Karlstads universitet 651 88 Karlstad

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Fakulteten för hälsa, natur- och teknikvetenskap Miljö- och energisystem

Kim Holmlund

Heat recovery solutions for mine

ventilation systems

Värmeåtervinningslösningar för gruvventilationssystem

Examensarbete 22,5 hp Högskoleingenjörsprogrammet i energi- och miljöteknik

Juni 2015

Handledare: Magnus Ståhl

Examinator: Lena Stawreberg

Abstract Recently, Boliden Mineral AB acquired the Finnish copper mine Kylylahti. In

connection with that, potential improvements, for instance in the ventilation system, is

investigated. The supply air going into the mine must be heated during the colder

months of the year to prevent icing that can damage the supply air shaft and adjacent

structures. Currently, the heating is done with LPG, but there is a lot of available energy

to possibly make use of since the return air is saturated. A heat exchanger could

possibly decrease the operating cost, the carbon dioxide emissions and the noise that

spreads over the area and disturbs the neighbours. Recently, the mine Zinkgruvan made

a very successful investment in a plate heat exchanger, and a Canadian study shows that

also battery heat exchangers can be a good alternative. In this work, both kinds of heat

exchangers were evaluated. Calculations were made for the total number of hours

spanning between February 2014 and January 2015, which were later normal year

corrected. Tabulated values and given data, including airflows and ambient air

temperature, were used to calculate enthalpy in the airflows. Available energy, energy

demand and transferable energy with heat exchanging was calculated through changes

in enthalpy. To calculate the payback period, the investment cost for the two systems

were estimated and the change in operating cost were calculated. The change in carbon

dioxide emissions was also calculated. All calculations were made for eight different

cases, where factors such as required supply air temperature, return air temperature and

LPG-price was varied within reasonable ranges. The investment cost for the battery heat

exchanger system is approximately 500 000-700 000 euros and for the plate heat

exchanger system approximately 1-1,1 million euros. Both systems reduce the LPG-

demand, but the electricity demand is increased since the fans have to overcome larger

pressure drops. Both systems have a lower operating cost than the current system in all

tested cases, and in most cases the plate heat exchanger system has the lowest. For the

plate heat exchanger system, the shortest possible payback period is 3,2 years, and the

longest possible is over 72 years. With the current values the payback period is 12,5

years. For the battery heat exchanger system, the shortest possible payback period is 1,9

years, the longest possible is 9,4 years and 6,3 years with current values. But in the

future, the LPG-price will probably increase and the payback period then becomes 6,4

years for the plate heat exchanger system and 3,4 years for the battery heat exchanger

system. The plate heat exchanger is more efficient than the battery heat exchanger, but it

does not have the same short payback period since the investment cost for a plate heat

exchanger system is significantly higher. Compared to the current heating method, both

the evaluated systems reduce the carbon dioxide emissions. With a margin electricity

perspective, the reduction is low, only 10-13% of the current emissions. With the

method Boliden uses for similar assessments the reduction is greater and the plate heat

exchanger system gives the greatest reduction, around 60% in comparison to 55% for

the battery heat exchanger system. With other possible improvements, the conditions for

a heat exchanger system may change, very likely so that the payback period becomes

longer. The payback period must be compared to the expected life of the mine, which

currently is around 6 years. But before a decision can be made, an estimation of how

much a heat exchanger can reduce the noise level must be made.

Sammanfattning Nyligen köpte Boliden Mineral AB den finska koppargruvan Kylylahti och i samband

med det undersöks potentiella förbättringsmöjligheter, bland annat i

ventilationssystemet. Tilluften som går ned i gruvan måste under vinterhalvåret värmas

upp för att förhindra isbildning som kan skada tilluftsschaktet och angränsande

strukturer. För nuvarande görs uppvärmningen med gasol men det finns mycket energi

att ta tillvara på eftersom frånluften är mättad. En värmeväxlare skulle förhoppningsvis

kunna sänka driftkostnaden, koldioxidutsläppen och även minska bullret som i nuläget

sprids över området och stör boende. Nyligen gjorde gruvan Zinkgruvan en mycket

lyckad investering i en plattvärmeväxlare och en kanadensisk studie som gjorts visar att

även batterivärmeväxlare kan vara ett bra alternativ. I detta arbete gjordes beräkningar

på båda. Beräkningar har gjorts för alla timmar under perioden februari 2014 till januari

2015, vilket sedan normalårskorrigerades. Det användes tabellvärden och givna data för

bland annat luftflöden och utomhustemperatur för att beräkna entalpier i luftflödena.

Med hjälp av entalpierna har sedan tillgänglig energi, energibehov och vad som kan

överföras med värmeväxlare beräknats. Investeringskostnad för de två systemen

uppskattades och ändring i driftkostnad beräknades för att slutligen kunna beräkna

återbetalningstid. Även ändringen i koldioxidutsläpp har beräknats. Beräkningarna har

gjorts för åtta olika fall, där faktorer som bland annat uppvärmningstemperatur,

frånluftstempertur och gasolpris varierats inom rimliga intervall. Investeringskostnaden

för batterivärmeväxlarsystemet är ungefär 500 000-700 000 € och för

plattvärmeväxlarsystemet ungefär 1 000 000-1 100 000 €. Båda värmeväxlarsystemen

minskar behovet av gasol, men istället ökar elbehovet eftersom fläktarna måste jobba

mot ett högre tryckfall. Båda systemen har i alla testade fall lägre driftkostnad än

nuvarande system. I de flesta fall har plattvärmeväxlarsystemet lägre driftkostnad än

batterivärmeväxlarsystemet. För plattvärmeväxlaren är den kortast möjliga

återbetalningstiden 3,2 år och i värsta fall är den över 72 år. Med de värden som gäller

just nu är återbetalningstiden 12,5 år. För batterivärmeväxlaren är den kortast möjliga

återbetalningstiden 1,9 år, i värsta fall är den 9,4 år och med de värden som gäller just

nu 6,3 år. Men i framtiden väntas gasolpriset stiga och återbetalningstiden blir då istället

6,4 år för plattvärmeväxlarsystemet och 3,4 år för batterivärmeväxlarsystemet.

Plattvärmeväxlaren är effektivare än batterivärmeväxlaren, men får aldrig lika kort

återbetalningstid eftersom ett plattvärmeväxlarsystem är mycket dyrare att investera i.

Båda lösningarna sänker koldioxidutsläppen. Med ett marginalelsperspektiv är

sänkningen ganska låg, bara 10-13% av de nuvarande utsläppen. Med den metod

Boliden använder för liknande bedömningar är sänkningen större och

plattvärmeväxlaren ger störst sänkning, cirka 60%, jämfört med 55% för

batterivärmeväxlaren. Med andra möjliga förbättringar kan förutsättningarna för ett

värmeväxlarsystem förändras, mycket troligt så att återbetalningstiden blir längre.

Återbetalningstiden måste jämföras med gruvans förväntade livslängd, som för

nuvarande är cirka 6 år. Men innan beslut måste även uppskattning av hur en

värmeväxlarlösning kan sänka bullernivån göras vilket ligger utanför detta arbete.

Preface

This work has been done in cooperation with Boliden Mineral AB during the spring of

2015 as part of the authors bachelor of science in environmental and energy engineering

at Karlstad University.

A big thank you to the supervisors Magnus Ståhl at Karlstad University and Andreas

Markström at Boliden for all the time and support, and also a big thank you to Eero

Tommila at Boliden Kylylahti for all the time and providing of necessary information.

Thank you also to Karl-Erik Rånman at Boliden, Claes Arvidsson and Lars Roar

Thoresen at Gupex, Percy Danielsson at Luvata and Mats Walther at Zinkgruvan for all

the information and help.

This thesis has been presented orally to an audience familiar to the subject. Thereafter

the work has been discussed at a special seminar. The author of this work has actively

participated in the seminar as an opponent of another thesis.

May 2015

Karlstad

Kim Holmlund

Nomenclature a Provisioning excess [€]

Specific heat capacity [J/kg,°C]

D Diameter [m]

d Diameter [mm]

f Friction factor [-]

G Investment cost [€]

g Gravitation constant [m/s2]

H Total head [m]

h Enthalpy [J/kg]

Minor loss coefficient [-]

L Length [m]

Mass flow [kg/s]

n Rpm [-]

P Pressure [Pa]

p Interest rate [%]

Power [W]

Re Reynolds number [-]

T Temperature [°C]

t Payback-time [years]

u Elevation [m]

Volume flow [m3/s]

v Fluid velocity [m/s]

Density [kg/m3]

Roughness [mm]

Efficiency [%]

Teperature transfer efficiency [%]

Dynamic viscosity [kg/m,s]

Relative humidity [%]

Specific humidity [kgvapor/kgdry air]

Subscripts Index Förklaring

v Vapor

g Water

a Dry air

Loss, tot Total loss

heat exc Heat exchanger

norm Normal month

Vocabulary Ambient air = The untreated outdoor air

Exhaust air = The air that passed the heat exchanger and is expelled into the atmosphere

LPG= Liquefied petroleum gas

Return air = The air that is taken up from the mine. Without a heat exchanger it is

expelled into the atmosphere

Supply air = The air that has passed the heating system and is ventilated down to the

mine

Table of Contents

1. Introduction ............................................................................................................... 1

1.1 Purpose .................................................................................................................... 2

1.2 Aim ......................................................................................................................... 2

1.3 Limitations .............................................................................................................. 2

2. Theory ....................................................................................................................... 3

2.1 Current system in Kylylahti .................................................................................... 3

2.2 Ventilation and air conditioning in mines ............................................................... 5

2.3 Heat exchanging ..................................................................................................... 5

2.4 A study of the economy in battery heat exchanging in mine ventilation................ 7

2.5 Zinkgruvan – an example of plate heat exchanging in mine ventilation ................ 9

3. Method .................................................................................................................... 12

3.1 Energy demand ..................................................................................................... 12

3.1.1 Volume- and mass flows of air ...................................................................... 14

3.1.2 Normal correction with degree days .............................................................. 15

3.2 Available energy ................................................................................................... 17

3.3 Energy available through heat exchanging ........................................................... 18

3.4 Design ................................................................................................................... 19

3.5 Investment cost ..................................................................................................... 21

3.6 Operating cost ....................................................................................................... 22

3.6.1 LPG ................................................................................................................ 22

3.6.2 Fan operation .................................................................................................. 22

3.6.3 Pump operation .............................................................................................. 24

3.7 Payback period ...................................................................................................... 26

3.8 Carbon emissions .................................................................................................. 26

4. Results ..................................................................................................................... 28

4.1 Energy demand ..................................................................................................... 28

4.2 Available energy ................................................................................................... 29

4.3 Costs ...................................................................................................................... 30

4.4 Payback period ...................................................................................................... 32

4.5 Carbon dioxide emissions ..................................................................................... 33

4.6 Detailed results ..................................................................................................... 35

4.7 Continued sensitivity analysis .............................................................................. 36

5. Discussion ............................................................................................................... 39

6. Conclusion ............................................................................................................... 44

7. Recommendation ..................................................................................................... 45

8. Further work ............................................................................................................ 46

9. References ............................................................................................................... 47

Appendix

Fan characteristics……………………………………………………………………...A

Case variables…………………………………………………………………………..B

Fan speed – logged data ……………………………………………………..…………C

Product sheets for the battery heat exchanger……………………………….……...….D

Pump characteristics……………………………………………………………………E

1

1. Introduction

In October 2014 the Swedish mining and metal company Boliden Mineral AB bought

the Finnish mine Kylylahti. The mine was opened in 2012 and is located in the historic

mining area of Outokumpu near the town Polvijärvi, in eastern Finland. (Boliden

Mineral AB n.d. a) Boliden is a company operating in all the processes of the chain of

metal production - prospecting/exploration, mining, enrichment/concentrating, smelting

and recycling of metals (Boliden Mineral AB n.d. b). Boliden operates five mining

areas, three in Sweden, one in Ireland (Boliden Mineral AB n.d. c), and Kylylahti in

Finland which is the latest acquisition. Boliden’s main metals for business are zinc and

copper, but lead, gold and silver are also of importance (Boliden Mineral AB n.d. b).

Kylylahti, whose main metal is copper, is one of Boliden’s smaller mines (Boliden

Mineral AB n.d. a).

Kylylahti is an underground mine, with mining in progress at a depth down to 700 m

below the surface (Boliden Mineral AB n.d. a). In all underground mining, proper

ventilation is essential to ensure the right working conditions for both equipment and

workers. In Nordic climates heating of the ventilation air is needed during the colder

season, foremost to protect shafts and other equipment from freezing. Ice in the supply

airshaft can block the airflow and loose icicles can cause damage to essential adjacent

structures. (Hartman et al. 1997)

In Kylylahti the supply air is estimated to require heating to approximately 3 °C to

avoid icing. The present system consists of 2 LPG-burners. The heating system and

supply airshaft and the return airshaft are situated only 30 m apart which enable a

different heating solution with heat exchanging between the two airflows. Since the

possibility to route the process water to the surface through an airshaft is currently being

evaluated, another possible solution would be to heat the supply air with process water.

In the current system there is also a problem with noise from the return air fan affecting

the neighbors.

In 2013 another Swedish mining company, Zinkgruvan, rebuilt one of their systems for

heating of supply air, from only oil-burners to a plate heat exchanger system, using the

heat in the return air, with oil-burners as complement. The project was very successful,

reducing the oil-consumption with 97%, with a payback period of less than 2 years.

Heating of mines is generally done with fossil fuel (Hartman et al. 1997). Heating

methods used in Boliden’s underground mines in addition to LPG are battery heat

exchangers and oil burners (Markström 2015). One of Boliden’s own environmental

goals for the period 2009-2013 was to keep the increase of CO2-emission down to 3%.

The goal failed and the increase reached 4%. For 2018 there is a goal to keep the

amount of CO2 per tonne metal at the 2012-level of 0,77 tonne CO2/tonne metal. In

2013 the amount was estimated to have reached 0,78 tonne CO2/tonne metal. (Boliden

Mineral AB 2014).

Metal production is an energy-intensive business. Boliden’s energy demand for 2013

was 19206 TJ and the cost for energy constitutes 18% of the total costs of the company

(Boliden Mineral AB 2014). Ventilation is not the largest consumer of energy in the

2

whole chain of metal production but it is estimated by ABB (cited by Walker 2013) to

cause about 30% of a mine's total energy demand.

1.1 Purpose

In connection with the acquisition of the mine in Kylylahti, Boliden wants to investigate

whether the supply air heating can be made cheaper with heat recovery system

complemented with LPG, than with the existing heating system with only LPG. A

profitable solution with a heat recovery system would be in line with Boliden’s

guidelines for increased efficiency, reduced/unchanged environmental impact/CO2-

emissions and social responsibility. The purpose of this work is to examine the

feasibility of building a heat recovery plant that could reduce both heating costs, carbon

emissions and reduce the noise from the return air fan.

1.2 Aim

By analyzing the energy content in the air, the energy demand for heating and efficiency

of the heat exchanger, two heat recovery systems between the supply air and return air

will be evaluated. To calculate a payback period, additional operating costs will be

calculated and cost of the investment will be estimated. In addition, the change in CO2-

emissions compared to the current system will be calculated.

1.3 Limitations

The work is limited to include only battery heat exchangers and plate heat exchangers

since they are well tested in mine ventilation.

3

2. Theory

2.1 Current system in Kylylahti

The mine Kylylahti is located in the village with the same name, situated outside of the

town of Polvijärvi in eastern Finland. The mine was opened in 2012 and was acquired

by the Swedish mining company Boliden Mineral AB in October 2014. Currently, the

mine has known ore corresponding to a life expectancy of around 6 years at current

production rate (Rånman 2015).

The ventilation station in its current form generates a lot of noise, which reaches and

disturbs the neighbours living in Kylylahti and Polvijärvi. Since the station will have to

undergo some form of a rebuild to fix the noise problem, there is also an opportunity to

improve the heating system in the process. An overview of the ventilation station is

shown in figure 1.

All the air to the Kylylahti mine is heated at the same ventilation station. The main fans

and the heating system are positioned above ground. The purposes of the main fans are

to convey the air between ground level and underground. Further down in the

ventilation system there are booster fans to overcome the higher pressure in the smaller

ducts to ensure that the air is distributed in the drifts. Water purification ponds are

located approximately 550 m from the ventilation station.

Figure 1. Overview of the ventilation station in Kylylahti. On the left side is the burner house,

leading supply air to the supply air fan and through a bend before it is sent underground. On the

right side is the return air fan and evasé. (Photo: Boliden Kylylahti)

The supply airshaft is 4 m in diameter and the return airshaft is 3,5 m, and the distance

between them is approximately 30 m. The current volume flow of supply air is 120 m3/s

and of return air 90 m3/s. Two-three times a day the flows are temporarily increased for

two hours to dilute blasting fumes. During those hours the volume flows are 180 m3/s

4

and 135 m3/s. (Tommila 2015) The difference in supply and return airflows ensures

airflow from underground up through the decline to keep it ventilated and ice-free.

Both fans are axial fans. The installed power is 900 kW for the supply air fan and 800

kW for the return air fan, both with maximum speed of 1000 rpm, see appendix A.

During normal operation, the differential pressure has been calculated in this work to

approximately 609 Pa over the supply air fan and 724 Pa over the return air fan, during

diluting operation the pressures are approximately 1371 Pa and 1629 Pa respectively.

Because of the high moisture content and icing on the measuring equipment, the

pressure and volume airflow measurements are unreliable, but fan speed is logged

continuously (Tommila 2015).

The average annual temperature in Kylylahti is 2-3 °C (Finnish meteorological institute

n.d.). The regulation of supply air temperature is currently done manually. When the

temperature drops very low, the air is heated to approximately 9 °C to avoid freezing.

(Tommila 2015) In the mining district of Outokumpu the temperature gradient is about

1,3 °C/100 m, giving the rock a natural temperature of approximately 12 °C at 500 m

depth (Kukkonen et al. 2011). At the highest point of measure, 142 m below ground

level, the return air temperature is 8,5 °C and the relative humidity 92%. At ground

level, the relative humidity is estimated to be 100%. (Tommila 2015)

With the current system the heating is done by two automated, direct working LPG

burners, each with a maximum power of 3,2 MW. When burner 1 is operating at more

than 70% of maximum capacity, burner 2 starts and both operates at the same level.

Burner 2 shuts down when the operating level drops to less than 42%. (Tommila 2015)

The burners are normally operating between October and April. In the period February

2014 to January 2015, approximately 415 tonne of LPG was used to heat the supply air.

A principle sketch of the ventilation heating system is shown in figure 2.

Figure 2. The principle of the current heating system in Kylylahti

5

2.2 Ventilation and air conditioning in mines

The air in underground mines is naturally stagnant and can be contaminated. To ensure

standards for safety and suitable working conditions for both workers and machines,

ventilation is necessary. Fresh air is needed both for workers, machines and vehicles

(running on diesel) and the extraction of air removes exhaust gases from machines,

vehicles and blasting fumes. (Hartman et al. 1997)

In colder climates it is usually necessary to heat the supply air during the colder season.

The primary purpose of heating the supply air is to protect the shaft and equipment.

Ground water and water conducting fractures in the rock keeps the walls in the supply

airshaft wet. To keep the water from freezing, the supply airflow needs to be slightly

above the freezing point so the water does not freeze from the evaporative cooling

effect. Build-up of ice obstructs the airflow and increases the resistance considerably.

There is also the risk of ice falling and damaging adjacent equipment. In many cases a

supply air temperature of 1 °C is enough to avoid icing (Hartman et al. 1997), but in

practice it is heated further to keep a safety margin, often to 3-4 °C (Brake 2013).

The air going down to the mine gains heat in several ways. Going down the air is heated

naturally through autocompression, due to conversion of potential energy to thermal

energy. This heat is lost when the air goes up again, due to decompression, which

means it can’t be used in a heat exchange system for heating the supply air. Going

down, the air also absorbs heat from the rock. 15 meters below the surface, the rock

temperature is considered constant and independent of the surface temperature. From

there the temperature increase with depth with an approximately uniform rate called the

temperature gradient. But in a similar way the air loses heat on the way up due to the

decrease in rock temperature with height. While down in the mine there are many ways

for the air to gain heat, including machinery, lights and blasting. But their effect is

minor compared to autocompression and the geothermal energy. When the return air

reaches the surface it is warmer than when going down mainly due to the absorption of

geothermal energy while down in the mine. The return air is also generally saturated.

The high humidity is due to a combination of the humid environment and water from

the combustion in diesel engines. As the air rises, it cools down, increasing the relative

humidity to 100%. (Hartman et al. 1997)

2.3 Heat exchanging

Three different kinds of heat exchangers are suitable for ventilation purposes when

heating supply air with return air: rotating heat exchangers, battery heat exchangers and

plate heat exchangers (Warfvinge & Dahlblom 2010). In mine ventilation, both plate

heat exchangers and battery heat exchangers are well tested. Rotating heat exchangers

was tested in the old Boliden mine Långdal between the years 1985-1990. There was a

problem with under dimensioned bearings but the project managed to decrease the oil

consumption to 31%, and the payback period was estimated to be about 9 years. But the

underground operation was ended in 1992 and no further tests have been done since.

(Markström 2015)

6

A battery heat exchanger, also called closed-loop glycol circuit, consists of two sets of

tube and fin heat exchangers. Every heat exchanger consists of several rows of the fluid-

carrying tubes connected by rows of fins. The principle of a battery heat exchanger is

shown in figure 3.

Figure 3. The principle of a battery heat exchanger

In a battery heat exchanger, the warm flow and the cold flow can be kept separated,

which prevents any leakage between them. A circulation pump keeps a fluid, usually

water with antifreeze, flowing between the two airflows, transporting the heat. For

conventional ventilation systems the temperature transfer efficiency is relatively low

(50%) and the pressure drop is relatively high (200 Pa). The efficiency can be increased

with more rows, but that also increases the pressure drop. To defrost a battery heat

exchanger, a part of the fluid flow can be by-passed the supply air battery or the fluid

can be stopped all together. (Warfvinge & Dahlblom 2010)

Plate heat exchangers also go by the name cross-flow heat exchangers. In a plate heat

exchanger, the warm flow and the cold flow meet cross stream on alternate sides of a set

of thin plates, meaning that the return airflow and the supply airflow need to be routed

together. The principle of a plate heat exchanger is shown in figure 4. Plate heat

exchangers are a fairly easy construction with no moving parts and minimal leakage.

For conventional ventilation systems the pressure drop is relatively high (150 Pa) and

the temperature transfer efficiency is relatively low (50-60%). (Warfvinge & Dahlblom

2010) Due to its construction, the plate heat exchanger always has a “cold corner”

where the yet unheated air and the chilled exhaust air meet, leading to a bigger risk of

freezing. Methods for defrosting a plate heat exchanger is to by-pass (parts of) the

ambient air or to partially and sequentially block it. (Fläkt Woods n.d.)

7

Figure 4. The principle of a plate heat exchanger

Temperature transfer efficiency is a measure of air heat exchangers heat exchange

ability - a relation between the real change in temperature and the maximum available

change. It is a constant dependent on factors including the heat exchangers size and the

heat transfer coefficients. (Warfvinge & Dahlblom 2010) According to industry

standard the temperature transfer efficiency shall be provided for specific conditions,

which are an ambient air temperature of 5 °C and a return air temperature of 25 °C and

relative humidity of 27% (Svensk ventilation 2012). The temperature transfer efficiency

is calculated with (1), and the equation can conversely be used to calculate what

temperature the flow will have after heating with a given heat exchanger (Warfvinge &

Dahlblom 2010).

Since the return air from mines usually is saturated, the temperature transfer efficiency

can be misleading because it does not take the higher humidity and condensation on the

return side into consideration. For that purpose a wet temperature transfer efficiency

provided by the supplier can be used to accommodate for these special conditions.

2.4 A study of the economy in battery heat exchanging in mine ventilation

A study by Dello Sbarba et al. (2012) was conducted in Canada, calculating energy

savings and payback period for battery heat exchanger systems in mine ventilation.

They used a computer software application in Microsoft ExcelTM

which was made

specifically to calculate feasibility in heat recovery in mining ventilation.

The software has two main components. The first is an energy analysis using common

psychometric and thermodynamic equations and efficiencies of the heat exchangers.

8

The other component is estimating the investment cost depending on several

parameters.

Some parameters were constant during the calculations. For instance, the return air was

assumed to be 13 °C and of 100% relative humidity. The required supply air

temperature was chosen to 1,5 °C and the electricity price was assumed to be 0,08

$/kWh.

Variable parameters were amongst others: fuel price, airflow rates and the distance

between the shafts. Supply and return airflows were assumed to be of the same size, but

different rates were tested. Three different Canadian regions were included, using

monthly average temperatures in the calculations.

Investment cost included heat exchangers and the installation, piping material and

labour, pump and automated washing system. It was pointed out that the cost of the heat

exchangers usually is most significant.

An increased pressure drop of 250 Pa was used to calculate additional fan operating

costs, assuming the heat exchangers are removed during the months when they are not

used. Pressure drop in the piping system was estimated to calculate operating cost of the

circulation pump. The original heating method was assumed to be gas, either natural gas

or propane, and the fuel price therefore varied from 8-20 $/GJ.

Since there were several variables, a few different cases were calculated. The case

closest resembling the conditions at Kylylahti were in a region with an average annual

temperature of 2 °C. The smallest distance tested in the study was 200 m. With airflows

of 200 m3/s and an investment cost of 1,2 million Canadian dollars, the payback period

was approximately 3-10 years depending on fuel price. For a fuel price of 20 $/GJ, the

pay-back time was ~3 years with a net saving of ~400 000 $/year, and for a fuel price of

13 $/GJ the payback period was ~5 years with a net saving of ~250 000 $/year.

Through the analysis of the results from the Canadian study, it was observed that longer

distance between the airshafts and lower fuel price made the payback period longer.

Higher airflow rates while the distance is longer and the fuel price lower, made the

payback period longer by making the investment cost bigger. Higher airflow rates could

also have shorter payback period compared to lower rates, due to the bigger energy

savings, if the distance is shorter and the fuel price higher. It was also noticed that small

variations in the variables could have a big effect on the results, and the investment cost

is only a rough estimate. Therefore it is implicated that the results of the study can’t be

used to draw conclusions for real, specific cases. Calculations have to be made for every

specific case. How a ventilation station with battery heat exchangers can be constructed

is partially shown in figure 5.

9

Figure 5. A ventilation station with battery heat exchanging

(Photo: Andreas Markström and Arjun Mohan)

2.5 Zinkgruvan – an example of plate heat exchanging in mine ventilation

Located in the small community of Zinkgruvan close to Askersund, Sweden, is the mine

Zinkgruvan owned by the mining company Lundin Mining. In 2013 the heating system

for one of the ventilation stations at Zinkgruvan was rebuilt. The old system consisted

of six oil-burners and the new system consisted of a plate heat exchanger. The plate heat

exchanger is shown in figure 6. The oil-burners were retained for additional heating

during low temperatures when the heat exchanger does not suffice. The project is

considered to be very successful with a payback period of only 1,9 years. The estimated

need for fuel-oil with the old heating system was about 310 m3/year , and the need after

installing the heat exchanger is estimated to be 10 m3/year (~97% reduction). (Walther

2015)

10

Figure 6. The heat exchanger in Zinkgruvan. Exhaust air enters the heat exchanger at the bottom

left, and the heated supply air leaves at the right

The heat exchanger is made up of 12 containers. The ambient air is drawn into the heat

exchanger from the sides and the return air comes in from the bottom. The airflows are

shown in figure 7 where the grey squares represent the heat exchanger plates.

(Arvidsson 2015)

Figure 7. Section principle sketch of the plate heat exchanger in Zinkgruvan

11

In the Zinkgruvan mine both the supply airflow and the return flow is 150 m3/s. The

supply air fan has an installed power of 500 kW (Walther 2015) and is working against

a differential pressure of approximately 4 000 Pa, which of the pressure drop over the

heat exchanger is estimated to 200 Pa (Arvidsson 2015). The mine operates at a depth to

about 1200 m. When the exhaust air reaches the surface it has a temperature of 11 °C

and the relative humidity is 100%. To avoid icing in the supply airshaft the air is heated

to 3 °C when it is colder outside. (Walther 2015) In Zinkgruvan the average annual

temperature is 6 °C (SMHI 2014). The heat exchanger’s temperature transfer efficiency

is 50% and the wet temperature transfer efficiency is 60% (Arvidsson 2015). The heat

exchanger can heat the supply airflow to 3 °C as long as the ambient air temperature

exceeds -9 °C. If it is colder the oil burners will start. (Walther 2015)

Depending on how far the return air is cooled, the vapor condensates at a rate of 1200-

3600 kg/h. A small inclination helps the condensate to flow backwards to the return

airshaft. Due to the wide space of 15 mm between the plates and the heavy

condensation, accumulation of deposits is presumed not to affect the heat transfer.

(Walther 2015)

When it is so cold that the oil burners start, air is drawn directly in to the burners

without passing the heat exchanger. This leads to a smaller airflow passing the heat

exchanger, decreasing the risk for icing in it. If it is so cold that this function is not

enough to prevent freezing, the total flow through the supply air fan (and consequently

the heat exchanger) can be temporarily reduced. (Thoresen 2015) The winter of 2013-

2014 was considered a mild winter. During that winter there was only one incident of

icing in the heat exchanger. It was resolved by adjusting the ambient airflow which was

distributed unevenly. (Walther 2015)

The old system had a problem with noise disturbing the closest neighbors. The noise

was composed by both noise from the exhaust air fan, dissipated from 800 m below the

surface, and by suction noise from the supply air fan. (Walther 2015) During the

rebuild, a building lined with soundproofing material was built around the supply air fan

and shaft, and the ducts and the bottom half of the heat exchanger are lined with

soundproofing perforated sheet metal (Arvidsson 2015). Together these measures

reduced the noise to an acceptable level (Walther 2015).

12

3. Method

Enthalpies and energy transfer has been calculated for every hour of a year-long period.

Since the data of fan usage before February 2014 is not available, all calculations have

been done for the period of February 2014 to January 2015. To take different variables

into account, the calculations have also been done for several cases. To do calculations

for all possible combinations of variables would be excessive, so eight different

combinations were chosen – constituting eight different cases. A complete list of what

values of variables were used in each case is attached in appendix B. In case 1 the

variables were chosen to give the shortest possible payback period within reasonable

variations in variables – the best case scenario, and in case 8 the variables were chosen

to give the longest possible payback period – the worst case scenario.

All calculations have been made in Microsoft Excel ™.

First, the energy demand was calculated, followed by maximum available energy and

energy that can be transferred through heat exchanging. After that, rough designs of the

systems were made in order to estimate investment cost and calculate change in

operating cost (constituted by changed fan operation, addition of a pump and change in

LPG-consumption). Lastly, the payback period and change in carbon emissions were

calculated.

3.1 Energy demand

To calculate the energy needed to heat ambient air to the required supply air

temperature, an energy balance for the mass flow of air has to be made. The energy

balance for a mass flow can be expressed as (2) to consider the energy content of both

the dry air and the vapor. (Cengel & Boles 2015)

( )

Enthalpy, h, of air consisting of dry air and vapor, is calculated with (3). Enthalpy for

dry air, ha, is calculated with (4) (Cengel & Boles 2015) and enthalpy for water in vapor

form, hg, and Cp is tabulated values that has been included in the model. For values not

included in the tables, a linear relation has been used and values were interpolated. The

tables that have been used are available in Cengel & Ghajar (2011) and Cengel & Boles

(2015).

To calculate the energy demand, the enthalpy for ambient air and the enthalpy for

ambient air heated to the required temperature, were used.

For the ambient air, hourly values for temperature and relative humidity were obtained

from the Finnish Meteorological Institute, collected from the weather station at Joensuu

13

airport, approximately 30 km from Kylylahti. Figure 8 shows daily average values for

ambient air temperature and relative humidity for February 2014 to January 2015.

Figure 8. Ambient air temperature and relative humidity at Joensuu airport from February 2014 to

January 2015

To convert relative humidity to specific humidity, and vice versa, (5) (Cengel & Boles

2015) was used. Pg is a tabulated value available in Cengel & Ghajar (2011) and Cengel

& Boles (2015). On the supply air side, the total pressure is equal to atmospheric

pressure, and on the return air side, the total pressure is constituted of atmospheric

pressure and the return air fan differential pressure.

Since the volume flow and pressure measurements are unreliable, the fan speed has

been used to calculate the fan differential pressure. Logged data for fan speed, see

appendix C, and given operating points from the logged data, see table 1, were used in

the affinity law (6) to calculate the differential pressure. The operating points are from

the program Ventsim™, except for the differential pressure for the supply air fan, which

is from logged data (from 11 July 08.00).

(

)

Table 1. Given operating points for the two main fans

Given operating points

Fan speed Airflow Differential pressure

Supply air fan 55% 123 m3/s 640 Pa

Return air fan 63% 96 m3/s 820 Pa

0,0

20,0

40,0

60,0

80,0

100,0

120,0

-30,0

-20,0

-10,0

0,0

10,0

20,0

30,0

Re

lati

ve h

um

idit

y [%

]

Am

bie

nt

air

tem

pe

ratu

re [

°C]

Climate data 2014/2015

Ambient air temperature Relative humidity

14

To calculate the enthalpy for the supply air heated to the required temperature, different

temperatures were used: 2, 3, 4, and 6 °C. These temperatures chosen with the

assumption that the supply air temperature regulation will be automated and other

measures will be taken to reduce the amount of free water in the supply airshaft, making

it probable that the supply air temperature in the future will be lower than today. Since

no condensation occurs on the supply side and it is presumed no free water is present,

the specific humidity stays constant through the heating process.

3.1.1 Volume- and mass flows of air

To calculate energy transfer, both volume flow and density has to be known. Volume

airflow was calculated with (7), known operating points in table 1, and logged data for

fan speed, appendix C.

The calculated airflows has been adjusted such that if the ambient air temperature is

lower than the required supply air temperature, the supply airflow is at least the same as

the return airflow. This is to prevent ambient air causing icing in the ramp, but also

because the energy demand appears to be 0 if there is no flow. If the return air fan is

disabled the air can still get out through the ramp, although no energy is available for

heat exchanging. The supply airflows had to be adjusted for 140 of the 8 760 hours of

the period.

In all other cases the variation in airflow has not been changed since variation and

stoppages is how the system works normally. During the logged period the ventilation

strategy during blasting has changed from having the fans disabled to keeping them on,

there has been problem with fogging and icing in the ramp and there has been problems

with vibrations in one of the fans which was resolved by increasing the rpms for both

fans.

After the flows have been adjusted the average supply airflow is 132,4 m3/s and average

return airflow is 103,1 m3/s. How the daily average values of airflows have varied over

the period February 2014 to January 2015, is shown in figure 9.

15

Figure 9. Daily average volume flows of supply air and return air

The density of moist air was calculated with (8) (McPherson 1993) where Pv first has

been calculated with (9).

3.1.2 Normal correction with degree days

To make the calculated energy demand more generalised it can be normal corrected. To

do that, degree days is used to calculate a correction factor which tells how much colder

or warmer a month has been compared to the same month during a normal year.

The number of degree days is calculated with the difference between a daily mean

temperature and a balance temperature. The balance temperature represents the

temperature to which active heating is necessary. To calculate the number of degree

days, the difference between the daily mean temperature and the balance temperature is

added for all the days of the month which are colder than the balance temperature. Days

whose mean temperature is higher than the balance temperature have no effect on the

number of degree days. (Heincke et al. 2010)

To calculate the number of degree days, several different balance temperatures have

been used - the different required supply air temperatures. The normal temperatures that

have been used are daily averages based on the period 1981-2010, from the weather

station Joensuu airport. In figure 10 the difference between the daily averages from the

period February 2014 to January 2015, and the same months during 1981-2010 is

0

20

40

60

80

100

120

140

160

180

[m3/s] Airflow

Supply air Return air

16

shown. The spring of 2014 was warmer than the normal year, but the autumn and

January 2015 was colder than normal.

Figure 10. Daily average temperature of the normal period of 1981-2010 and of the period

February 2014 to January 2015

To calculate the correction factor with (10), a statistical value of degree days for the

month and year of the calculated energy consumption and a normal value of degree days

for the same month (degree daysnorm) is used. (SMHI n.d.)

If the correction factor is more than 1 it means that the specific month has been colder

than the normal month and vice versa.

Calculated correction factors are shown in table 2 with the number of degree days. The

energy demand that needs to be met by LPG is then adjusted with (11) to a normal

consumption - what the energy demand would have been during a normal month.

(SMHI n.d.)

For months without any degree days during a normal year, the calculated energy

demand for the period was cut out. With the required supply air temperature 6 °C, there

is two degree days in June a normal year, but for the period 2014/2015 there is none.

Since no correction factor can be calculated, the energy demand for that month was not

adjusted.

-30,0

-20,0

-10,0

0,0

10,0

20,0

30,0

Am

bie

nt

air

tem

pe

ratu

re [

°C]

Daily average temperature

2014/2015 Normal period

17

Table 2. Degree days for normal months and during the period 2014/2015, and corrections factors,

corresponding to the tested balance temperatures Balance

Temp. 2°C 3°C 4°C 6°C

Norm.

2014/

2015

Corr.

factor Norm.

2014/

2015

Corr.

factor Norm.

2014/

2015

Corr.

factor Norm.

2014/

2015

Corr.

factor

Feb 322 109 0,340 350 136 0,390 378 164 0,434 434 220 0,507

Mar 361 69 0,191 392 92 0,235 423 120 0,284 485 182 0,375

Apr 198 30 0,154 228 44 0,194 258 61 0,235 318 100 0,315

May 33 2 0,067 53 7 0,125 77 14 0,183 137 34 0,251

Jun 0 0 - 0 0 - 0 0 - 2 0 -

Jul 0 0 - 0 0 - 0 0 - 0 0 -

Aug 0 0 - 0 0 - 0 0 - 0 0 -

Sep 0 1 - 0 2 - 0 4 - 0 10 -

Oct 0 59 - 0 75 - 0 93 - 0 142 -

Nov 5 91 20,231 14 117 8,651 28 143 5,199 71 198 2,797

Dec 135 196 1,451 166 227 1,367 197 258 1,309 259 320 1,235

Jan 292 312 1,068 323 343 1,062 354 374 1,056 416 436 1,048

3.2 Available energy

To calculate the largest possible energy transfer without freezing the condensate, the

enthalpy for return air and the enthalpy for return air chilled to the lowest allowable

exhaust air temperature, were used in (2).

The enthalpy for the return air was calculated with (3), (4) and (5). The condition of the

return air is known through individual measurements and observations. The highest

point of measurement of return air condition is 142 m below the surface, where

occasional measures during early 2015 indicates a temperature of 8,5 °C and a relative

humidity of around 92%. Since free water is present, a relative humidity of 100% at the

surface level is likely and the temperature is probably 8,5 °C or slightly lower. Since the

return air temperature can show a slight variation over the year and is affected by how

the air is routed down in the mine, enthalpy is calculated for several different states; at

the highest point of measurement and at a temperature of 8,5±1 °C (7,5, 8, 8,5, 9, 9,5)

and a relative humidity of 100%.

For the exhaust air, the enthalpy was calculated for when the air was chilled to the

lowest allowable exhaust air temperature. To keep a margin from the freezing point, the

lowest allowable temperature was a bit higher, both 0,5 and 1 °C was tested. If

condensation occurs, the relative humidity is 100%. If there is no condensation, the

specific humidity is the same as for the return air. Whether condensation occurred or not

was estimated by first calculating Pv with (12) (Cengel & Boles 2015). Pv was

calculated for the same specific humidity as the return air, and the temperature

corresponding to that Pv was estimated with tabulated values from Cengel & Ghajar

2011 and Cengel & Boles 2015. That temperature was then compared to an

approximated temperature of the exhaust air to determine whether condensation

18

occurred. The exhaust air temperature was roughly calculated with (2), (3) and (4)

where the air temperature was the unknown factor. (The enthalpy of the return air was

already calculated and the other factors are for saturated air at 4 °C.)

3.3 Energy available through heat exchanging

For the calculations on heat exchanging, two different heat exchangers have been used.

One is a battery heat exchanger made by the company Luvata. Product sheets for the

battery system are included in appendix D. The other one is a plate heat exchanger made

by the company Gupex. There are no product sheets available for the plate heat

exchanger, so all relevant data is estimated by Claes Arvidsson (2015) from Gupex.

Both heat exchangers have been dimensioned for the average airflows.

Possible energy transfer through heat exchanging was calculated with (2) as the

difference between ambient air enthalpy (already calculated in section 3.1) and the

enthalpy for air heated through heat exchanging. The enthalpy for air heated through

heat exchanging was calculated with (3) and (4) where the heated air temperature is

unknown. To calculate what temperature the supply air reaches with heat exchanging,

wet temperature efficiencies for each heat exchanger were used in (1). The specific

humidity is still presumed to be the same as for the ambient air.

The wet temperature transfer efficiency for the battery set is included in appendix D and

is shown in table 3. It is assumed to vary linearly between the stated values. The

temperature threshold to prevent icing on the batteries is -3,2 °C. That means as long as

the ambient air temperature is between 0°C and -3,2 °C the variation between b and c is

used to calculate the wet temperature transfer efficiency. When the outdoor temperature

falls below -3,2 °C the variation between b and e is used. The batteries on the return air

side were chosen with bigger distance between the fins to decrease the risk for

accumulation of particles, but experience shows that similar systems usually require

washing. Since washing can be hard to do during the winter, there is a risk for a

decrease in efficiency. How much the efficiency will be affected is unknown, but an

estimation was made that particle build-up can reduce the efficiency by 10-20%

(Danielsson 2015). Therefore, reduced temperature transfer efficiencies of 10% and

20% were tested in the continued sensitivity analysis.

19

Table 3. Wet temperature efficiency for the battery heat exchanger at different ambient air

temperatures. (*) marks the dimensioning case when the efficiency has been reduced.

a b c d e

TAmbient air

[°C]

5 0 -5 -30 -30*

TSupply air

[°C]

6,6 4,3 2,1 -14,7 -23,9

52% 53% 54% 40% 16%

The wet temperature transfer efficiency for the plate heat exchanger is assumed to be

55%, for all ambient air temperatures (Arvidsson 2015). An efficiency of 60% is more

reasonable if both airflows are of the same size (Arvidsson 2015) but has also been

tested in this work. The heavy condensation is presumed to prevent any accumulation of

particles on the plates, keeping the efficiency unaffected by particle build-up.

To get how much energy transfer in the heat exchanger that is possible in reality, the

possible energy transfer through heat exchanging is compared to the largest possible

energy transfer without freezing the condensate (already calculated in section 3.2). If the

possible heat transfer through heat exchanging is larger than the largest possible energy

transfer without freezing the condensate, the condensate will in reality freeze in the heat

exchanger. Since ice in the heat exchanger is unacceptable, the real possible energy

transfer through heat exchanging must be limited to the largest possible energy transfer

without freezing the condensate during those occasions.

To then get how much energy is transferred through the heat exchanger and how much

of the energy demand (already calculated in section 3.1) that must be met by LPG, the

real possible energy transfer through heat exchanging is compared to the energy

demand. If the energy demand is larger than the real possible energy transfer, the

transfer is limited to the real possible transfer through heat exchanging and the rest of

the energy demand has to be met by LPG. If the real possible transfer through heat

exchanging exceeds the demand, the heat exchanger meets the whole demand.

The energy demand that could not be met by heat exchanging, and instead had to be met

by LPG, was normal corrected with (11).

3.4 Design

To calculate investment cost and the change in operation cost, rough designs of the

systems had to be made.

In Kylylahti the return air fan is positioned vertically and above ground. This limits the

possible configurations of the ducts, especially for the plate heat exchanger. To turn the

fan horizontally would presumably demand reconstruction of the fan to the extent that

the cost would be comparable to buying a new fan (Andrés 2015).

20

To achieve a configuration of the battery heat exchanger system which is practical from

an aerodynamical standpoint, the return airflow needs to be bent 90°. That way it can be

directed in to an exhaust building built with heat exchanger batteries mounted in the

walls, which the flow has to pass through. The principle for the supply air side is the

same but there is no redirection necessary. The principle for the battery heat exchanger

system is shown in figure 11. On the return air side there needs to be three heat

exchangers, each 5 m wide and 2,4 m high. For each wall to hold one heat exchanger

the exhaust building needs to be approximately 6 m wide and to reach the return airflow

it needs to be 8 m high. On the supply air side the burner building needs to be rebuilt

and extended to hold four heat exchanger of the same size. The long side of the building

needs to be approximately 12 m, the short sides 6 m and all 4 m high. To circulate the

fluid, an estimation of 170 m of piping is necessary.

Figure 11. Overview of a possible battery heat exchanger system

The configuration of the specific type of plate heat exchanger that is used in this work

demands the return air to flow in from below. In this case that requires the whole heat

exchanger to be placed above the return air fan, 3,5-4 m above ground. The heat

exchanger is approximately 5 m x 7,5 m x 25 m. A duct would then lead the heated

supply air to the burner building which has to be rebuild to receive the supply airflow

for further heating and passing it on through the supply air fan. The principle for the

plate heat exchanger system is shown in figure 12.

21

Figure 12. Overview of a possible plate heat exchanger system

The condensation in the heat exchangers is presumably too polluted to be discharged

without purification. The options are either to let the condensate flow down the return

airshaft or to lead it to the ponds where the mine’s process water is purified. If the

condensate is allowed to flow back down the shaft, it can be pumped up with the

process water, but this approach will result in a bigger strain for the exhaust air fan and

unnecessary pumping. In this work it is assumed the condensate will be led by a ditch

from the ventilation station to the ponds.

3.5 Investment cost

All investment costs are estimated in SEK, and have been recalculated to Euro. Three

different exchange rates were tested: 8,5, 9,0, and 9,5 SEK/Euro. An additional cost of

10% has been added to the total investment cost for unforeseen expenses.

The investment cost for the batteries were estimated to approximately 375 000 SEK

(Danielsson 2015) each. The price vas varied with ±50 000 SEK in the calculations. The

cost for transportation of the batteries to Kylylahti was estimated to approximately

40 000 SEK (Danielsson 2015).

To build the supply and exhaust buildings needed for the battery heat exchanger system,

the prices have been estimated with Wikells NYB (2008) to 500 SEK/m2 for floors, 1

200 SEK/m2 for walls and 750 SEK/m

2 for roofs, which gives a total cost of

484 800 SEK for both buildings. The cost for redirecting the return airflow has been

estimated to 500 000 SEK.

The price for the circulation pump was estimated to 80 000 SEK and the price for a

frequency converter to the pump was estimated to 25 000 SEK. The cost for the 170 m

of piping needed for the circulating fluid was estimated to 4 800 SEK/m with Wikells

22

VVS (2007), whereof about 60m of pipes need to be dug down into the ground at a cost

of 1 900 SEK/m estimated with Wikells NYB (2008). In total the cost for piping is

estimated to 930 000 SEK.

The investment cost for the plate heat exchanger, ducts and installation has been

estimated to approximately 8,6 MSEK (Arvidsson 2015). The cost was varied with

±100 000 SEK in the calculations.

In addition, there is a cost of 130 SEK/m, estimated with Wikells NYB (2008),

71 500 SEK in total, to excavate a ditch for redirecting the condensate regardless of the

type of system.

3.6 Operating cost

The operations will change with the installation of a heat exchanger. In this work, the

change in fan power, the addition of a pump (only for the battery system) and the

change in LPG-demand is included. The change in operating cost was calculated as the

increase (current cost not included) in electricity costs and the decrease in LPG-costs –

from the cost without a heat exchanger to the cost with a heat exchanger under the same

conditions. The cost for maintenance was not included.

Different LPG- and electricity prices were tested;

LPG: 0,5, 0,6, 0,7, 0,8 and 0,9 €/kg.

Electricity: 60, 65, 70 and 75 €/MWh.

3.6.1 LPG

The energy demand that can’t be met with heat exchanging, require heating with LPG.

The energy demand with and without heat exchanger that has to be met with LPG was

normal corrected, then the LPG-demand was calculated with the lower heating value

46,4 MJ/kg (Teboil 2013). Since the LPG-burners are direct working, the efficiency is

assumed to be 100%.

Since the regulation of the supply air temperature is done manually and several different

supply air temperatures have been tested, the LPG-demand with a heat exchanger has

been compared to a calculated LPG-demand instead of the real demand of the

2014/2015 period.

3.6.2 Fan operation

To calculate the difference in the power needed by the fans, caused by the changes in

the system, the change in pressure needs to be calculated first. Since the only change is

caused by additional losses, all but the losses is cut from the right side of (13) (Cengel

& Cimbala 2010) which leaves (14).

23

The pressure change can then be calculated as (15) (Cengel & Cimbala 2010) where the

losses is made up by losses caused by the heat exchanger and friction and minor losses

in the ducts.

The loss caused by the heat exchanger is provided by the supplier. For the plate heat

exchanger system, the pressure drop is estimated to 400-450 Pa for both sides of the

heat exchanger for the dimensioning airflows (Arvidsson 2015). That includes both the

heat exchanger and the required ducts. Both values have been tested in the calculations

for operating costs. For the battery heat exchanger, the pressure drop is around 300 Pa

on the supply air side and 350 Pa on the return air side for average airflows, see

appendix D.

For the plate heat exchanger, the friction losses are included in the estimation of

pressure drop of the heat exchanger and for the battery heat exchanger there is no

section of straight duct long enough for friction losses.

Minor losses are calculated with (16) (Cengel & Cimbala 2010) where KL is the loss

coefficient. KL-coefficients that have been used in this work are from Cengel &

Cimbala (2010). Air velocity is calculated from the annual average of the volume flows

and the cross section area.

[ ]

For the battery system a 90° bend is required to re-route the return airflow. The

bend is assumed to be flanged, and the diameter is 3,5 m. The resulting pressure

drop is approximately 25 Pa.

After the 90° bend there is a sudden expansion into the exhaust building, into a

rectangular cross section with the measures 4x6 m. The resulting pressure drop

is approximately 30 Pa.

For the battery system there is a sudden contraction from the supply building

into the burner building, from an area of 12x4 m to 9x4 m or 4,5x4 m depending

on if the door to the second burner is open. The pressure drop was approximated

to 5 Pa which is an average of whether the door is open or not.

For the supply airflow in the current system there is an inlet to an area of 9x4 m

or 4,5x4 m depending on whether the door is open or not, and the average

pressure drop is approximated to 10 Pa.

For the return airflow in the current system the pressure drop for the evasé is

approximated to 80 Pa.

24

The pressure drops that represent the inlet and outlet of the heat exchangers are

expected to be included in the heat exchangers pressure drop. To compensate for the

loss of the old minor losses of the inlet and outlet when the heat exchangers are applied,

the pressure drops for the heat exchanger systems are reduced with the pressure drops

calculated for the inlet and the evasé.

The change in fan pressure applies for every hour of the year as long as there is a flow

through the fan.

The increased need in fan power caused by the increase in pressure is calculated

with (17) (Alvarez 2003) and the calculated values for volume flows were used.

The motor efficiency was assumed to be 95%. The fan efficiency has been estimated

using the affinity laws (6) and (7) and the given operating points which are shown in

table 1, to calculate the maximum operating points and deduce the fan efficiency from

the fan performance curves, see appendix A. The fan efficiency was estimated to 71%

for the supply air fan and 76% for the return air fan, regardless of flow, but they have

been varied with ±2% in the calculations.

3.6.3 Pump operation

In order to circulate the fluid for the battery heat exchanger system, a circulation pump

is necessary.

Total head, Hpump, for pumps is calculated with (18) (Cengel & Cimbala 2010), but

since the pump is working in a closed system and the fluid velocity is constant through

the system, everything but Hloss,tot can be cut from the right side of the equation, leaving

(19) (Cengel & Cimbala 2010).

Hloss is then calculated as (20) (Cengel & Cimbala 2010).

The loss caused by the heat exchanger batteries is provided by the supplier. The

pressure drop for full flow of the fluid is 97 kPa through the supply air batteries, and

101kPa through the return air batteries, see appendix D.

It was assumed that as long as there is an energy need the fluid flow is full, 34,5 l/s, see

appendix D. Approximately 170 m of piping is necessary to circulate the fluid, how the

25

piping is designed is shown in figure 13. The pipes were assumed to be of stainless

steel, insulated, with a diameter of 129 mm.

Figure 13. Schematic diagram of the piping connecting the batteries

The Reynolds number was calculated with (21) (Cengel & Cimbala 2010) to around

201 500. For internal flow in pipes, a Reynolds number higher than 4 000 means that

the flow is turbulent (Cengel & Cimbala 2010).

Friction loss through the pipes is calculated with (22) (Cengel & Cimbala 2010). Since

the flow is turbulent, the friction factor is calculated with (23) (Cengel & Cimbala

2010), where Ɛ was assumed to be 0,002, with a result of 0,016.

√ (

⁄

√ )

26

Minor losses are calculated with (24) (Cengel & Cimbala 2010).

[ ]

All bends and branches are assumed to be flanged. Every battery has two connections in

each direction, and therefore two expansions and two contractions. One ball valve is

connected to each heat exchanger and five threaded unions are assumed to be needed in

each direction. The loss coefficient associated to the connection to the pump is

estimated to KL=3. In total, the flow is estimated to pass through an average of minor

losses of KL=15, where the different KL-coefficient are from Cengel & Cimbala (2010).

The total head loss in the circulating fluid system was calculated to approximately 33 m.

The pump is assumed to be turned off when there is no need for heating. The power

demanded by the pump is deduced from the pump characteristic curve, appendix E, to

15 kW.

3.7 Payback period

To assess the profitability of the investment the payback-method (25) (Björnsson n.d.)

is used. It is a simple method often used as an initial sifting-tool, calculating how much

time it takes for the investment to be paid back due to the reduced expenditures. The

payback-method is an approximate method since there is no consideration for factors as

change in energy prices or the life span of the investment. By taking the interest rate

into account the method becomes more accurate because the provisioning excess is

converted to (a form of) current value. The interest represents the most prosperous

alternative investment. (Grubbström & Lundquist 2005)

(

)

The interest rate used was 10%, which is what is usually used for similar assessments at

Boliden (Rånman 2015).

There are no official guidelines for what an acceptable payback period is, but 5-6 years

are common for energy investments. Longer periods can sometimes be accepted, and

shorter periods are of course preferential. Mines often have potential for a longer life

span than what is currently identified, but so far, the mine has known ore corresponding

to a life expectancy of around 6 years at current production rate. (Rånman 2015).

3.8 Carbon emissions

The increase in power consumption and the decrease in LPG-consumption implicate a

change in the carbon emission that can be attributed to the heating done at the

ventilation station. The calculations on carbon emissions include the change in power

27

demand to the fans, the power demand to the circulation pump (only for the battery

system) and the change in LPG-consumption.

There are several ways to assess carbon dioxide emissions from a change in electricity

use, of which margin electricity and average electricity are two common standpoints

(Sköldberg et al. 2006). In this work, two approaches to the margin perspective and the

method Boliden usually uses will be used.

Margin electricity can be understood as the production methods that will be increased or

decreased to account for the changed electricity used. It accounts for that production

methods has different regulation possibilities and prices. A problem with margin

electricity is that the emission factors are difficult to estimate. With the average

electricity perspective, the change in emissions is instead ascribed to all electricity use

(not just the changed). With this method, all emissions can be added and will

correspond to the real amount. If the average electricity perspective is used, the system

boundary is important. One possible boundary is the national border of the country

where the electricity is used, for instance with the argument that that is the production

which can be regulated with national instruments. The Nordic region is another possible

boundary since electricity moves freely within the Nordic electricity market. There are

also arguments for using Northern Europe as system boundary. (Sköldberg et al. 2006)

To account for carbon dioxide emissions from electricity use, Boliden practices the

Scope 2 Guidance within the GHG Protocol Corporate Standard which is developed by

World Resources Institute. More specifically, that means that Boliden practices a

location based method, using a national average emission factor. This method is also

used for assessment of changed emissions. (Ryman 2015)

The changes in carbon emission caused by the changed electricity demand have been

calculated with three different emission factors in this work:

An emission factor for short term margin electricity of 969 kg/MWh

(Naturvårdsverket 2015).

An emission factor for long term margin electricity of 375 kg/MWh

(Naturvårdsverket 2015).

A location based emission factor of 0,191 kg CO2/kWh, which is for Finland in

2014. (Department for Environment Food and Rural affairs n.d.)

The change in carbon emissions due to the change in LPG-consumption was calculated

with the emission factor 2997,855 kg CO2/ton LPG (Naturvårdsverket 2015).

28

4. Results

First, the calculated energy demand and maximum available energy are declared. That is

followed by costs, payback period and carbon dioxide emissions for the eight tested

cases. Case 3 is deemed as the case best corresponding to current circumstances and

case 5 is deemed as best representing the future. More extensive results are declared for

the heat exchanger system with the shortest payback period, in case 3 and 5. Lastly, the

impact of individual variables on the payback period for the heat exchanger system with

the shortest payback period, in case 5, is shown.

4.1 Energy demand

The calculated energy demand for heating is shown in table 4, both for the period

2014/2015 and corrected with the degree days-method to a normalized demand. Four

different required supply air temperatures was used, and for all temperatures is the

normalized energy demand bigger than for the 2014/2015 period. Regardless of the

required supply air temperature, there is a demand for heating from November to May,

and during the period 2014-2015 there was also a significant demand in October.

Table 4. Energy demand depending on required supply air temperature Energy demand [MWh]

Supply air temp.

2°C 3°C 4°C 6°C

2014/

2015 Norm.

2014/

2015 Norm.

2014/

2015 Norm.

2014/

2015 Norm.

Feb 489 1437 608 1560 731 1685 979 1932

Mar 345 1808 450 1915 567 1995 820 2186

Apr 178 1154 235 1210 305 1299 472 1500

May 38 561 57 455 83 451 160 639

Jun 0 0 0 0 1 0 8 39

Jul 0 0 0 0 0 0 0 0

Aug 0 0 0 0 0 0 0 0

Sep 5 0 13 0 25 0 57 0

Oct 273 0 339 0 414 0 605 0

Nov 385 19 488 56 595 114 818 292

Dec 773 533 895 654 1020 779 1269 1028

Jan 1262 1181 1387 1306 1513 1433 1764 1684

Total 3743 6693 4472 7156 5254 7756 6952 9300

The real LPG-consumption during the period February 2014 to January 2015 is

estimated to 415 tonne, which implies a heating demand for the same period of roughly

5350 MWh. The real energy demand for the period closely matches the calculated

energy demand for a supply air temperature of 4°C.

Figure 14 shows a correlation between the energy demand for heating and the difference

between the required supply air temperature and ambient air temperature. The heating

demand has a strong dependency on ambient air temperature, but is also influenced by

humidity and variation in airflow.

29

Figure 14. Normalized heating demand and the difference between a required supply air

temperature of 3°C and ambient air temperature

4.2 Available energy

Tables 5 and 6 shows maximum available energy, which is how much energy that can

be extracted if the return air is cooled to the lowest allowable exhaust air temperature.

The maximum available energy shows a variation over the year which is due to the

variations in return airflow.

Table 5. Maximum available energy for different states of

return air, cooled to either 1 °C or 0,5 °C

Maximum available energy [MWh/year]

State of

return air Cooled to 1°C Cooled to 0,5°C

T=8,5°C

=92% 15180 16216

T=9,5°C

=100% 19126 20157

T=9,0°C

=100% 17950 18983

T=8,5°C

=100% 16771 17806

T=8,0°C

=100% 15588 16625

T=7,5°C

=100% 14401 15440

-20,0

-15,0

-10,0

-5,0

0,0

5,0

10,0

15,0

0

500

1000

1500

(Tsu

pp

ly-T

amb

ien

t) [

°C]

Ene

rgy

de

man

d [

MW

h]

Energy demand and temperature difference

Energy demand Temperature difference

30

Table 6. Maximum available energy over the course of a year

Maximum available energy [MWh]

Month T=8,5 °C, =100 %, cooled to 0,5 °C

Feb 1555

Mar 1706

Apr 1557

May 1689

Jun 1495

Jul 1524

Aug 1512

Sep 1398

Oct 1427

Nov 1385

Dec 1309

Jan 1250

Total 17807

Figure 15 shows a correlation between the maximum available energy and the return air

volume flow. Maximum available energy is almost exclusively dependent on the return

air volume flow, since the state of the return air is assumed to be constant over the

course of the day and of the year.

Figure 15. Return air volume flow and the maximum available energy

4.3 Costs

The reasonable range of the total investment cost for the two systems are shown in table

7. The battery heat exchanger system is the cheapest, about 50-60% of the cost for the

plate heat exchanger system.

0

200

400

600

800

1000

1200

1400

1600

1800

0

20

40

60

80

100

120

140A

vaila

ble

en

erg

y [M

Wh

]

Re

turn

air

flo

w [

m3

/s]

Available energy and return airflow

Return air flow Available energy

31

Table 7. Range of total investment cost for the two heat exchanger systems

Total investment cost [€]

Lowest possible Mid-range Highest possible

Battery

system 511 000 582 000 662 000

Plate

system 992 000 1 060 000 1 135 000

The operating cost for the system, with or without a heat exchanger, in all eight tested

cases, is shown in figure 16. For operation without heat exchanger the operating cost is

only constituted of the cost for LPG to heat to the required supply air temperature, and

for the plate system and the battery system the operating cost is constituted of the

increase in electricity cost and the cost for LPG. The operating cost with a heat

exchanger is in all tested cases lower than without a heat exchanger, and in all cases but

number 8 is the plate system the one with lowest operating cost.

Figure 16. Operating cost without a heat exchanger compared to with a heat exchanger

How the operating cost is divided between LPG and electricity is shown in figure 17 for

the battery system and in figure 18 for the plate system. For the plate heat exchanger

system the size of the electricity is dependent on both the assumed pressure drop over

the heat exchanger (400-450 Pa) and the fan efficiency. For the battery heat exchanger

system, the assumed pressure drop is the same in all tested cases and the only affecting

variable is the fan efficiency. For the battery system, LPG constitutes the biggest part of

the operation cost in all cases, but for the plate system, electricity constitutes the biggest

part in case 7 and 8. Regardless of which pressure drop was used for the plate heat

exchanger calculations, the plate heat exchanger has a bigger increase in electricity

demand than the battery system, but it also has a bigger decrease in LPG-demand.

0

100 000

200 000

300 000

400 000

500 000

600 000

700 000

1 2 3 4 5 6 7 8

Op

era

tin

g co

st [€

/ye

ar]

Operating cost

No heat exchanger Plate system Battery system

32

Figure 17. Operating cost for the battery heat exchanger system

Figure 18. Operating cost for the plate heat exchanger system

4.4 Payback period

The payback period for the two systems is shown in figure 19. In all the tested cases, the

plate heat exchanger system has the longest payback period and the battery system has

the shortest. In cases 1-6 the battery system’s payback period is only 49-59% of the

plate heat exchanger’s and in case 7 it is only 41%. In case 8 the payback period for the

plate heat exchanger system is over 72 years, but both systems are profitable in all

tested cases. In most cases the payback period of the battery heat exchanger system is

shorter than both the most common payback period within the company and the

anticipated life expectancy, the exceptions being case 3 and the worst case scenario. For

the plate heat exchanger on the other side, the payback period is longer than both the

0

50000

100000

150000

200000

250000

300000

350000

1 2 3 4 5 6 7 8

Op

era

tin

g co

st [€

/ye

ar]

Operating cost - battery system

Electricity LPG

0

50000

100000

150000

200000

250000

300000

350000

1 2 3 4 5 6 7 8

Op

era

tin

g co

st [€

/ye

ar]

Operating cost - plate system

Electricity LPG

33

common period and the anticipated life expectancy in most cases, the exceptions being

case 1,2 and 6.

Figure 19. Payback period for the two heat exchanger systems

4.5 Carbon dioxide emissions

The annual carbon dioxide emissions from the system without a heat exchanger, and for

the two heat exchanger systems, are shown in figures 20-22, from the perspective of

short term margin and long term margin electricity, and the GHG protocol company

standard. Carbon dioxide emissions included are from the use of LPG and the change in

CO2 from the changed electricity use.

Regardless of what emission factor is used, both heat exchanger systems give less

emission than heating with only LPG, in all tested cases. That means that the decrease

in carbon dioxide from LPG is bigger than the increase in carbon dioxide from the

increased electricity use.

Comparing the three perspectives, the short term margin electricity gives the generally

smallest decrease with a heat exchanger, while the GHG protocol standard gives the

generally biggest decrease. Generally, the decrease compared to the system without a

heat exchanger is bigger than the difference between the two heat exchanger systems.

From a short term margin electricity perspective, the plate heat exchanger system gives

the biggest decrease in cases 1-3, the systems are equal in case 5, and in case 4 and 6-8

the battery heat exchanger system gives the biggest decrease. In cases 1-3 and 5, the

pressure drop over the plate heat exchanger has been assumed to 400 Pa, in cases 4 and

6-8 it was assumed to 450 Pa. The emissions caused by the plate system are in general

90% of the emissions from the unchanged system, and emissions caused by the battery

system are in general 87%.

3,20

4,67

12,49

7,41 6,44 5,88

10,86

1,89 2,62

6,25

3,66 3,40 3,24 4,43

9,38

0123456789

10111213

1 2 3 4 5 6 7 8

Pay

bac

k p

eri

od

[ye

ars]

Payback period

Plate system Battery system

>72

34

Figure 20. CO2-emissions for the three different systems from a short term margin electricity

perspective

From a long term margin electricity perspective, the plate heat exchanger gives the

biggest decrease in all tested cases except for case 8 where the two systems are equal.

Emissions from the plate system are in general 51% of the ones from the unchanged

system and emissions from the battery system 55%.

Figure 21. CO2-emissions for the three different systems from a long term margin electricity

perspective

With the GHG protocol standard the plate heat exchanger gives the biggest decrease in

all tested cases. The emissions from the plate heat exchanger system are in general 40%

of the emissions from the current system, and the emissions from the battery heat

exchanger system are in general 45%.

0

500

1000

1500

2000

2500

1 2 3 4 5 6 7 8

CO

2 [

ton

ne

/ye

ar]

CO2-emissions (short term marigin electricity)

No heat exchanger Plate system Battery system

0

500

1000

1500

2000

2500

1 2 3 4 5 6 7 8

CO

2 [

ton

ne

/ye

ar]

CO2-emissions (long term marigin electricity)

No heat exchanger Plate system Battery system

35

Figure 22. CO2-emissions for the three different systems calculated with the GHG protocol

standard

4.6 Detailed results

More detailed results are presented for the battery heat exchanger system for case 3 and

case 5 in table 8 and table 9.

Table 8. Detailed results for the battery system in case 3

Battery system, case 3

No heat exchanger Battery system

Heating demand [MWh] 7755,5

Energy transfer through heat exchanging [MWh] - 5010,6

Energy from burners [MWh] 7755,5 2745,0

Electricity demand [MWh] - 962

Electricity cost [€] 0 62502

LPG-demand [kg] 601721 212971

LPG-cost [€] 300860

106486

Operating cost [€] 168987

Investment cost [€] - 591881

Payback period [years] - 6,25

CO2-emissions (short term margin electricity) [tonne]

1804

1570

CO2-emissions (long term margin electricity) [tonne] 999

CO2-emissions (“Boliden method”) [tonne] 822

0

500

1000

1500

2000

2500

1 2 3 4 5 6 7 8

CO

2 [

ton

ne

/ye

ar]

CO2-emissions (GHG protocol standard)

No heat exchanger Plate system Battery system

36

Table 9. Detailed results for the battery system in case 5

Battery system, case 5

No heat exchanger Battery system

Heating demand [MWh] 7156,6

Energy transfer through heat exchanging [MWh] - 4997,4

Energy from burners [MWh] 7156,6 2159,2

Electricity demand [MWh] - 945

Electricity cost [€] 0 61394

LPG-demand [kg] 555255 167526

LPG-cost [€] 388679

117268

Operating cost [€] 178662

Investment cost [€] - 581986

Payback period [years] - 3,40

CO2-emissions (short term margin electricity) [tonne]

1665

1417

CO2-emissions (long term margin electricity) [tonne] 856

CO2-emissions (“Boliden method”) [tonne] 683

4.7 Continued sensitivity analysis

In all eight cases, both values for lowest allowable exhaust air temperature and both

values for relative humidity of the return air, were tested. It had very little impact what

values were used. Neither had any noticeable effect on the payback period of the battery

heat exchanger system, while the payback period of the plate heat exchanger system

changed only marginally, with the biggest impact in cases 3 and 7. The value of return

air relative humidity had bigger impact than lowest allowable exhaust air temperature,

but the change of both had the biggest impact. In case 3, the payback period of the plate

system was increased by 12 weeks, and in case 10 by nine weeks. Because of the

limited impact, the relative humidity of 100% and the lowest allowable exhaust

temperature of 0,5 °C were used for all cases but case 8.

In the continued sensitivity analysis, the variables from case 5 were used, while

changing one of the variables at the time. The values of the unchanged variables are

shown in appendix B.

Table 10 shows how the payback period changed with different values on the required

supply air temperature, and table 11 how it changed with different values return air

temperature. The payback period is shortest for a required temperature of 3 °C, how the

payback period changes with different required temperatures is due to the pattern of

ambient air temperatures - how common certain temperatures are. The payback period

is lower for higher return air temperatures. This is expected since more energy is

37

available for energy transfer in the heat exchanger. Compared to the highest tested

return air temperature, the payback period is half a year longer for the lowest tested

temperature.

Table 10. Change in payback period with different required supply air temperatures

Battery system, case 8: required supply air temperature

Required supply air

temperature [°C]

6 4 3 2

Payback period

[years]

3,69 3,41 3,40 3,46

Table 11. Change in payback period with different return air temperatures

Battery system, case 5: return air temperature

Return air

temperature [°C]

9,5 9,0 8,5 8,0 7,5

Payback period

[years]

3,20 3,29 3,40 3,54 3,71

Table 12 shows how the payback period changed when the temperature transfer

efficiency of the battery heat exchanger was reduced due to the assumption of particle

accumulation.

Table 12. Change in payback period with reduced

temperature transfer efficiency due to particle accumulation

Battery system, case 5:

heat exchanger efficiency reduction

Reduction in ηtemp Payback period [years]

0% 3,40

10% 3,79

20% 4,37

The impact of the fan efficiency is shown in table 13, and within the reasonable range

used, the difference in payback period is about three weeks from the lowest to the

highest efficiency.

Table 13. Payback period for different values on fan efficiency

Battery system, case 5: fan efficiency

Fan efficiency, η Payback period [years]

Ηfan,supply =0,69, ηfan,return=0,74 3,44

Ηfan,supply=0,73, ηfan,return=0,78 3,37

38

The impact of the price for the heat exchanger on payback period is shown in table 14.

Within the range used, the difference in payback period is 31 weeks from the lowest to

the highest price.

Table 14. Payback period for different prices on heat exchanger batteries

Battery system, case 5: price heat exchanger

Price for each battery [SEK] Payback period [years]

425 000 (supply side)

425 000 (return side)

3,70

325 000 (supply side)

325 000 (return side)

3,11

Table 15 show how the payback period is affected by different LPG-prices, electricity

prices and the exchange rate between Euro and SEK. The payback period is lower with

a higher price on LPG, a lower price on electricity and a high exchange rate. Changing

the electricity price has a lower effect when the price for LPG is higher. Changing the

exchange rate has bigger effect when the price for LPG is low. Changing the price for

LPG has a slightly lower effect when the electricity price is lower.

Table 15. The payback period for the battery system with all combinations of LPG-price, electricity

price and exchange rate. LPG-price is varied in the y-direction, electricity price in x-direction and

the exchange rate varies within the cells.

Battery system, case 5: electricity price, LPG-price and exchange rate

Exchange rate

(8,5/9,0/9,5)

[SEK/€] Electricity price [€/MWh]

LPG-price

[€/kg]

75 70 65 60

0,5 7,29/6,72/6,24 6,91/6,38/5,93 6,57/6,07/5,65 6,26/5,79/5,39

0,6 5,03/4,68/4,37 4,85/4,51/4,22 4,68/4,36/4,08 4,52/4,21/3,94

0,7 3,85/3,60/3,37 3,74/3,50/3,28 3,64/3,40/3,20 3,55/3,32/3,11

0,8 3,12/2,92/2,75 3,05/2,86/2,69 2,99/2,80/2,63 2,92/2,74/2,57

0,9 2,63/2,46/2,32 2,58/2,42/2,28 2,53/2,37/2,23 2,48/2,33/2,19

39

5. Discussion

In the enthalpy and energy calculations, a few necessary simplifications and

assumptions were made. The battery heat exchangers can only be dimensioned for a

relative humidity of 80% on the return air side, meaning that the wet temperature

transfer efficiency most likely is a bit higher. How the efficiency for the batteries is

reduced depending on the ambient air temperature to prevent freezing was included in

the product sheets. As seen in table 12, the accumulation of particles on the battery heat

exchanger may have considerable effect on the payback period; increasing it from 3,40

to 4,37 years if the temperature transfer efficiency is reduced by 20%. The reduction of

the battery heat exchanger system has only been made for case 5, and is not a variable

that has been changed in the other cases. If it had been, it would reasonably have had

the same effect on the payback period. How the efficiency will be reduced in reality is

impossible to know without experience from the particular system. In the calculations,

the reduction is assumed to be effective all year round. In reality the reduction will

increase slowly to reach that level perhaps first during midwinter or even later. If

washing is done regularly, perhaps even during warmer winter days, the effect should

not be as substantial. For the plate heat exchanger, the wet temperature transfer

efficiency was assumed to be the same for all ambient air temperatures. This assumption

might work as long as the ambient airflow through the heat exchanger is reduced when

it is colder outside; the rest of the flow has to be bypassed straight to the burners. For

technical reasons the whole volume flow was used in the calculations, which means that

the whole flow would be heated in the heat exchanger, a bit more efficiently than in

reality, but not as high. The efficiencies (for both systems) are also a bit uncertain since

they are based on the average flow, which means they probably will vary slightly as the

volume flows do.

Since the regulation is manual, it has been difficult to estimate an appropriate supply air

temperature for the calculations. In the calculations, the same required supply air

temperature has been used all year round, and with the future in mind, a temperature of

3°C was used in case 5, which is assumed to be the most reasonable case. But in the

continued sensitivity analysis, where the effect of only changing the supply air

temperature, it is shown that the payback period not necessarily is neither extended or

shortened by increasing or decreasing the supply air temperature. That the payback

period is shortest for a temperature of 3-4 °C, and increases whether the temperature is

increased or reduced, is probably due to the climate data for the location.

The explanation for the low impact of the return air humidity and the lowest allowable

exhaust temperature on the payback period can be explained by the low wet temperature

transfer efficiency. The battery heat exchanger system has such low wet temperature

transfer efficiency at lower temperatures, that the return airflow seldom is cooled as low

as the lowest allowable temperature. The relative humidity influence the maximum

transferable heat – not how much is actually transferred, under most conditions the heat

transfer is limited by the wet temperature transfer efficiency to avoid freezing.

As seen in table 15 in the sensitivity analysis, the economic conditions have a big effect

on the payback period. When LPG-price, electricity price and exchange rate is varied

for the battery heat exchanger system in case 5, the payback period can vary between

40

the extremes 2,19-7,29 years. A high price on LPG means the value of the saving is

bigger. The price Kylylahti mine pays for LPG has dropped more than 50% in less than

1,5 years, from about 0,8 €/kg to under 0,4 €/kg. This price may probably increase again

and that is why a slightly higher price of 0,7 €/kg in case 5 was deemed most

representative of the future. A low price on electricity means that the increase in

operating cost constituted by electricity is not as expensive. In most cases, an electricity

price of 65 €/MWh was used. If the average electricity price is 60 €/MWh instead, the

payback period for case 5 becomes 3,22 years instead of 3,40. A high exchange rate

without changing the investment cost in SEK means that the investment cost in Euro is

lower.

To calculate the investment cost, a few assumptions and estimations were made, for

example about the handling of the condensate. The geographical conditions at the site

might not allow for just a ditch to transport condensate to the ponds - pumping might be

necessary if the ponds are located higher than the ventilation station. If there is a risk of

the condensate freezing, it might have to be led to the ponds by underground piping

instead of a ditch. Luvata is only the manufacturer of the battery heat exchangers, which

means all other work will have to be done by others. All costs have been estimated

separately, but since it all has to be conducted by a contractor, it might become more

expensive. Gupex on the other hand is the contractor, making the price for the plate heat

exchanger system more reliable. To ensure the investment cost is not too low, 10% for

unforeseen cost were added for both systems. In case 5, the price of each battery was

assumed to 375 000 SEK giving a payback period of 3,40 years. If instead a higher

price of 425 000 SEK is used, which could represent other increased costs, the payback

period becomes 3,70 years.

Considering some of the simplifications and estimations that has been made, it can be

argued that the enthalpy calculations and using tabulated values for every hour of the

year have been unnecessarily meticulous, not improving the accuracy much but taking a

lot of time. But either way, the results can’t be used to predict the future, but merely to

give pointers to how it might look.

Since Kylylahti is a new mine within the Boliden organization, a lot of possible

improvements are investigated. But since they are beyond the scope of this project, their

possible effects have not been included to any greater extent, although some

consideration has been given to the required supply air temperature.

There is a possibility that the return air fan which is located at the surface can be

removed due to changes within the ventilation system underground. That would

reasonably lower the investment cost and payback period for the plate heat exchanger

system, further than used in this work, since it could be built at ground level instead of

several meters above ground. It would not lower the investment cost for the battery heat

exchanger as much, since the only change would be that the exhaust building would be

lower. If the return air fan would be removed, that would also eliminate – or

significantly decrease – the noise problem. If noise is not an issue, that would open up

the possibility to heat supply air with the process water that is pumped up from the

mine.

41

It is highly plausible that the airflows will change with changes to the underground

ventilation system and the regulation. It is favorably for the payback period if both

flows are of the same size, but since this ventilation operates for the whole mine, the

return airflow must be lower to keep the ramp ventilated. If the flows are reduced, that

would extend the payback period for a heat exchanger system since the heating demand

would decrease – and therefore also the possible save. This reasoning is supported by

the finds in the study by Dello Sbarba et al. (2012).

The regulation of supply air temperature is done manually, and therefore there is no set

required supply air temperature. In this work, the calculation suggest that the heating

strategy can be compared to a required supply air temperature of 4 °C since the real

LPG-demand for the period February 2014 to January 2015 most closely correspond to

heating to 4 °C for that period. It is not an unreasonable temperature compared to

theory, but perhaps a couple degrees higher than what is usually practiced. But when it

is colder, the supply air is heated far higher than theoretically necessary. It is not fully

determined why it is practically necessary, there are several plausible reasons that likely

interact. It may be caused by a lot of free water in the supply airshaft. It might also be

caused by unreliable measurements in combination with insufficient mixing so a part of

the flow is heated unnecessary high while another part is so cold it freezes. The manual

regulation likely has an effect as well. It is probable that several actions will be taken

which directly or indirectly will lower or stabilize the supply air temperature. That

would lead to a longer payback period for heat exchanger systems of the same reason

reduced airflows would.

Comparing the results for the battery heat exchanger system from this work and the

results from the case closest to Kylylahti in the study by Dello Sbarba et al. (2012), the

payback period is shorter in the study than in this work. Comparing case 3 in this work

to the closest case in the study, the investment cost in the study was 130% of what was

used in this work, but the fuel price was almost the same, resulting in a payback period

of 6,3 years in this work, compared to roughly 5 years in the Dello Sbarba et al. study.

Comparing case 5 in this work to the one closest in the study, the investment cost in the

study was about 150% of what was used in this work and the fuel prices was of equal

size, resulting in a payback period of 3,4 years in this work compared to just below

3 years in the study. Generally the investment cost in the study was higher, most likely

due to the substantially longer distance (200m) between the shafts. A lower electricity

price and lower pressure drop contributes to a smaller increase in fan operation cost than

for Kylylahti. Since the return air temperature was much higher in the study (13 °C)

than in Kylylahti (~8,5 °C), a lot more energy can be saved with a heat exchanger,

although the slightly lower supply air temperature of 1,5 °C also means the possible

saving is smaller. That the payback period in both comparisons is lower in the study,

although the investment cost is higher, can be explained by the change in operation cost.

The savings in operation cost in the study is in both comparisons 140% of the savings in

this work.

Comparing the results from this work to the plate heat exchanger project in Zinkgruvan,

a plate heat exchanger in Kylylahti does not reach the same short payback period in any

tested case. The payback period in Zinkgruvan was 1,9 years, the best case scenario in

Kylylahti gives a payback period of 3,2 years. During the current circumstances (case 3)

42

the payback period would be 12,5 years, and with future circumstances (case 5) it would

be 6,4 years. Zinkgruvan has a higher average annual temperature than Kylylahti, which

would reasonably lower the demand for heating and therefore also the possible

savings, which would give a longer payback period compared to a colder location. But

there are many reasons a plate heat exchanger can’t be as successful in Kylylahti as in

Zinkgruvan. In Zinkgruvan, heating was done with oil before the heat exchanger was

installed, and oil is not direct acting. Kylylahti uses LPG which is direct acting, and

therefore the saving in fuel will not be as substantial as in Zinkgruvan. Since both flows

in Zinkgruvan are of the same size, more energy is available compared to the demand,

than if the return airflow is lower. That is why the wet temperature transfer efficiency is

higher for the system in Zinkgruvan (60%) than for the plate heat exchanger in

Kylylahti (55%). In addition to that, the return air temperature is a couple degrees

higher in Zinkgruvan (11 °C) than in Kylylahti (~8,5 °C), also making more energy

available.

Comparing the two heat exchangers, they both have pros and cons. The plate heat

exchanger has no moving parts and requires almost no maintenance. The battery heat

exchanger on the other hand has a pump that requires maintenance and the circulation

fluid needs to be refilled. In the plate heat exchanger system, a bypass function can be

built in to the burner building so the pressure drop on the supply air side can be

decreased when no heating is necessary. But there is no apparent possibility to bypass

the air on the return air side, meaning the pressure drop over the heat exchanger will be

the same all year round. In the battery heat exchanger system, a bypass function can

easily be built in to both the supply and the exhaust building. Doing so would reduce the

pressure drop during the time of the year when heating is not needed, and therefore

make the increase in operating cost lower and the pay-back period shorter. In the plate

heat exchanger, the heavy condensation is presumed to prevent any accumulation of

particles on the plates, which upholds the temperature transfer efficiency without

washing. In the battery heat exchanger the fins are corrugated to increase temperature

transfer efficiency, but it also increases the tendency for particle accumulation. To

reduce the tendency for accumulation, heat exchangers with bigger distance between the

fins were chosen for the return air side. The fins are also thicker, which means they can

withstand high pressure washing. But manual washing is impractical to do often and

accumulation of particles will reduce the heat transfer – and it is unknown exactly how

big the impact would be. A possible alternative would be an automated washing system

of the kind mentioned in the study by Dello Sbarba et al. (2012).

Both heat exchanger solutions will probably lower the noise level. The battery heat

exchanger system will bend the return airflow, preventing it from reaching so high

which currently is easing the spreading of the noise. While being led into the exhaust

building, the air velocity will also decrease significantly. The air velocity will also be

decreased within the plate heat exchanger and the noise reducing construction will

reduce the noise further. How big reduction these measures will have is unknown and

calculating noise reduction is beyond the scope of this project. But further work to

estimate the possible noise reduction is necessary before a heat exchanger could be

installed. If a heat exchanger is in place and the noise reduction is not sufficient, it will

be very difficult to take further noise-reducing actions.

43

To assess the change in carbon dioxide emissions, long term margin electricity is a less

appropriate method since it considers bigger changes in time, like new investments, in

how electricity is produced. Whether short term margin or the GHG protocol standard is

best depends on perspective. Regardless of which is used, the difference between the

two heat exchanger systems is small. With a short term margin electricity perspective

the decrease in carbon emissions compared to the current system is small, each system

gives the biggest decrease in about half the cases, and in case 5 the decrease is equal in

size. With the GHG protocol standard method the decrease compared to the current

system is significantly bigger, and the plate heat exchanger gives the biggest decrease in

emissions in all cases. Short term margin electricity is the method considered to most

fairly represent the real change in carbon emissions, but the GHG protocol standard is

the method used within the company for assessing changes, and therefore is the

perspective that should be used in this case as well.

Both tested heat exchanger solutions are profitable, but the battery system has a

considerably shorter payback period in all tested cases. The plate heat exchanger has a

higher wet temperature transfer efficiency, but not high enough to compensate for the

investment cost which is almost doubled the one for the battery heat exchanger system.

The only cases where the plate heat exchanger has a payback period noteworthy lower

than the anticipated life expectancy and the common “limits” is cases 1 and 2. Case 1 is

the best case scenario and case 2 is more reasonable but still fairly optimistic. The

payback period for the battery system on the other hand is in most cases considerably

lower than both the anticipated life expectancy and the common payback period. The

exceptions are cases 3 and 8. Case 3 is deemed most closely corresponding to present

circumstances and case 8 is the worst case scenario. The margin to life expectancy and

acceptable limits is still noteworthy even though individual variables are changed, with

the exception of the LPG-price. For a heat exchanger solution to be profitable within an

acceptable period, the LPG-price must increase. Even though case 3 with a payback

period of 6,25 years most closely correspond to the current situation, more plausible

cases such 4, 5 or 6 with payback periods of 3,66, 3,40 and 3,24 years should be

considered seriously. The price on LPG will likely increase and the supply air

temperature will probably be decreased. In case 5, an LPG-price of 0,7 €/kg was used, if

it would increase to 0,8 €/kg instead, the payback period would be 2,80 years. But

before any decision can be made, an estimation of how much a heat exchanger system

could lower the noise level is necessary and a comparison between a heat exchanger

system and only noise reducing actions must be made. With the changes the future

might hold, actions to reduce noise might no longer be necessary, but even without the

need for noise reduction, a heat recovery solution is still an interesting investment from

an energy saving perspective.

44

6. Conclusion

With the method Boliden uses for similar assessments, a plate heat exchanger system

gives the biggest reduction in carbon dioxide emissions compared to the current system.

The reduction is approximately 60% compared to 55% for a battery heat exchanger

system. A plate heat exchanger system also gives a bigger reduction in operating cost,

but the investment cost is so high – almost doubled the cost for the battery system – that

the payback period is never near as short as fort the battery system. The payback period

for a plate heat exchanger system is almost doubled the time for the battery heat

exchanger system, depending on what case is evaluated. With the current price on LPG,

the payback period is about 6,3 years for a battery system, but if the LPG price increases

to earlier levels the payback period is more in the range of 3,2-3,7 years. But other

energy saving measures in the ventilation system may easily mean that the payback

period becomes longer. Before a decision whether to invest in a heat exchanger system

or not, the noise reducing potential of the system should be evaluated.

45

7. Recommendation

If a heat exchanger is to be built in Kylylahti, a battery heat exchanger is preferred to a

plate heat exchanger, with regard to what system has the shortest payback period,

according to the calculations made in this work. But for a battery heat exchanger system

to become profitable within a reasonable timeframe, the LPG-price need to be at least

0,6 €/kg. Before a decision is made, an estimation of how much the noise level can be

reduced with installation of a heat exchanger needs to be made.

46

8. Further work

This work also raises a couple of questions that would be interesting to look further

into.

It is known which factors determine how far the supply air needs to be heated to avoid

icing in the supply air shaft, but how much influence they have in a particular system is

hard to estimate. A model to simulate energy transfer within a supply air shaft would

give a better understanding of specific factors influence but also provide a tool to easier

estimate what effect certain improvements may have.

How much the efficiency of a battery heat exchanger is affected by accumulation and

build-up of particles under these specific circumstances is unknown. A better

understanding of the effect could perhaps be achieved by analyzing already existing

similar systems.

Heating the supply air with process water is also an interesting possibility that arises if

the process water is redirected through one of the airshafts. How to enable maximum

heat recovery without freezing the process water is a problem that could be investigated

with dynamic modeling.

47

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Brake, D.J. (2013). Mine Ventilation: A practitioner’s manual. Brisbane: Mine

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Cengel, Y. & Boles, M. (2015). Thermodynamics: An engineering approach. New

York: McGraw Hill.

Cengel, Y. & Ghajar, A. (2011). Heat and mass transfer: Fundamentals and

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recovery systems for mine ventilation. International Journal of Mining, Reclamation

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Grubbström, R., & Lundquist, J. (2005). Investering och finansiering: Metodik och

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luftfororeninga/ [2015-03-30]

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SMHI (2014). Normal årsmedeltemperatur. [Electronic]. Available:

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%20du%20med%20SMHI%20Graddagar.pdf [2015-02-19]

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content/uploads/2014/07/Riktlinjer_Verkningsgrad_-_English.pdf [2015-02-20]

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Warfvinge, C. & Dahlblom, M. (2010). Projektering av VVS-installationer. Lund:

Studentlitteratur

Wikells byggberäkningar AB (2008). Wikells sektionsfakta – NYB 08/09. Växjö:

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Walther, Mats, at Zinkgruvan Mining AB (Zinkgruvan). Interview. (2015-02-06)

Appendix A 1(2)

Fan characteristics

Provided by Carlos Andrés at Zitrón

Appendix A 2(2)

Fan characteristics

Provided by Carlos Andrés at Zitrón

Appendix B 1(5)

Case variables

Variable Variance

Required supply air temperature 2 / 3 / 4 /6 [°C]

Return air temperature 7,5 / 8 / 8,5 / 9 / 9,5 [°C]

Relative humidity in return air 92 / 100 [%]

Lowest allowable exhaust air temperature 0,5 / 1,0 [°C]

Temperature transfer efficiency, (plate heat

exchanger)

55 / 60 [%]

Pressure drop, (plate heat exchanger) 400 / 450 [Pa]

Fan efficiency, (supply air fan) 69 / 71 / 73 [%]

Fan efficiency, (return air fan) 74 / 76 / 78 [%]

Price plate heat exchanger 8 600 000±100 000 [SEK]

Price battery heat exchanger à 375 000±50 000 [SEK] (supply

side)

à 375 000±50 000 [SEK] (return

side)

Price electricity 60 / 65 / 70 / 75 [€/MWh]

Price LPG 0,5 / 0,6 / 0,7 / 0,8 / 0,9 [€/kg]

Exchange rate 8,5 / 9,0 / 9,5 [SEK/€]

Case 1

Variable Chosen value

Required supply air temperature 6 °C

Return air temperature 9,5 °C

Relative humidity in return air 100 %

Lowest allowable exhaust air temperature 0,5 °C

Temperature transfer efficiency, (plate heat

exchanger)

60 %

Pressure drop, (plate heat exchanger) 400 Pa

Fan efficiency (supply air fan) 73 %

Fan efficiency (return air fan) 78 %

Price plate heat exchanger 8,5 MSEK

Price battery heat exchanger à 325 000 SEK (supply side)

à 325 000 SEK (return side)

Price electricity 60 €/MWh

Price LPG 0,9 €/kg

Exchange rate 9,5 SEK/€

Appendix B 2(5)

Case variables

Case 2

Variable Chosen value

Required supply air temperature 6 °C

Return air temperature 9 °C

Relative humidity in return air 100 %

Lowest allowable exhaust air temperature 0,5 °C

Temperature transfer efficiency, (plate heat

exchanger)

55 %

Pressure drop, (plate heat exchanger) 400 Pa

Fan efficiency (supply air fan) 73 %

Fan efficiency (return air fan) 78 %

Price plate heat exchanger 8,6 MSEK

Price battery heat exchanger à 375 000 SEK (supply side)

à 375 000 SEK (return side)

Price electricity 65 €/MWh

Price LPG 0,8 €/kg

Exchange rate 9,5 SEK/€

Case 3

Variable Chosen value

Required supply air temperature 4 °C

Return air temperature 8,5 °C

Relative humidity in return air 100 %

Lowest allowable exhaust air temperature 0,5 °C

Temperature transfer efficiency, (plate heat

exchanger)

55 %

Pressure drop, (plate heat exchanger) 400 Pa

Fan efficiency (supply air fan) 71 %

Fan efficiency (return air fan) 76 %

Price plate heat exchanger 8,7 MSEK

Price battery heat exchanger à 425 000 SEK (supply side)

à 425 000 SEK (return side)

Price electricity 65 €/MWh

Price LPG 0,5 €/kg

Exchange rate 9,5 SEK/€

Appendix B 3(5)

Case variables

Case 4

Variable Chosen value

Required supply air temperature 4 °C

Return air temperature 8,0 °C

Relative humidity in return air 100 %

Lowest allowable exhaust air temperature 0,5 °C

Temperature transfer efficiency, (plate heat

exchanger)

55 %

Pressure drop, (plate heat exchanger) 450 Pa

Fan efficiency (supply air fan) 69 %

Fan efficiency (return air fan) 74 %

Price plate heat exchanger 8,6 MSEK

Price battery heat exchanger à 375 000 SEK (supply side)

à 375 000 SEK (return side)

Price electricity 65 €/MWh

Price LPG 0,7 €/kg

Exchange rate 9,0 SEK/€

Case 5

Variable Chosen value

Required supply air temperature 3 °C

Return air temperature 8,5 °C

Relative humidity in return air 100 %

Lowest allowable exhaust air temperature 0,5 °C

Temperature transfer efficiency, (plate heat

exchanger)

55 %

Pressure drop, (plate heat exchanger) 400 Pa

Fan efficiency (supply air fan) 71 %

Fan efficiency (return air fan) 76 %

Price plate heat exchanger 8,6 MSEK

Price battery heat exchanger à 375 000 SEK (supply side)

à 375 000 SEK (return side)

Price electricity 65 €/MWh

Price LPG 0,7 €/kg

Exchange rate 9,0 SEK/€

Appendix B 4(5)

Case variables

Case 6

Variable Chosen value

Required supply air temperature 3 °C

Return air temperature 8,5 °C

Relative humidity in return air 100 %

Lowest allowable exhaust air temperature 0,5 °C

Temperature transfer efficiency, (plate heat

exchanger)

55 %

Pressure drop, (plate heat exchanger) 450 Pa

Fan efficiency (supply air fan) 71 %

Fan efficiency (return air fan) 76 %

Price plate heat exchanger 8,7 MSEK

Price battery heat exchanger à 425 000 SEK (supply side)

à 425 000 SEK (return side)

Price electricity 65 €/MWh

Price LPG 0,8 €/kg

Exchange rate 8,5 SEK/€

Case 7

Variable Chosen value

Required supply air temperature 2 °C

Return air temperature 8,5 °C

Relative humidity in return air 100 %

Lowest allowable exhaust air temperature 0,5 °C

Temperature transfer efficiency, (plate heat

exchanger)

55 %

Pressure drop, (plate heat exchanger) 450 Pa

Fan efficiency (supply air fan) 73 %

Fan efficiency (return air fan) 78 %

Price plate heat exchanger 8,5 MSEK

Price battery heat exchanger à 325 000 SEK (supply side)

à 325 000 SEK (return side)

Price electricity 70 €/MWh

Price LPG 0,6 €/kg

Exchange rate 8,5 SEK/€

Appendix B 5(5)

Case variables

Case 8

Variable Chosen value

Required supply air temperature 2 °C

Return air temperature 7,5 °C

Relative humidity in return air 92 %

Lowest allowable exhaust air temperature 1,0 °C

Temperature transfer efficiency, (plate heat

exchanger)

55 %

Pressure drop, (plate heat exchanger) 450 Pa

Fan efficiency (supply air fan) 69 %

Fan efficiency (return air fan) 74 %

Price plate heat exchanger 8,7 MSEK

Price battery heat exchanger à 425 000 SEK (supply side)

à 425 000 SEK (return side)

Price electricity 75 €/MWh

Price LPG 0,5 €/kg

Exchange rate 8,5 SEK/€

Appendix C

Fan speed – logged data

Provided by Eero Tommila at Boliden Kylylahti

0

200

400

600

800

1000

1200

RPM Supply fan rpm

Supply air fan

0

200

400

600

800

1000

1200

RPM Return fan rpm

Return air fan

Appendix D 1(6)

Product sheets for the battery heat exchanger

Provided by Percy Danielsson at Luvata

Appendix D 2(6)

Product sheets for the battery heat exchanger

Provided by Percy Danielsson at Luvata

Appendix D 3(6)

Product sheets for the battery heat exchanger

Provided by Percy Danielsson at Luvata

Appendix D 4(6)

Product sheets for the battery heat exchanger

Provided by Percy Danielsson at Luvata

Appendix D 5(6)

Product sheets for the battery heat exchanger

Provided by Percy Danielsson at Luvata

Appendix D 6(6)

Product sheets for the battery heat exchanger

Provided by Percy Danielsson at Luvata

Appendix E

Pump Characteristics

Provided by Thomas Nyström at Zander och Ingeström