HEAT RECOVERY SYSTEMS IN UNDERGROUND MINE VENTILATION

©Hugo Dello Sbarba

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

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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

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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 Main piping system .................................................................................................... 52

5.07.1 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 Piping accessories ....................................................................................................... 59

5.09.1 Strainer .................................................................................................................... 60

5.09.2 Air bleed lines ......................................................................................................... 60

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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

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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

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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 5-25: Heat recovery system economics, case 1 ................................................................. 71

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Figure 5-26: Heat recovery system economics, case 2..................................................................71








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

©Hugo Dello Sbarba

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.

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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.

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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

8

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 air-

side 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.

9

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.

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.

Figure 1-9: Cross flow tube and fin HE (Jun-Jie & Wen-Quan, 2005)

10

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.

11

Chapter 5 covers technical considerations in the design and building of each of the components

of the heat recovery system.

12

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.

13

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

14

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

15

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; closed-

loop 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

16

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.

17

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).

18

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

2001 2002 2003 2005 2006 2007 2009 2010 2011

$0

$5

$10

$15

$20

$25

$30

Canadian average natural gas and propane price since 2001

Propane

Natural gas

year

pri

ce

($

/GJ)

19

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.

20

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.

21

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.

22

Nomenclature

A Area (m2) T Temperature (°C)

C Flow stream heat capacity rate (kW °C-1

) dbT Dry bulb temperature (°C)

pc Specific heat of fluid at constant pressure (kJ kg

-1 K

-1) (kJkg

-1 °C

- wbT Wet bulb temperature (°C)

id Pipe inside diameter (m) mu Mean velocity across cross-sectional area (m s

-1)

f Friction factor, dimensionless UA Overall thermal conductance (W °C-1

)

h Head loss (kPa) W Humidity ratio (kghumidity/kgdry air)

k Thermal conductivity (kW m-1

°C-1

) Z Altitude (m)

K

Pressure loss coefficient, dimensionless Greek symbols

fgh Specific enthalpy of phase change (kJ kg-1

) Heat exchanger efficiency, dimensionless

L Length (m) fan Fan efficiency, dimensionless

m Mass flow rate (kg s-1

) (kg s-1

mot Fan electric motor efficiency, dimensionless

daysn Number of calendar days within a given a month

Density (kg m-3

)

NTU Number of transfer units, dimensionless Difference

Nu Nusselt number, dimensionless Kinematic viscosity (m2 s

-1)

bP Barometric pressure (kPa) Fraction of mixture, dimensionless

sensP Power recovered from sensible heat (kW) Subscript

latP Power recovered from latent heat (kW) c Cold fluid

Pr Prandtl number, dimensionless f Film

Q Volumetric flow rate (m3 s

-1) h Hot fluid

R Thermal resistance (°C W-1

) i Inlet

Re Reynolds number, dimensionless o Outlet

S Cost ($CAN)

4.02 Calculation of heat capacity rate of air at exhaust

The heat capacity rate of exhaust air is given by Equation 4-1.

airdryairairpair QcC

,, Eq. 4-1

The barometric pressure is first estimated from the altitude of the mine site.

1907510*3.101

Z

bP

Eq. 4-2

The density of dry air is then calculated from Equation 4-3.

dryair

bdryair

T

P

,,

287.0 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.

23

airpc , 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.

24

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:

)( ,, icihairsenstot TTCPP 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.

25

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:

)( ,,min

,, icihh

ihoh TTC

CTT 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.

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

Figure 4-3: Outlet conditions predicted below saturation line

26

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.

)( ,, icihairsens TTCP Eq. 4-6

)(, outinairdryfglat WWmhP

Eq. 4-7

senslattot PPP 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:

3.237

27.17exp6105.0

dry

drydry

T

TP Eq. 4-9

3.237

27.17exp6105.0

wet

wetwet

T

TP Eq. 4-10

)(000644.0 wetdrybwetu TTPPP Eq. 4-11

ub

u

PP

PW

622.0 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:

drywetu PPP Eq. 4-13

And the saturation line function is determined from Equation 4-14.

3.237

27.17exp6105.0

3.237

27.17exp6105.0

622.0622.0

dry

dryb

dry

dry

dryb

dry

T

TP

T

T

PP

PW 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.

27

To determine if condensation occurs, the assumed wall temperature is inserted into Equation 4-

14. 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:

walldryi

wallexhi

TT

WWa

,

, Eq. 4-15

The value of W when T is equal to 0 is given as:

wallwall WaTb Eq. 4-16

The function of the line is found from Equation 4-17.

baTW db 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.

dbfg Th 805.12501 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.

28

Relative

Humidity

Power

recovered

Mass flow rate

of condensate

% kW kg/s

0 417.4 0

10 417.6 0

20 417.9 0

30 418.5 0

40 418.8 0

50 427.9 0

60 449.0 0.068

70 477.4 0.126

80 511.4 0.191

90 551.9 0.248

100 591.4 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.

)( ,,min icih

fgcondtot

TTC

hmP

Eq. 4-18

Relative

Humidity

Efficiency of

HE

%

0 0.675

10 0.675

20 0.675

30 0.675

40 0.675

50 0.69

60 0.59

70 0.52

80 0.45

90 0.39

100 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,

29

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

30

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 multi-

pass 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.

mglymglymair

wall TTT

T ,,,

5.3

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.

glycol

toticoc

C

PTT ,, Eq. 4-19

From the assumption that Cglycol is 1.5 times larger than Cair; 5.0max

min C

C

the mean temperature of

the fluid can thus be approximated from the following equations (Shah & Sekulic, 2003):

2

,,,

ohihmh

TTT

Eq. 4-20

2

,,,

icocmc

TTT

Eq. 4-21

31

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.

ihic TC

qT ,

min,

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:

glycol

glycol

glycol

mQ

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 trade-

offs between the operating and capital cost of the piping system. Equation 4-24 is used to obtain

the pipe diameter. 5.0

4

m

gly

u

QD 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.

32

NPS Diameter

(in) (in)

External Internal

2 2.38 2.07

4 4.5 4.03

6 6.63 6.07

8 8.63 7.98

10 10.75 10.02

12 12.75 11.94

14 14 13.13

16 16 15

18 18 16.88

20 20 18.81

24 24 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.

2

4

i

md

Qu

Eq. 4-25

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:

2)(

2mglycol

ipip

uK

d

fLh

Eq. 4-26

33

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.

2,

,

4

bib

bmdn

Qu

Eq. 4-27

34

The total head loss across the manifolds is then found using Equation 4-28.

2

2033060

2,

,

bmglycoltttt

bibmani

uffff

d

fLnh

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.

)(

)(

power possible Maximum

Power Actual

,,min

,,

icih

ohihh

TTC

TTC

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.

accmanipipHEstotp hhhhh , Eq. 4-30

totpGlycolusef hQP ,

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.

pumpmot

usefact

PP

Eq. 4-32

35

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.

day

PS

nS usefkWh

dayspump

sec86400

sec3600 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 non-

operating 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.

ooairiiairfanuse PQPQP ,,, Eq. 4-34

The fan and motor efficiency are taken from the user inputs. The actual power is calculated from

Equation 4-35:

motfan

fanusefanact

PP

,, 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.

dayP

SnS fanact

kWhdaysfan

sec86400

sec3600, Eq. 4-35

To calculate the net energy cost balance per year the Equation 4-36 is used:

fanpumpsavbal SSSS Eq. 4-36

36

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.

tot

airpairin

P

cmT

,,

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.

ambTTT min Eq. 4-38

sec86400/,,

monthdaysairpairinsav nTCmE 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.

sec86400 daystotsav nPE Eq. 4-40

37

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.

savfuelsav ESS 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.

)( ,,, hglyambhpipininsloss TTAkP 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

38

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.

TAkP wpippvcloss , =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/m2 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, gainP =1.17 MW

To determine the glycol mixture temperature increase, Equation 4-43 is used.

39

gly

gain

C

PT 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.

fg

latcond

h

Pm

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:

2

54321

15.27315.27315.273

TA

TA

TAAAPx

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.

40

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.

41

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)

42

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.

43

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 Area (m2) mu Mean velocity across cross-sectional area (m s

-1)

C Flow stream heat capacity rate (kW °C-1

) sV Total system’s volume (m

3)

id Pipe outside diameter (m) tV Expansion tank’s volume (m

3)

od

Pipe inside diameter (m) W Width (m)

h Head loss (kPa) pipeW Weight of pipe

k Thermal conductivity (kW m-1

°C-1

) Greek symbols

K

Pressure loss coefficient, dimensionless Linear coefficient of thermal expansion (m/ m K)

gh Specific enthalpy of phase change (kJ kg-1

) Heat exchanger efficiency, dimensionless

L Length (m) fan Fan efficiency, dimensionless

m Mass flow rate (kg s-1

) (kg s-1

mot Fan electric motor efficiency, dimensionless

daysn Number of days within a given a month Density (kg m-3

)

bP Barometric pressure (kPa) Difference

sensP Power recovered from sensible heat (kW) Kinematic viscosity (m2 s

-1)

latP Power recovered from latent heat (kW) h Specific volume of glycol/water at higher temperature (m

3/kg)

Pr Prandtl number, dimensionless c Specific volume of glycol/water at lower temperature (m

3/kg)

Q Volumetric flow rate (m3 s

-1) Fraction of mixture

Re Reynolds number, dimensionless Subscript

S Cost ($CAN) c Cold fluid

t Pipe thickness (m) f Film

Temperature (°C) h Hot fluid

dbT Dry bulb temperature (°C) i Inlet

wbT Wet bulb temperature (°C) o Outlet

44

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.

17.30

QIntNcoils 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

45

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 0.6

Weld-o-let 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 35 1.6

5 ft (3 m) of 4” dia. 48.25 1.46

Ball valve 720 0.421

Flexible hose flanged 580 1.667

Total 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.

46

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:

coilscoilreq NAA ,1 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;

11tan

2

actreq AA

L Eq. 5-3

Figure 5-3: Surface building extension

47

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.

The actual size of the building are then found using the following equations:

'24

reqAW Eq. 5-4

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).

cos,

LL fwall Eq. 5-5

Where L

BW 2/)(arctan

Eq. 5-6

Where actAB Eq. 5-7

The area of the side walls is found using Equation 5-8.

2

)(cos

cos2

BHL

BL

Awalls

Eq. 5-8

The area of the roof is found from Equation 5-9.

Figure 5-4: Surface building extension building top and side view

48

)(

coscosBW

LB

LAroof

Eq. 5-9

The floor area is found from Equation 5-10.

)( BWLLBAfloor 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.

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.

22

3 11

2outin

cRAA

QP 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.

Figure 5-5: Exhaust ventilation collar at surface (Gagnon, 2011)

49

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.

WLL fwallsfds ,2 Eq. 5-12

50

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: WLA 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

51

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 non-

operating 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

52

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.

121

pippip

LIntn 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.

1

3

11

21elb

pipcps nInt

LIntn 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.

elbelbcpscpsgrpippippiplabpiplab HnHnHnHLSS ,2, Eq. 5-15

elbelbcpscpspippippipmat SnSnSLS , 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

53

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:

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.

NPS Depth

in ft

2 to 10 3

12 to 18 4

20 to 24 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.

L

ddepthV o

12

)"4(2037.0 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:

8

VHIntN exc

days +1 Eq. 5-18

54

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 Depth

in in

2 to 10 3

12 to 18 6

20 to 24 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.

depthWLV 037.0 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.

55

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:

)(91.15 tdtW opipe Eq. 5-20 (Engineering toolbox)

Then to calculate the weight of the glycol mixture per meters within the pipe; Equation 5-21 is

used;

4

2i

mixgly

dW

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;

pipglysp

WW

LoadMaxL

. Eq. 5-22

The total number of supports required is the total pipe length divided by the maximum support

span.

56

1sup

sp

tot

L

LIntn Eq. 5-23

The weight per length and maximum span for different NPS is shown in Table 5-5.

NPS Weight

Maximum Span

in kg/m m

2 7.8 220

4 25.1 68.7

6 48.8 35.4

8 78.0 22.1

10 116.2 14.9

12 159.2 10.8

14 189.8 12.7

16 241.5 9.96

18 315.1 9.65

20 380.3 7.99

24 539.9 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.

lab

f

t RHSS

nS sup1sup,

sup,

supsup,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.

57

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.

inselbelbinspiplabinslab HnHLSS ,, Eq. 5-25

inselbelbinspipinsmat SnSLS ,, 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:

58

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 head

gpm 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.

1

pump

sys

pumph

hIntN 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.

59

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.

60

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.

61

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) :

TVV

c

hst

31 Eq. 5-28

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:

st VV 00768.0 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:

Ld

V opip 15.12

4

2

Eq. 5-30

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.

425

2o

coilsmani

dNV

Eq. 5-31

62

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: 1562.10 tantan, VSmat Eq. 5-32

For labour time: 424.1012.0 tantan, VHlab 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.

63

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

galglytotglygly SVXS / 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.

64

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.

65

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.

66

The flat spray covers a length of ; 5.37tan3.02 = 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; 20tan8.12 =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.

67

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 three-

way 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.

150

noz

val

NIntN 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

68

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 4 (Fig. 5-19)Airflow: 400m3/s, distance: 200m



Case 7(Fig. 5-22) Airflow: 600m3/s, distance: 200m



69

Figure 5-16: Cost distribution of components case 1 Figure 5-17: Cost distribution of components case 2




70

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%

71

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 m

3/s & 600 m

3/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 4 (Fig. 5-28)Airflow: 400m3/s, distance: 200m, 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



Figure 5-25: Heat recovery system economics, case 1 Figure 5-26: Heat recovery system economics, case 2

BC Qc NWT BC Qc NWT BC Qc NWT

8 $/GJ 13 $/GJ 20 $/GJ

$0

$500,000

$1,000,000

$1,500,000

$2,000,000

$2,500,000

0

2

4

6

8

10

12

14

Cost data of return air heat recovery system (Q: 200m3/s, dist.: 200 m, capital cost: 1,196,000$)

Payback period (yrs) Gross energy savings ($/yr) Net energy savings ($/yr)

yrs


8 $/GJ 13 $/GJ 20 $/GJ

$0

$500,000

$1,000,000

$1,500,000

$2,000,000

$2,500,000

0

2

4

6

8

10

12

14



yrs

39 25 15 15

72




Figure 5-24: Heat recovery system economics, case 9


8 $/GJ 13 $/GJ 20 $/GJ

$0

$500,000

$1,000,000

$1,500,000

$2,000,000

$2,500,000

0

2

4

6

8

10

12

14



yrs


8 $/GJ 13 $/GJ 20 $/GJ

$0

$500,000

$1,000,000

$1,500,000

$2,000,000

$2,500,000

0

2

4

6

8

10

12

14



yrs


8 $/GJ 13 $/GJ 20 $/GJ

$0

$500,000

$1,000,000

$1,500,000

$2,000,000

$2,500,000

0

2

4

6

8

10

12

14



yrs


8 $/GJ 13 $/GJ 20 $/GJ

$0

$500,000

$1,000,000

$1,500,000

$2,000,000

$2,500,000

0

2

4

6

8

10

12

14



yrs


8 $/GJ 13 $/GJ 20 $/GJ

$0

$500,000

$1,000,000

$1,500,000

$2,000,000

$2,500,000

0

2

4

6

8

10

12

14



yrs


8 $/GJ 13 $/GJ 20 $/GJ

$0

$500,000

$1,000,000

$1,500,000

$2,000,000

$2,500,000

0

2

4

6

8

10

12

14



yrs


8 $/GJ 13 $/GJ 20 $/GJ

$0

$500,000

$1,000,000

$1,500,000

$2,000,000

$2,500,000

0

2

4

6

8

10

12

14



yrs

52 30 29 87 21 15

33 20 75

43 23 21

43 23 22 16

36 21

93

73

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 de-

watering 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.

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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

75

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

76

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

77

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

78

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.

*RFE Eq. 4-1

Where:

F : Factor of merit *R : Tower capacity factor

If R=<1, Ew

If R=>1, Ea

Where inainw

outwinw

wTT

TT

,,

,,

Water efficiency Eq. 4-2

79

inwina

outainaa

TT

TT

,,

,,

Air efficiency Eq. 4-3

inainw

inainwwp

a

w

w

a

SS

TTC

m

mR

,,

,,,

(McPherson, 1993) Eq. 4-5

Where:

inwS ,: Sigma heat of saturated air at same temperature as the inlet water (J/kg dry air)

inaS , : Sigma heat of inlet air at exhaust (J/kg dry air)

inwT , : Inlet water temperature

inaT , : Inlet air temperature

wpC , : Specific heat capacity rate of water at constant pressure

wm : Mass flow rate of water

am : Mass flow rate of air

)386.25.2502( wetw TL

wetwin TWLS 1005

Where:

wetT : 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.

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:

Figure 6-6: Plate heat exchanger (Made-in-China.com, 2010) (IQS inc.)

81

To determine the pressure drop across the PHE, from (Shah & Sekulic, 2003) Equation 6-6 is

used:

e

pp

D

fLGnGP

2

4

2

5.1 2

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:

2)4/( p

pD

mG

0A

mG

bwNA p 0

eGD

Re

Where

b : Distance between chevron plates

pD : Diameter of ports

eD : Hydraulic diameter between chevron plates

25.0Re8.0 f : friction factor

L : Length of chevron plates

m : Mass flow rate of fluid

pN : Number of fluid passes

pn : 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.

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 tube-

fin 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.

Direct contact HE at exhaust

83

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 TCm p

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

84

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

85

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.

86

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

87

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.

88

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

89

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.

90

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.

91

Nomenclature

fA Secondary surface area (fin) (m2)

uP Actual vapour pressure (kPa)

frA

HE frontal area (m2)

wP Saturated vapour pressure at wet bulb temperature (kPa)

idA

Cross-sectional inside tube area (m2) Pr

Prandlt number, dimensionless

oA Minimum free flow area (m2) Q Volumetric flow rate (m

3 s

-1)

pA Primary surface area (tube) (m2)

hr Hydraulic radius (m)

tA Total heat transfer surface area (m2) Re

Reynolds number, dimensionless

C

Heat capacity rate (kW K-1

) R Thermal resistance (K W-1

)

minC Min heat capacity rate between the two fluids (kW K

-1)

dt Tube thickness (m)

maxC Max heat capacity rate between the 2 fluids (kW K

-1)

T Temperature (°C)

pc Specific heat at const. pres (kJ kg-1

K-1

) dbT Dry bulb temperature (°C)

HD

Hydraulic diameter (m) wbT Wet bulb temperature (°C)

id Tube inside diameter (m) lmT

Log mean temperature difference

od Tube outside diameter (m) UA Overall thermal conductance (W °C

-1)

f Friction factor, dimensionless

mu Air mean velocity (m s-1

)

pF Fin pitch (m) V Volume occupied by HE (m3)

G

Mass flux (kg m-2

s-1

) W Humidity ration (kgdry vapour/kgdry air)

cg

Gravitational acceleration (9.8 m s-2

) lX Longitudinal tube pitch (m)

h

Heat transfer coefficient (W m-2

°C-1

) tX Transversal tube pitch (m)

j

j Colburn factor, dimensionless Z Altitude (m)

k

Thermal conductivity (W m-1

°C-1

) Greek symbol

cK Entrance contraction-loss coefficient, ( ) Difference

eK Exit loss coefficient, dimensionless

HE efficiency, dimensionless

1L Width of HE (m)

p Single pass HE efficiency, dimensionless

2L Length of HE (m)

Fin thickness (m)

3L Height of HE (m) f Fin efficiency, dimensionless

m

Mass flow rate of water (kg s-1

) o Overall fin efficiency, dimensionless

cN

Number of circuits Dynamic viscosity (Pa s)

fN Number of fins Density (kg m-3

)

RN Number of tube rows Ratio of total surface area on one side of the

heat exchanger to the total volume on both side of the heat exchanger, (m

-1)

RtN / Number of tubes per row Ratio of free-flow area to frontal area on one

side of the heat exchanger, dimensionless

tN Total number of tubes Subscript

NTU Number of Transfer Units c Cold fluid

Nu Nusselt number, dimensionless h Hot fluid

bP Barometric pressure (kPa) i Inlet

dP Saturated vapour pressure at dry bulb temperature (kPa)

m Mean

92

o Outlet

7.02 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

93

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: RRtt NNN / Eq. 7-1

Number of fluid passes: c

tp

N

NN Eq. 7-2

o Note: this number should return an integer; otherwise, some input variables

should be changed to obtain a realistic value.

Width: lR XNL 2 Eq. 7-3

Note: Xl/2 is added at both ends.

Height: 3L

o For inline arrangement: Rtt NXL /3 Eq. 7-4

o For staggered arrangement: 2

/3t

Rtt

XNXL Eq. 7-5

Note: Takes into account the Xt/2 at both ends.

94

Note: As mentioned previously, the user should ensure that it corresponds to the available

height of the heat exchanger casing.

Total number of fins: Lff NLN /1 Eq. 7-6

Tube outside diameter: dio tdd Eq. 7-7

Frontal area: 31LLAfr Eq. 7-8

Primary heat transfer surface area:

)4

(2)(2

3211 to

tfop Nd

LLNLNLdA

Eq. 7-10

Secondary heat transfer area (fin surface):

131

2

32 24

2 LNLLNNd

LLA ffto

f

Eq. 7-11

The total heat transfer surface: pft AAA Eq. 7-12

Minimum free flow area

o For Inline arrangement:

[( ) ( ) ]

Eq. 7-13

o For staggered arrangement

The following variables must first be calculated

( ) ( ) Eq. 7-14

[(

)

]

( ) Eq. 7-15

{

Eq. 7-16

To then find the minimum free flow area

[(

) ( ) ( ) ] Eq. 7-17

Volume occupied by the heat exchanger: Eq. 7-18

Hydraulic Diameter of the heat exchanger

The following variables must first be calculated:

Eq. 7-19

Eq. 7-20

To then find the hydraulic diameter

Eq. 7-21

95

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: fr

mA

Qu Eq. 7-22

Mass flow rate of dry air: Qm

Eq. 7-23

7.04 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:

mcC p Eq. 7-24

)( ,,min

,, icihh

ihoh TTC

CTT Eq. 7-25

96

)( ,,min

,, icihc

icoc TTC

CTT Eq. 7-26

The cp’s of the two fluids were approximated to be at a temperature in between the inlet hot and

cold fluid temperatures. Thus the cp of air is evaluated to be 1.005 Kkg

kJ

and that of water 4.18

Kkg

kJ

. The specific heats of water and air are relatively constant at the operating temperatures

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 5.0max

min C

C

2

,,,

ohihmh

TTT

Eq. 7-27

2

,,,

icocmc

TTT

Eq. 7-28

For 5.0max

min C

Cand Cmin is the hot fluid and Cmax cold fluid:

lmmcmh TTT ,, Eq. 7-29

)/()(ln

)()(

,,,,

,,,,

mcohmcih

mcohmcihlm

TTTT

TTTTT

Eq. 7-30

2

,,,

icocmc

TTT

Eq. 7-31

For 5.0max

min C

Cand Cmin is the cold fluid and Cmax is hot fluid:

lmmhmc TTT ,, Eq. 7-32

)/()(ln

)()(

,,,,

,,,,

icmhocmh

icmhocmhlm

TTTT

TTTTT

Eq. 7-33

2

,,,

ohihmh

TTT

Eq. 7-34

97

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)

omd

duoRe Eq. 7-35

HmD

DuHRe 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.

cdm

NA

Qu

i

Eq. 7-37

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

98

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. 362.0

28.0Re15.0

to

o

A

Aj Eq. 7-38

Where

211

41

fo

t

o

o

t

o

l

to

oNd

X

d

d

X

d

X

A

A 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). 212.0

0897.0

0449.0

392.0Re394.0

o

p

od

d

FN

dj

o

Eq. 7-40

Then using the following relation, the Nusselt number is found from Equation 7-37.

Re

Pr 3/1

Nu

j Eq. 7-41

Using the Nusselt number, the heat transfer coefficient is found from Equation 7-38.

H

h

D

kNuh

Eq. 7-42

99

From the heat transfer coefficient and the following equations, it is possible to calculate the fin

efficiency. 5.0

tl

e

XXr

2

oo

dr

1/2

*

*

2, ( ) , exp(0.13 1.3863),

, / ,2

nae e

f

e f e o f e o

hm m r n m

k

r r r r r

2for 0893.09107.0 ** rrb

2for n r17125.09706.0 ** rb e

445.0*)(2.2570.6for tanh

rf Eq. 7-43

445.0*-b )(2.2570.6for )( rma ef Eq. 7-44

From the fin efficiency it is then possible to calculate the overall efficiency from Equation 7-45:

)1(1 f

t

f

oA

A 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=0.023Re0.8

Pr0.4

Eq. 7-46

The heat transfer coefficient can be found using the following equation:

o

c

d

kNuh

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:

cc

hAR

)(

1 Eq. 7-48

The air side

hoh

hAR

)(

1

Eq. 7-49

The wall resistance

100

Lk

rrR

t

iow

2

)/ln( 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 1

,,,

11

chh

mh

c

mcwm

RRR

T

R

TT Eq. 7-51

From the resistance of the system, it is possible to determine the overall thermal conductance UA

of the system.

cwh RRRUA

1

Eq. 7-52

From the thermal conductance, the number of transfer units (NTU) can be found from the

following equation:

minC

UANTU 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

*/*)exp(1exp1 CCNTUp Eq. 7-54

101

For Cmin unmixed and Cmax mixed

)exp(1*exp1(*

1NTUC

Cp Eq. 7-55

For both fluids unmixed

1

1

1

!

)1(

)!1(

1)(

)(**)1(exp)exp(1

nn

j

n

n

n

j

p

yj

jn

nyP

NTUPCNTUCNTU

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.

CC

C

n

p

p

n

p

p

1

1

11

1

(Overall counter flow) Eq. 7-57

*

*

1

)))1((1(1

C

C np

(Overall parallel flow) 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

102

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.

103

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;

lossesexit and entrance

222

losses changedensity lossesfriction

2

'1'12

'12

1

2

o

iec

ico

i

m

ihic

KKg

G

r

Lf

g

Gp

Eq. 7-58

Where

edgeleadingmuG

' 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.

m

ihic r

Lf

g

Gp

1

2

2

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:

197.0

0935.0

104.0

418.0Re039.1

o

p

od

d

FN

d

tf

o 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.

104

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.

1

2

1

2

1

22 P

u

u

f

fP

m

m

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.

0

50

100

150

200

250

300

350

400

450

500

0 2 4 6 8 10 12 14

Pre

ssu

re d

rop

(P

a)

Air speed (m/s)

Face air speed vs pressure drop

n5

n4

n2

n3

n1

n6

105

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

)( 12 PPQW in

Eq. 8-1

Where

Q is the volumetric flow rate of air

The change in specific enthalpy will be:

m

Whh

in12 Eq. 8-2

Where

m is the mass flow rate of air

For an underground mine, the rate of work would become:

gZmW in

Eq. 8-3

Where Z is the change in elevation and g the constant gravitational acceleration (9.81 m/s2)

106

Qm Eq. 8-4

The increase in pressure for an underground mine is calculated as follows:

12 PPgZ 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.

19075103.101

Z

bP

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.

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:

100012

gZhh Eq. 8-7

0

10

20

30

40

50

60

70

80

90

0 1000 2000 3000 4000 5000 6000

ΔP

(kP

a)

Depth (m)

Change in pressure with negative Altitude

Exponential

Linear

107

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.

1

2

1

2

1

n

n

T

T

P

P 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).

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

0

10

20

30

40

50

60

0 1000 2000 3000 4000 5000 6000

Dry

bu

lb t

em

pe

ratu

re c

han

ge (

°C)

Depth (m)

Temperature increase with depth for intake temperature of 20°C

Polytropic

Linear enthalpy

108

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.

)()( 12lg1212 WWhTTCmhh p

Eq. 8-9

Where

lgh : Heat of evaporation

W : Humidity ratio

As the humidity content of air increases ( 12 WW ), the change in dry bulb temperature ( 12 TT )

decreases since the change in enthalpy ( 12 hh ) 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.

109

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 s

2 m

-8

Flow rate of air: 50 m3/s

2RQP 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):

C

Cm

T

p

9.1125

Eq. 8-11

110

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.

111

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.

112

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 vapour-

compression 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

113

cycle, the totality of the work from the compressor is rejected into heat at the condenser (line 2-

3-4). 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, )1ln( NTU (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.

)( 513

hh

Pm cool

NH

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.

)( 123hhmW NH

Eq. 9-2

The amount of power rejected from the condenser is calculated with Equation 9-3.

Wqq evapcond Eq. 9-3

Finally point 4 to 5 can be determined to size the expansion valve.

114

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.

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

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)

115

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

116

performance COP. It is defined as the total cooling capacity of the evaporator divided by the

compressor work as shown in Equation 9-4.

12

51

hh

hhCOP

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,

117

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, high-

pressure 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

118

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.

119

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

120

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.

121

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.

)0(, ambairpairair TCmq

Eq. 9-5

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.

wfginwwpww mhTCmq )0( ,, Eq. 9-6

Using Equation 9-5 and 9-6, the total mass of ice formed is found from Equation 9-7.

fginwwp

airw

hTC

qm

)0( ,,

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

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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.

123

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.

124

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.

125

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

126

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.

127

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I

APPENDIX A

Intake coil specifications sheet

II

Exhaust coil specifications

III

Specifications of Adjustable Saddle Support, Anvil International©.

IV

Foamglas® Insulation specification sheet

V

Plate heat exchanger quote, Thermofin©

VI


VII


VIII

Cities factor Newfoundland Peterborough 117.1 Saint-Hyacinthe 113.2

Alberta Corner Brook 118.7 Sarnia 116.8 Sherbrooke 113.8

Calgary 129.4 St-Johns 119.2 Sault Ste Marie 111.7 Sorel 114

Edmonton 130.6 Northwest territories St. Catharines 111.4 St Jerome 113.4

Fort McMurray 126 Yellowknife 120.5 Sudbury 111.4 Trois-rivieres 114

Lethbridge 119.9 Nova Scotia Thunder Bay 112.9 Saskatchewan

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 112 Regina 114.2

British Columbia New Glasgow 114.2 PEI Saskatoon 112.7

Kamloops 116.4 Sydney 111.6 Charlottetown 116.6 Yukon

Prince George 117.6 Truro 114.2 Summerside 116.1 Whitehorse 112

Vancouver 128.7 Yarmouth 114.1 Quebec

Victoria 117.7 Ontario Cap-de-la-Madelaine 113.8

Manitoba Barrie 118 Charlesbourg 113.8

Brandon 115 Brantford 117.1 Chicoutimi 112.8

Portage la Prairie 115 Cornwall 116.9 Gatineau 113.4

Winnipeg 127.7 Hamilton 121.9 Granby 113.7

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 117 Quebec 119.4

Newcastle 113.2 Ottawa 121.5 Rimouski 113.3

St. John 115.9 Owen Sound 118.1 Rouyn-Noranda 113.4 Canadian location factors for capital cost approximation (RSMeans Co., 2010)

Average soil temperature (°C)

Depth 1981 1984 1978

cm January February January February January February Total average

5 -1.3 -0.75 -1.3 -1.3 -3.3 -2.8 -1.8

10 -0.8 -0.4 -0.3 -0.3 -2.8 -2.8 -1.23

20 -0.45 -0.2 -0.25 -0.2 -2.2 -1.7 -0.83

50 0.3 0 0.5 0.0 -0.6 -1.1 -0.15

100 1.5 0.8 2.1 1.2 1.1 0.6 1.22

150 2.5 1.6 3.0 2.2 2.2 1.7 2.2

300 5.2 4.1 5.9 4.8 5.6 5.0 5.1

Soil Temperature Data, Canada, for Val d’Or (Centre climatologique Canadien, 1977-1984)

IX

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 ft $/ft2 floor hrs/ft2 floor

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 Labour and equipment

$/ft2 hrs/ft2

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 labour

$/ft2 hrs/ft2

0.38 0.007

Table 4 Insulation, 1.5” thickness, R5, Vynil/scrim/foil, ref: R.S. Means bldg construction 2008, page 400

X

Main piping system

Nature of soil Labour required to remove 1 B.C.Y.

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 pipe pipe 1 groove 2 grooves couplings couplings elbow elbow

in $/ft hours/ft hrs hrs $ hrs $ hrs

2 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

XI

NPS unit

price Labour

in. $ hours

2 12 0.1328

4 21.2 0.1392

6 31.28 0.1424

8 47.2 0.1456

10 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

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

NPS Ins.

Foamglas® Insulation

Elbow Ins.

Foamglas®

Elbow

Ins.

in $/ft hrs/ft $ 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

XII

Pump and MCC cost

Flow Head Motor required Pump Price

m3/s ft hp $

0.06309

(1000 gpm)

50 25 11500

100 50 11500

200 100 11500

300 150 11500

0.12618 (2000 gpm)

50 50 11600

100 100 11600

200 200 13700

300 300 18800

0.31545

(5000 gpm)

50 50 13700

100 100 13700

150 150 21400

0.6309

(10000 gpm)

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

Rotational

speed unit price

hp rpm $

25 1200 1894

3600 1150

50 1200 4519

3600 2911

100 1200 7952

3600 5481

150 1200 10826

3600 8926

200 1200 13631

3600 11904

300 1200 22517

3600 17946

400 1200 23566

3600 22465

500 3600 30969

Table 12 Electric motor cost, ref: (CostMine, 2004)

XIII

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.

XIV

Piping accessories

NPS unit price Labour

in. $ 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: 194921887 xy ,

For labour time: 014.0115.0 xy


in. $ 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

XV


in. $ 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


in. $ 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

Material cost: 6003.226 xy

Labour hours: 093.029.0 xy

XVI


in. $ 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


in. $ 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


in. $ 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;

XVII


in. $ 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 Labour

gal $ 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. unit price Labour

in. $ 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

XVIII

Automated washing system cost

Soil nature equipment Labour

$/ft 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 Labour

$/ft 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 Labour

$/ft hrs

2.35 0.0947

Table 30 Motor feeder systems, 115 V, 1 hp, ref: (CostMine, 2004) RSmeans electrical 2010 page 372

XIX

APPENDIX C

-For air

μ: Dynamic viscosity: From 0 to 25°C

μ = -2.3413*(E-13)*T3 - 2.0916*(E-11)*T

2 + 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)*T

3 + 1.1458*(E-06)*T

2 - 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)*T

3 + 1.0723*(E-06)*T

2 - 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)*T

3 + 1.1013*(E-02)*T

2 - 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*T

5-2.162*10

-6*T

4+ 5.543*10

-5*T

3-6.233*10

4*T

2 +7.7938*10

-

4*T+4.21

ρ: density of water: from 0 to 80°C

ρ =-4.684*10-8

*T4+3.456*10

-5*T

3-7.575*10

-3*T

2+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

XX

Design no.1 (Laronde design case)

HE geometry inputs

Xl m 0.0413

Xt m 0.0381

L1 m 1.9

fin thickness δ m 0.0003

Number of fins per m m 305

tube inside diameter m 0.01466

tube wall thickness unit/m 0.00061

number of tubes per row(1st row) m 46

number of tube rows m 8

no. of circuits 92

Inlet air Inputs

Vol. flow rate (across frontal area) m3/s 32.1

Dry bulb temperature °C 18

Wet bulb Temperature °C 18

Altitude m 335

Inlet running fluid Inputs

Inlet temperature °C 1.5

total mass flow rate kg/s 14.7

Design no. 2

HE geometry inputs

Xl m 0.05

Xt m 0.05

L1 m 2.5







no. of circuits 35

Inlet air Inputs

Vol. flow rate (across frontal area) m3/s 22



Altitude m 0



total mass flow rate kg/s 8

XXI

Design no. 3

HE geometry inputs

Xl m 0.05

Xt m 0.04

L1 m 5.6







no. of circuits 35

Inlet air Inputs




Altitude m 0




Design no.4

HE geometry inputs

Xl m 0.03

Xt m 0.04

L1 m 7.62







no. of circuits 36

Inlet air Inputs




Altitude m 335




XXII

Design no. 5

HE geometry inputs

Xl m 0.06

Xt m 0.04

L1 m 1.5







no. of circuits 36

Inlet air Inputs




Altitude m 0




Design no. 6

HE geometry inputs

Xl m 0.0413

Xt m 0.0381

L1 m 1.9







no. of circuits 92

Inlet air Inputs

Vol. flow rate (across frontal area) m3/s 32.1



Altitude m 335



total mass flow rate kg/s 14.7

Documents

HEAT RECOVERY SYSTEMS IN UNDERGROUND MINE VENTILATION