Figure 9-5: Diagram of natural heating system (Sylvestre, 1999)
The Creighton and Kidd Creek mine both located in the province of Ontario, Canada are presently using such a system. At the Creighton mine, the intake fresh air downstream of the natural heat exchanger is said to be kept at approximately 3°C throughout the whole year (Sylvestre, 1999).
9.07
Ice stopes
Similar to the natural heating and cooling system previously described, the so-called ice stopes also use the advantage of large temperature differences between warm and cold periods of the year to transfer and store heat. During cold periods, air is carried to an underground stope where water is sprayed through the cold air. As the water comes in contact with the cold air it transfers its latent heat to the air. Ice forms within the stope warming the intake air to approximately 0°C (Sylvestre, 1999). During warmer periods, as fresh air flows through the stope, warm air transfers its heat to melt the ice, cooling the ambient air. The melted ice is then pumped to surface in order to avoid flooding. Stobie mine located in Sudbury, Ontario has implemented such a system and is still presently using it (Cornthwaite, 2009). The diagram of the ice stope system is shown in Figure 9-6.
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Figure 9-6: Diagram of ice stopes (McPherson, 1993)
A similar concept was proposed to heat air at surface by creating ice blocks during winter to then store them and melt the ice blocks during warmer periods to cool the air (Howes & Hortin, 2005). However, no practical applications have been reported in the literature. A similar design is proposed in the section 9.09.
9.08
Ice conveying to the underground levels
Carrying ice to the underground level has been realised by several mines in South Africa. Ice machines are located on surface and ice is sent through large diameter pipes to the underground levels. The advantage of sending ice to the undergrounds is that the amount of water that is required to be pumped back to the surface is much less than with chilled water pipes systems. Large amount of energy is required to melt the ice (latent heat transfer), it is due to this effect that the amount of water required to cool the air is much less. It is said that the quantity of water to be pumped back to the surface is approximately a quarter of the quantity required for chilled water pipes. It was mentioned that in general, the concept is usually feasible for mines with a depth of 3000 m or more despite the expensive cost of ice making (Sheer, Butterworth, & Ramsden, 2001). It was discovered that steel pipes were unsuitable for ice conveying but plastic piping have been so far successful. There are many considerations when designing ice conveying systems but the experience that the South Africans have acquired throughout the years could be applied to Canadian mines. The main advantage of the Canadian climate is that the snow could be collected and stored during the winter assuming that the region has enough snow accumulations; therefore the expensive capital and operating cost of ice making machines would be eliminated. The cost would include the collection the snow, (most probably with the use of snow blowers) and the installations to store it.
9.09 New design proposal: surface ice formation, ice storage for ice conveying to the underground levels As intake air requires to be heated during the winter period, one way to warm the air is to spray water as at the Stobie mine ice stope outlined in section 9.07. The idea is as follows; prior to entering the intake shaft on surface, water would be sprayed inside a building to heat the intake 119
air, the ice would then fall onto a conveyor belt to be discharged in a large insulated silo. The silo would be connected to the ice conveying system carrying the ice underground during the warm season. The schematic of the system is shown in Figure 9-7.
Figure 9-7: Surface ice formation for heating, ice storage for ice conveying to the underground levels
The main consideration with this design is to ensure that the ice does not turn into large blocks inside the silo. In that case, it would be very difficult for the ice to be carried inside the pipes as it would get stuck inside the silo or the piping system. The formation of ice and its behaviour within the silo should be studied. If large ice blocks are inevitably formed inside the silo from compression, one possibility would be to use a shredder to break the ice blocks prior to the discharge in the pipes. An alternative design would be to use the ice to cool the intake air on surface instead of underground. Intake air would flow through ice within a compartment on surface. A proposed design would be to have the air flowing below the silo so that ice is discharged with gravity within the compartment. Another possibility would be to use a conveyor belt to move the ice within the silo back to the intake shaft surface building. This system could also be feasible even if no cooling is required, ice would be piled up and melted to the ambient air during the warm season and no silo would be required. 120
In order to implement this design, the total annual volume of ice produced should first be calculated to size the silo. First, the required heat demand of mine fresh air should be calculated from the predicted ambient temperature. The amount of power to deliver to the air so that it achieves a 0°C temperature is calculated from Equation 9-5.
Eq. 9-5
qair m air C p,air (0 Tamb )
The heat transferred from the water is from both sensible and latent heat. The sensible heat transfer depends on water temperature when leaving the sprays and mass flow rate. As water leaves the spray, it cools until it reaches a 0°C temperature. Then, the water will turn into ice and the latent heat of fusion will be transferred to the air (hfg= 333 kJ/kg). The total heat transferred to the air will therefore be equal to Equation 9-6. The first part of the equation calculates the sensible heat and the second the latent heat transfer.
q w m w C p, w (Tw,in 0) h fg m w
Eq. 9-6
Using Equation 9-5 and 9-6, the total mass of ice formed is found from Equation 9-7.
qair mw C p, w (Tw,in 0) h fg
Eq. 9-7
Calculation example of volume of ice formation Assumptions: Air flow rate of 300 m3/s (360 kg/s) Ambient air temperature of -20°C Water inlet temperature of 5°C Cp,w=4.2 kJ kg-1 °C-1 Cp,air=1.005 kJ kg-1 °C-1 From Equation 9-7, the mass flow rate of water formed would be: 18 kg/s. For an ice density of approximately 920 kg/m3, the rate of volume of ice generated will be: 0.0196 m3/s. Assuming 30 days at these conditions, the total volume of ice formed would be: 50 700 m3
9.10
Conclusions
New and existing innovative design proposals were introduced which could be interesting for future research projects. A simplified explanation of the vapour compression cycle was also introduced to better understand the fundamentals of cooling systems. Each operating underground mine has different characteristics and all different options for cooling systems should be evaluated. The use of natural cooling systems is very interesting especially as it involves minimal operating costs. If a large amount of broken rocks that connects underground levels to surface is present, NHEA would most likely be the best option to start off with. In the case that large stopes at proximity to the surface are available, an ice stope system 121
could be a great possibility as it involves a minimal capital cost and low maintenance. Another natural cooling system is the storage of ice on surface for conveying to the underground levels during the warmer period. The storage of ice can be from snow accumulations or intake air heating with water sprays. If a natural cooling system is not possible, sufficient or feasible, several possible mechanical refrigeration designs are available. There are many factors to take into consideration when designing a refrigeration plant such as air conditions (humidity and temperature), rock temperatures and the mining operation parameters such as the geometry of the mine, the mining method, the mine life etc. It was also found that surface mechanical cooling plants were proven to be inefficient at underground mine sites in milder climates. As more and more Canadian mines will require the use of refrigeration plants, lessons learned and experiences must be shared to avoid the use of inefficient or insufficient machines resulting in higher capital and operating costs. For any mine site purchasing such systems, knowledge of the refrigeration plants technologies must be acquired to ensure that the system provided is the most efficient and feasible design available.
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CHAPTER 10. GENERAL CONCLUSIONS AND RECOMMENDATIONS FOR FURTHER WORK 10.01 Overview This work has demonstrated that exhaust air heat recovery systems in underground mines have great potential. The factors influencing the economics of an exhaust air heat recovery system have been narrowed down to create a feasibility study software tool. The software tool enables any underground mine site to quickly evaluate the return on investment of the installation of an exhaust air heat recovery system. The fictitious case studies showed that there are several mines that should adopt such a system as it would result in a fast return on investment with several millions of dollars in energy cost savings throughout the life of the mine. The two existing projects; at Kiena and Williams mines, have proven that these systems can run effectively with minimal maintenance required. As energy prices will inevitably rise, these systems will become increasingly attractive. The implementation of a heat recovery system should be studied during the development phase of the mine as surface ventilation installations should be designed to accommodate the components of the system. In some cases where heating cost is expected to be elevated, the study of a heat recovery system could even influence the positioning of the ventilation shafts. Heat recovered from compressor coolers and mine water can be combined with an exhaust air heat recovery system to maximize the efficiency of the system. On the other hand it can sometimes be more economically viable to recover heat solely from these other heat sources as the exhaust air heat would be of too low grade or/and located too far from the heat sink. These systems have been found to exist in both configurations. Combining a heat recovery system with surface building facilities heating could improve the efficiency of the system as it would be utilized for a longer period of the year. It could also be more economically viable to solely use the heat recovery system for space heating of buildings. An extensive energy assessment should be conducted at the mine site to evaluate the best possible design.
10.02 Main goals and objectives The main goal of this study was to evaluate the feasibility of heat recovery systems in underground mines and investigate the means of reducing the adverse effects of adiabatic compression. The most common way to alleviate the adverse effect of adiabatic compression is the use of cooling plants; some novel cooling systems were therefore described within this thesis. To achieve these objectives, the following partial goals were defined:
Research existing heat recovery systems or under study projects in underground mines. Investigate different designs of heat recovery systems. 123
Research the theory on performances of tube and fin HEs. Choose the most efficient and feasible exhaust air heat recovery design and study all of its components into detail. Development of a software tool to evaluate the feasibility of the exhaust air heat recovery system design for underground mines. Research on available technologies of mechanical cooling systems in underground mines. List recommendations and novel cooling systems in underground mines. Research on the theory of adiabatic compression. Research and develop new ideas for mine cooling other than mechanical or natural cooling.
The main accomplishments of this research can be summarized as follow:
Most of the existing systems and projects under study on heat recovery have been evaluated and recommendations for mine application have been formulated. A software tool has been developed to evaluate the feasibility of exhaust air heat recovery systems at any underground mine site. A software tool has been developed to design tube and fin HEs as they are a major component of the exhaust air heat recovery system. Alternate designs of closed-loop glycol circuit have been outlined with some technical information available and recommendations. Existing heat recovery projects and studies have been listed with recommendations. The theory of adiabatic compression has been reviewed, ideas to reduce its adverse effect have been proposed. Novel cooling systems for underground mines have been listed with several design proposals and recommendations for future Canadian mine applications.
10.03 Recommendations for further work Among the several topics that have been investigated in this thesis, there are a few where further research could be carried out in the future. 10.03.1 Energy calculations and tube and fin HE: In calculating energy savings, it was assumed that the ambient air conditions are the same for one full month. A study should be performed to determine the accuracy of this method. A more accurate way could be to estimate the total number of hours within a certain temperature range from historical weather data. Also, it should be determined how many years of historical records should be considered in order to have a better approximation of temperatures, 5 years were chosen in this thesis. The natural gas companies might have such information as they try to predict the annual consumption of their customers. The software developed could be supplemented with a pipe heat loss calculation software tool to determine the best insulation material to choose from. The heat losses would take into consideration underground pipes and pipes exposed to ambient air. 124
The fouling effect on the heat transfer at exhaust should be studied to obtain the actual merit of an automated washing system. As the pressure drop on the air-side of the HE will be in most cases the major portion of the operating costs of the system, more research should be performed on correlations for pressure drop and friction factors. In order to optimize the design of the tube and fin HE, the HE design software should provide accurate results for both the efficiency and pressure drop. The geometry could then be chosen from the optimal trade-offs between the heat transfer and the pressure drop. The HE design software could then be implemented within the feasibility study software and validate the optimal HE geometry. Further research should be performed on HE performances under humid conditions to validate the present assumptions on the HE efficiency. 10.03.2 Capital cost approximation and design recommendations Another improvement on the feasibility study software tool would be the review of the capital cost of the pumping system. Manufacturers should be contacted to obtain a wider range of prices. Additional options within the software should include the capital cost of heat recovery using other heat sources (mine water, compressor coolers) as well as the economics of surface building heating. It is expected that quite a few exhaust air heat recovery systems will be implemented within the years to come. For a new project construction, the design recommendations outlined in CHAPTER 5 should be reviewed. Also, any component which was not included in this work could be added to the feasibility study software tool. Costing data should also be reviewed and modified accordingly. 10.03.3 Alternative designs The economics of alternative designs should be studied more profoundly to derive their cost benefits. The design of spray chambers at exhaust and glycol tube and fin HE at intake should be prioritized as this was found to be very promising. More research should be carried out on the spray chambers and its effect on the air cleaning process; how effective it is to remove pollutants within the air, which pollutants are removed within the air and how beneficial it is for a clean environment. More research can be carried out for other heat sources available mine sites. It is however difficult to generalize them for all mine sites. More and more energy evaluations should be performed and documented at mine sites to continuously build up the knowledge base on heat recovery systems. 10.03.4 Cooling Some more research should be performed on the efficiency of novel natural heating and cooling systems in Northern climates of the world. These systems are very efficient and could become the number one way to cool and heat the intake fresh air all across Northern regions of the world. The Creighton’s natural heat exchanger area and Stobie’s ice stopes were created from existing mining configurations hence the cost to accommodate these systems was relatively low. It would be interesting to study the effect of modifying the geometry of the mine, for example conduct 125
additional blasting to create an NHEA at a mine site. The economics of the new design proposals; ice formation and storage on surface, ice conveying to the underground, should also be evaluated and reviewed.
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McCallum. (1969, October). Design of mine air heating plant at Strathcona mine, Falconbridge Nickel Mines Ltd. 62-65. McPherson, M. J. (1993). Subsurface Ventilation and Environmental Engineering. London: Chapman & Hall. McQuiston, F., Parker, J., & Spitler, J. (2005). Heating ventilating and air conditioning (sixth edition ed.). John Wiley and Sons Inc. Mining-Technology. (n.d.). Retrieved June 20, 2009, from http://www.miningtechnology.com/contractor_images/bluhm/1_Fans.jpg Mon, M. S., & Gross, U. (2004). Numerical study of fin-spacing effects in annular-finned tube heat exchangers. International Journal of Heat and Mass Transfer , 47, 1953-1964. Moreby, R. (2007). Metal mine ventilation short courses. Australia: University of New South Wales. NRCAN. (2011). Petroleum Products and Crude Oil Prices. Retrieved 04 2011, from Nat. Resources Canada: http://www.nrcan.gc.ca/eneene/sources/pripri/index-eng.php Page, J. S. (1999). John S. Estimator's piping man-hour manual (5th Edition ed.). Gulf professional Publishing. Pelletier, M.-A. (2009, December). Private communication. Hemlo, Ontario: Barrick Gold. Pyhäsalmi. (2010, April). Pioneering underground innovation. International mining , 12-18. Quirion, C. (2009). Private communication. Rouyn-Noranda: Agnico-Eagle, Mine Laronde. Radermacher, R., & Hwang, Y. (2005). Vapor Compression Heat Pumps: With Refrigerant Mixtures. CRC Pr I LIc. Ramsden, R., & Sheer, T. (2001). Design and simulation of ultra-deep mine cooling systems. In Raymond, J. (2006). Low-Temperature Geothermal Potential of the Gaspe Mines, Murdochville. Master's Thesis, Universite Laval, Departement de Geologie. Raymond, J., Therrien, R., & Gosselin, L. (2010). Low-temperature geothermal energy in mining environments. CIM , 1 (2). Rieger, S. (1921). The genesis and classification of cold soils. New york: Academic Press. RSMeans Co. (2010). Assemblies cost data. Kingston, MA. RSMeans Co. (2008). Construction cost data. Kingston, MA. RSMeans Co. (2010). Electrical cost data. Kingston, MA. RSMeans Co. (2010). Mechanical cost data. Kingston, MA. RSMeans Co. (2010). Site work cost data. Kingston, MA. Ruiter, H. d. (1992, October). Energy savings in mine. Mining Magazine , 240-245. Sabau, A. (2010, December). Private communication. Shaddock, M. (2010). Private Communication. Sudbury, Ont.: Barrick Gold, Williams mine. Shah, R. K., & Sekulic, D. P. (2003). Fundamentals of heat exchanger design. John Wiley and Sons Inc,. Sheer, T. J., Butterworth, M., & Ramsden, R. (2001). Ice as a coolant for deep mines. In S. Smith, D. C., & Arthur, H. L. (1996). Mine Ventilation: Waste Heat Recovery. CIM Bulletin , 89 (998), 126-130. Sylvestre, M. J.-G. (1999). M. Sc. thesis: Heating and ventilation's study of INCO's Creighton mine. McGill University, Department of Mining and Metallurgical Engineering. Templeton, J. (1950). The application of the heat pump to the mining industry. 53, 235-247. Thomas, B. (2008). Private communication. Wisconsin: Industrial Heat Transfer Inc.
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130
APPENDIX A
Intake coil specifications sheet
I
Exhaust coil specifications
II
Specifications of Adjustable Saddle Support, Anvil International©.
III
Foamglas® Insulation specification sheet
IV
Plate heat exchanger quote, Thermofin©
V
Plate heat exchanger quote, Thermofin©
VI
Plate heat exchanger quote, Thermofin©
VII
Cities
factor
Alberta
Newfoundland
Peterborough
117.1
Saint-Hyacinthe
113.2 113.8
Corner Brook
118.7
Sarnia
116.8
Sherbrooke
Calgary
129.4
St-Johns
119.2
Sault Ste Marie
111.7
Sorel
Edmonton
130.6
Northwest territories
St. Catharines
111.4
St Jerome
Sudbury
111.4
Trois-rivieres
Thunder Bay
112.9
Saskatchewan
Fort McMurray
126
Yellowknife
120.5
114 113.4 114
Lethbridge
119.9
Nova Scotia
Lloydminster
115.1
Bridgewater
114.8
Timmins
117.3
Moose Jaw
112.2
Medicine Hat
115.3
Dartmouth
116.2
Toroonto
122.5
Prince Albert
111.1
Red deer
115.9
Halifax
117.4
Windsor
Regina
114.2
Saskatoon
112.7
British Columbia
112
New Glasgow
114.2
PEI
Kamloops
116.4
Sydney
111.6
Charlottetown
116.6
Yukon
Prince George
117.6
Truro
114.2
Summerside
116.1
Whitehorse
Vancouver
128.7
Yarmouth
114.1
Quebec
Victoria
117.7
Ontario
Manitoba
Barrie
118
Cap-de-la-Madelaine
113.8
Charlesbourg
113.8
Brandon
115
Brantford
117.1
Chicoutimi
112.8
Portage la Prairie
115
Cornwall
116.9
Gatineau
113.4
127.7
Hamilton
121.9
Granby
113.7
Winnipeg New Brunswick
Kingston
117.8
Hull
113.6
Bathurst
113.2
Kitchener
113.3
Joliette
114.1
Dalhousie
113.2
London
120.6
Laval
113.5
Fredericton
115.8
North Bay
117.1
Montreal
120.8
Moncton
113.5
Oshawa
Quebec
119.4
Newcastle
113.2
Ottawa
121.5
Rimouski
113.3
St. John
115.9
Owen Sound
118.1
Rouyn-Noranda
113.4
117
112
Canadian location factors for capital cost approximation (RSMeans Co., 2010)
Depth cm 5 10 20 50 100 150 300
1981 January -1.3 -0.8 -0.45 0.3 1.5 2.5 5.2
February -0.75 -0.4 -0.2 0 0.8 1.6 4.1
Average soil temperature (°C) 1984 January February -1.3 -0.3 -0.25 0.5 2.1 3.0 5.9
-1.3 -0.3 -0.2 0.0 1.2 2.2 4.8
1978 January -3.3 -2.8 -2.2 -0.6 1.1 2.2 5.6
February
Total average
-2.8 -2.8 -1.7 -1.1 0.6 1.7 5.0
Soil Temperature Data, Canada, for Val d’Or (Centre climatologique Canadien, 1977-1984)
VIII
-1.8 -1.23 -0.83 -0.15 1.22 2.2 5.1
APPENDIX B HE building extension cost Includes: Trench excavation, hand trim, compacted backfill, formwork (4 uses), keyway form (4 uses), reinforcing, dowels, concrete, place concrete, direct chute, screed finish. Material
Labour and equipment
$/ft
hrs/ft 7.35
12.75
Table 1 : Standard foundation strip footing; load 2.6 KLF, soil capacity 3 KSF, 16” wide x 8” deep plain, ref : R.S . Means assemblies 2010, page 6
Height
Width
Price
labour
$/ft floor
hrs/ft2 floor
2
ft
ft
10
10
8.30
5.50
14
14
8.80
0.160
16
16
9.30
0.175
20
20
10.25
0.204
24
24
9.40
0.175
24
40
9.40
0.175
24
100
7.85
0.093
24
120
6.70
0.079
24
150
6.05
0.073
24
200
5.70
0.063
Table 2 Pre-engineered steel buildings: ref: R.S. means bldg construction 2008, page 397 Material 2
Labour and equipment hrs/ft2
$/ft
1.84
0.041
Table 3 Slab on grade 4” thick, non industrial and non reinforced, ref: R.S. Means assembly 2010, page 24 Unit price 2
labour hrs/ft2
$/ft
0.38
0.007
Table 4 Insulation, 1.5” thickness, R5, Vynil/scrim/foil, ref: R.S. Means bldg construction 2008, page 400
IX
Main piping system Labour required to remove 1 B.C.Y.
Nature of soil
hrs Common earth
0.080
Loam and sandy clay
0.074
Sand and gravel
0.073
Dense hard clay
0.091
Table 5 Trench digging labour time with respect to the soil nature for ½ C.Y. excavator, ref: R.S. bldg construction 2008 page 284 Note: (B.C.Y.) Bank Cubic Yards is defined as it lies in its natural/undisturbed state prior to extraction from the earth.
Labour
Material
hrs/L.C.Y.
$/L.C.Y.
0.160
35
Table 6 Utility bedding for pipe and conduit, not incl. compaction, Crushed stone ¾” to ½” , ref: R.S. Means site work 2010, page 301 Labour hrs/L.C.Y. 0.235
Table 7 By Hand backfill, compaction in 6” layers, vibrating plate., ref: R.S. Means site work 2010, page 299 Note: It will be assumed that the B.C.Y. of the trench excavated is equal to the L.C.Y. backfilled
NPS in 2
pipe $/ft
pipe hours/ft
1 groove hrs
2 grooves
couplings
couplings
elbow
elbow
hrs
$
hrs
$
hrs
2.2
0.15
0.138
0.276
16.4
0.16
21.5
0.32
4
9.65
0.292
0.186
0.372
31
0.32
41.5
0.64
6
19.67
0.5
0.2
0.4
53
0.48
116
0.96
8
30.4
0.565
0.242
0.484
83.5
0.571
243
1.143
10
45
0.674
0.276
0.552
149
0.686
440
1.333
12
57
0.774
0.348
0.696
167
0.75
705
1.6
14
56
1.08
0.533
1.066
192
1
835
2
16
67.6
1.271
0.571
1.142
250
1.2
1075
2.182
18
77.4
1.543
0.593
1.186
289
1.333
1375
2.133
20
94.3
1.8
0.64
1.28
395
1.5
1825
2.462
24
175
2.16
0.696
1.392
505
1.846
2625
2.909
Table 8 Labour and material cost of pipe, couplings, pipe grooving and elbows, ref: R.S. Means mech 2010 Pipe price sch40 without coupling and hanger; page 167, couplings rigid style; page169, grooving; page 174, elbow; page 168
X
unit price
NPS in.
$
Labour hours
2
12
0.1328
4
21.2
0.1392
6
31.28
0.1424
8
47.2
0.1456
64
0.1456
12
72.8
0.1488
14
140.8
0.1488
16
150.4
0.152
18
166.4
0.152
20
182.4
0.1568
24
222.4
0.16
10
Table 9 Man hours and unit price for: Saddles, pipe support, complete, adjustable, CI saddle, TYPE 36 RS, pipe insulation 2” thick with stanchion. ref: R.S. Means mech 2010
Ins. Foamglas®
NPS in
$/ft
Elbow Ins. Foamglas®
Insulation hrs/ft
$
Elbow Ins. hrs
2
5.15
0.114
14.85
0.342
4
6.95
0.128
28.78
0.384
6
11.05
0.145
43.88
0.435
8
14.3
0.168
59.1
0.504
10
14.75
0.188
98.38
0.564
12
22.5
0.2
139.42
0.6
14
25
0.213
181.18
0.639
16
28
0.229
240.22
0.687
18
30
0.246
279.24
0.738
20
35
0.267
410.52
0.801
24
50
0.291
569.6
0.873
Table 10 Glass cellular insulation material cost and labour time, ref: R.S. Means mech 2010 page 107
XI
Pump and MCC cost Flow
Head
Motor required
Pump Price
m3/s
ft
hp
$
0.06309 (1000 gpm)
0.12618 (2000 gpm)
0.31545 (5000 gpm) 0.6309 (10000 gpm)
50
25
11500
100
50
11500
200
100
11500
300
150
11500
50
50
11600
100
100
11600
200
200
13700
300
300
18800
50
50
13700
100
100
13700
150
150
21400
50
250
23300
100
500
23300
Table 11 Pump pricing and motor required Pumps, Process Medium duty, Centrifugal, ref (CostMine, 2004): Electric motor AC power hp
Rotational speed rpm
25 50 100 150 200 300 400 500
unit price $
1200
1894
3600
1150
1200
4519
3600
2911
1200
7952
3600
5481
1200
10826
3600
8926
1200
13631
3600
11904
1200
22517
3600
17946
1200
23566
3600
22465
3600
30969
Table 12 Electric motor cost, ref: (CostMine, 2004)
XII
Motor Power
unit price $
Labour
Lab time
$
hrs
10
1500
1525
20.93342
25
2500
1975
27.1105
50
8350
3200
43.92588
100
27600
4550
62.4571
150
33800
5850
80.30199
200
33600
7175
98.49005
300
59000
10305
141
400
78000
13258
182
500
98000
16211
222
Table 13 MCC material and cost installation ref: R.S. Means Assembly 2010
Labour time has been calculated assuming an electrician labour rate of 72.85$. For the 300, 400 and 500 hp motor the material cost and man hours has been assumed to be a linear function from the lower horsepower values: Material cost: y = 195.32x + 475.86 Man hours: y = 0.4054x + 19.386.
Price includes; Steel intermediate conduit, Wire Motor Power
unit price
Labour
Labour
hp
$/ft
$/ft
hrs/ft
10
3.19
8.05
0.111
25
6.95
11.85
0.163
50
23
16.7
0.229
100
50
30.5
0.419
150
70
41
0.563
200
99.5
61
0.837
300
149
85
1.17
400
200
112
1.53
500
251
139
1.90
Table 14 Motor feeder length, ref: R.S. Means Assembly 2010 page 365 Labour hours have been calculated assuming an electrician rate of 72.85$. For the 300, 400, 500 hp motor, a linear relationship is assumed.
XIII
Piping accessories NPS in.
unit price
Labour
$
hours
2
420
0.211
4
560
0.421
6
860
0.632
8
1350
0.889
10
1975
1.2
12
2550
1.5
14
8150
1.6
16
10100
1.714
18
14474
2.056
20
18248
2.286
24
25796
2.746
Table 15 Strainer tee type, ref: R.S. means mech 2010 page 172 For material cost: y 1887 x 19492 , For labour time: y 0.115x 0.014 NPS in.
unit price $
Labour hours
2
33
0.471
4
70
0.941
6
189
1.412
8
415
1.714
10
860
2
12
1200
2.4
14
1200
2.667
16
1275
3
18
1575
2.909
20
2250
3.2
24
3450
4
Table 16 Tee, painted, grooved joint, ref: R.S. Means mech 2010 page 169
XIV
NPS in.
unit price $
Labour hours
2
17.80
0.16
4
51
0.32
6
60
0.48
8
98
0.571
10
162
0.686
12
185
0.750
14
198
1
16
259
1.2
18
305
1.333
20
405
1.5
24
525
1.846
Table 17 Grooved Joint, Tee reducing, painted, ref: R.S. Means mech 2010, page 170 NPS in.
unit price $
Labour hours
2
199
0.258
4
249
0.421
6
490
0.632
8
670
0.889
10
1950
1.2
12
2300
1.5
14
2567.9
1.713
16
3020.5
1.971
18
3473.1
2.229
20
3925.7
2.487
24
4830.9
3.003
Table 18 Check valve, grooved joint, ref: R.S. Means mech2010, page 173
y 226.3x 600 Material cost: Labour hours: y 0.29 x 0.093
XV
NPS in.
unit price $
Labour hours
2
223
0.211
4
325
0.421
6
610
0.632
8
1600
0.889
10
2900
1.2
12
3325
1.5
14
4025
1.6
16
5825
1.714
18
7175
2.667
20
8825
2.909
24
11800
3.2
Table 19 Butterfly valve, grooved joint, with stainless steel trim, ref: R.S. Means mech 2010, page 173 NPS in.
unit price $
Labour hours
2
29.50
0.8
4
35
1.6
6
64
2.4
8
113
3.429
10
180
4
12
262
4.8
14
380
5.333
16
570
8
18
800
9.6
20
975
10.435
24
1300
12
Table 20 Welding neck flange, 150 lb ref: R.S. Means mech 2010, page 163 NPS in.
unit price $
Labour hours
2
320
1.6
4
525
5.333
6
865
8
8
1550
9.6
10
2825
10.909
12
3750
14.118
14
7175
18.462
16
10800
24
18
13900
30
20
20100
40
24
28700
48
Table 21 Gate valve OS&Y,125 lbs flanged ref: R.S. Means mech 2010 page 242;
XVI
NPS in.
unit price $
Labour hours
2
360
1.231
4
440
2
6
545
2.667
8
650
3.2
10
720
3.478
12
890
4
14
1100
4.211
16
1275
5.517
18
1425
6.4
20
1500
7.619
24
1750
8.889
Table 22 Expansion joints, 10” face to face, bellows type neoprene cover, flanged spool ref: RS Means mech 2010 page 275 Capacity
unit price
gal
$
Labour hrs
30
580
1.333
40
680
1.6
60
815
2
80
875
2.286
100
1175
2.667
120
1275
3.2
135
1325
3.556
175
2075
4
220
2350
4.444
240
2450
4.848
305
3450
5.333
400
4250
5.714
Table 23 Steel liquid expansion tank, ASME, painted, ref: R.S. Means mech 2010 page 276 Dia. in.
unit price
Labour
$
hours
4” x 3”
36.50
0.552
6” x 4”
58.50
0.923
8” x 6”
152
1.043
10” x 8”
310
1.200
12” x 10”
555
1.500
Table 24 Reducer concentric, painted, ref: R.S. Means mech 2010 page 170
XVII
Automated washing system cost Soil nature
equipment $/ft
Labour hrs/ft
Common earth
0.66
0.0268
Loam and sandy clay
0.64
0.0262
Sand and gravel
0.59
0.0249
Dense hard clay
0.7
0.0275
Table 25 Trenching 2’ wide, 2’ deep, using 3/8 C.Y. bucket ref: R.S. Means site work 2010 page 563 unit price $
Labour hours
2.11
0.071
Table 26 One hole vertical mounting malleable iron 2” pipe size, ref: Rsmeans mech 2010 page 95 unit price $/ft
Labour hrs/ft
14.25
0.0289
Table 27 Pipe stainless steel Schedule 10 type 304; RSmeans mech 2010 page 177 unit price $/ft 1.55
Table 28 Polyethylene, flexible, no couplings or hangers, 2” dia; Ref: RSmeans mech 2010
unit price $
Labour hrs
675
0.750
Table 29 Electric motor actuated two way screwed 2” dia ref: RSmeans mech 2010, page 260
unit price $/ft
Labour hrs
2.35
0.0947
Table 30 Motor feeder systems, 115 V, 1 hp, ref: (CostMine, 2004) RSmeans electrical 2010 page 372
XVIII
APPENDIX C -For air μ: Dynamic viscosity: From 0 to 25°C μ = -2.3413*(E-13)*T3 - 2.0916*(E-11)*T2 + 5.0503*(E-08)*T + 1.7197*(E-05) Pr: Prandlt number: From 0 to 80°C Pr = 2.604167*(E-10)*T4 - 3.125000*(E-08)*T3 + 1.1458*(E-06)*T2 - 1.125*(E-04)*T + 7.15*(E-01) k: Thermal conductivity From 0 to 80°C k=(7*(10-5)*T)+(2.43*(10-2)) -For water μ :Dynamic viscosity: From 0 to 80°C μ = 4.6535*(E-11)*T4 - 1.1120*(E-08)*T3 + 1.0723*(E-06)*T2 - 5.6190*(E-05)*T + 1.7775E-03 Pr: Prandtl number: From 0 to 80°C Pr = 5.5563*(E-07)*T4 - 1.2490*(E-04)*T3 + 1.1013*(E-02)*T2 - 5.0866*(E-01)*T + 1.3649*(E+01) Cp: Specific heat: From 0 to 25°C (correlations shall be developed for wider range of temperatures) Cp=-2.785*10-10*T6+3.968*10-8*T5-2.162*10-6*T4+ 5.543*10-5*T3-6.233*104*T2 +7.7938*104 *T+4.21 ρ: density of water: from 0 to 80°C ρ =-4.684*10-8*T4+3.456*10-5*T3-7.575*10-3*T2+6.123*10-2*T+998.5 k :thermal conductivity of water k = -7.2627*10-6*T2 + 1.8430(10-3)*T + 5.6908*(10-1) Note : All temperatures are in Celsius
XIX
Design no.1 (Laronde design case) HE geometry inputs Xl Xt L1 fin thickness δ Number of fins per m tube inside diameter tube wall thickness number of tubes per row(1st row) number of tube rows no. of circuits
m m m m m m unit/m m m
0.0413 0.0381 1.9 0.0003 305 0.01466 0.00061 46 8 92
Inlet air Inputs Vol. flow rate (across frontal area) Dry bulb temperature Wet bulb Temperature Altitude
m /s °C °C m
3
32.1 18 18 335
Inlet running fluid Inputs Inlet temperature total mass flow rate
°C kg/s
1.5 14.7
m m m m m m unit/m m m
0.05 0.05 2.5 0.0003 100 0.00952 0.00061 18 4 35
Design no. 2 HE geometry inputs Xl Xt L1 fin thickness δ Number of fins per m tube inside diameter tube wall thickness number of tubes per row(1st row) number of tube rows no. of circuits Inlet air Inputs Vol. flow rate (across frontal area) Dry bulb temperature Wet bulb Temperature Altitude
m /s °C °C m
3
22 15 10 0
Inlet running fluid Inputs Inlet temperature total mass flow rate
°C kg/s
1.5 8
XX
Design no. 3 HE geometry inputs Xl Xt L1 fin thickness δ Number of fins per m tube inside diameter tube wall thickness number of tubes per row(1st row) number of tube rows no. of circuits
m m m m m m unit/m m m
0.05 0.04 5.6 0.0003 250 0.018 0.00061 18 4 35
Inlet air Inputs Vol. flow rate (across frontal area) Dry bulb temperature Wet bulb Temperature Altitude
m /s °C °C m
3
30 22 10 0
Inlet running fluid Inputs Inlet temperature total mass flow rate
°C kg/s
1.5 10
m m m m m m unit/m m m
0.03 0.04 7.62 0.0003 200 0.012 0.00061 36 4 36
Design no.4 HE geometry inputs Xl Xt L1 fin thickness δ Number of fins per m tube inside diameter tube wall thickness number of tubes per row(1st row) number of tube rows no. of circuits Inlet air Inputs Vol. flow rate (across frontal area) Dry bulb temperature Wet bulb Temperature Altitude
m /s °C °C m
3
28 10 8 335
Inlet running fluid Inputs Inlet temperature total mass flow rate
°C kg/s
1.5 12
XXI
Design no. 5 HE geometry inputs Xl Xt L1 fin thickness δ Number of fins per m tube inside diameter tube wall thickness number of tubes per row(1st row) number of tube rows no. of circuits
m m m m m m unit/m m m
0.06 0.04 1.5 0.0003 200 0.02 0.00061 36 4 36
Inlet air Inputs Vol. flow rate (across frontal area) Dry bulb temperature Wet bulb Temperature Altitude
m /s °C °C m
3
35 20 8 0
Inlet running fluid Inputs Inlet temperature total mass flow rate
°C kg/s
1.5 20
m m m m m m unit/m m m
0.0413 0.0381 1.9 0.0003 250 0.01466 0.00061 46 8 92
Design no. 6 HE geometry inputs Xl Xt L1 fin thickness δ Number of fins per m tube inside diameter tube wall thickness number of tubes per row(1st row) number of tube rows no. of circuits Inlet air Inputs Vol. flow rate (across frontal area) Dry bulb temperature Wet bulb Temperature Altitude
m /s °C °C m
3
32.1 18 18 335
Inlet running fluid Inputs Inlet temperature total mass flow rate
°C kg/s
1.5 14.7
XXII
