heat recovery solutions for mine ventilation systems823877/fulltext01.pdfkarlstads universitet 651...
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Karlstads universitet 651 88 Karlstad
Tfn 054-700 10 00 Fax 054-700 14 60
[email protected] www.kau.se
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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Grubbström, R., & Lundquist, J. (2005). Investering och finansiering: Metodik och
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%20du%20med%20SMHI%20Graddagar.pdf [2015-02-19]
Svensk ventilation (2012). Temperature efficiency in heat recovery devices for supply
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Warfvinge, C. & Dahlblom, M. (2010). Projektering av VVS-installationer. Lund:
Studentlitteratur
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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