EPSRC THERMAL MANAGEMENT OF
INDUSTRIAL PROCESSES
Review of Industrial Condensing Boilers (Technology & Cost)
Case Study: Thermal Design of a condensing boiler in a Large
Scale Biomass District Heating Plant (40 MW)
(July 2010)
Report Prepared by:
SUWIC, Sheffield University
Researcher: Dr Q. Chen
Investigators:
Professor Jim Swithenbank
Professor Vida N Sharifi
Sheffield University Waste Incineration Centre
(SUWIC)
Department of Chemical and Process Engineering
Sheffield University
Executive Summary
A considerable amount of waste heat in boiler flue gases is in the form of latent heat
of water vapour. This energy cannot be recovered until the flue gases are cooled to a
temperature below the dew point. One of the main reasons why not much effort has
been spent by industry to recover latent heat from the flue gases is that it is difficult to
deal with the corrosion which arises when sulphuric acid or nitrate salts from flue
gases condense out at the surface of boilers and the flue ducts. Another reason is
that it is difficult to utilise the heat at low temperatures (i.e. approximately 150°C
which is the average flue gas temperature). Nevertheless, industrial condensing
boilers (condensers) are designed to recover this latent heat from the flue gases.
In accordance with our EPSRC grant proposal, Sheffield University has conducted
an extensive literature review of industrial condensing boilers, looking into various
technologies and the associated costs. In addition, extensive calculations have been
carried out as part of a case study to investigate the thermal design of a condensing
boiler in a large scale district heating plant (40 MW). This report presents the results
obtained from the above studies.
Some main findings from this study are as follows:
1.
By recovering the latent heat of water vapour in the flue gas through
condensing boilers, the whole heating system can achieve significantly higher
efficiency levels than using conventional boilers.
2.
In addition to waste heat recovery, condensing boilers can also be optimised for
emission abatement, especially for particle removal. The particle collection
mechanisms include inertial impaction, gravitational settling (for larger
particles), thermophoresis (induced by the temperature gradient between the
flue gas and the cool surface), and diffusiophoresis (by the steam condensation
on cool surfaces). Moreover, particle growth by water condensation also
changes the particle size distribution in the gas stream thus aiding collection.
3.
Two potential technical barriers for the condensing boiler application are
corrosion and return water temperatures. a) Highly corrosion-resistant
material is required for condensing boiler manufacture. b) In order to reduce
the return water temperature, an under-floor heating system or a high surface
area of the radiators is needed in combination with a condensing boiler in the
heating system.
4.
The thermal design of a single pass shell-and-tube condensing heat
exchanger/condenser shows that there is considerable thermal resistance on the
shell-side. This is due to; fouling, gas phase convective resistance and vapour
film interface resistance. For the ‘Case Study’ model boiler, approximately
4919m2 of total heat transfer area is required if stainless steel is used as a
construction material. If the heat transfer area is made of carbon steel, then
polypropylene could be used as the corrosion-resistant coating material outside
I
the tubes. The addition of a polypropylene coating increases the tube wall
thermal resistance; hence the required heat transfer area would be
approximately 5812m2.
5.
The estimated total capital cost for the condensing boiler design ranges from
$2,028,000 (carbon steel) to $4,908,000 (stainless steel). The application of
the condensing boiler increases the energy efficiency, leading to fuel savings of
up to 20%. The payback period is about 2 years for the carbon steel
condenser or 4 years for the stainless steel condenser.
6.
The condensing boiler requires a lower water return temperature and should be
used in conjunction with a heat pump or with an under-floor system or larger
radiators for building heating.
II
List of Contents
1. Introduction ............................................................................................................1
2. Literature Review: Condensing Boiler and its Application....................................3
2.1 Condensing Boilers.......................................................................................3
2.1.1 Modes of Condensation .....................................................................3
2.1.2 Gas-fired Condensing Boilers ...........................................................5
2.1.3 Flue Gas Condensers .........................................................................7
2.1.4 Advantages of Condensing Boilers..................................................10
2.1.5 Technical barriers.............................................................................13
2.1.6 Potential solutions............................................................................17
2.2 Heat Pumps.................................................................................................20
2.2.1 Compression heat pump ..................................................................20
2.2.2 Absorption heat pump......................................................................21
2.2.3 Application of heat pumps with a condensing boiler ......................23
2.3 Cost and Economic Issues ..........................................................................25
2.4 Application of Condensing Boilers and Heat Pumps in Heating Systems .29
2.4.1 Sodra Nas Vimmerby Energi AB Biomass District Heating Plant,
Sweden.........................................................................................................29
2.4.2 Kraftvarmeværk Waste Incineration Plant in Thisted Denmark......31
2.4.3 The Hedenverket Waste-to-Energy Plant at Karlstad, Sweden .......32
2.4.4 Davamyran Heat and Power Plant...................................................33
2.4.5 The Vestforbranding Waste to Energy Plant in Copenhagen,
Denmark.......................................................................................................34
2.4.6 Sonderborg Waste to Energy Plant, Denmark .................................36
3. Case Study: Condensing Boiler Design for a Biomass Fuelled Heating Plant ....37
3.1 Plant Description ........................................................................................37
3.2 Conditions of the Heating Plant..................................................................38
3.2.1 Fuel input .........................................................................................38
3.2.2 Process Parameters ..........................................................................39
3.2.3 Flue Gas Composition .....................................................................40
3.3 Condensing Boiler Design..........................................................................41
3.3.1 Heat Exchanger Selection................................................................41
3.3.2 Condensation Curve ........................................................................45
III
3.3.3 Thermal Design Methodology.........................................................47
3.3.4 Heat Transfer Coefficients...............................................................49
3.4 Size of the Condenser and the Pressure Drops ...........................................59
3.5 Cost Estimation...........................................................................................61
3.5.1 Capital Costs....................................................................................62
3.5.2 Operating and Maintenance Costs ...................................................65
3.5.3 Profitability......................................................................................65
4. Conclusions ..........................................................................................................67
References ................................................................................................................69
IV
1. Introduction
The flue gases from incineration plants normally carry off 15–40% of the heat
content of the fuel. Thus, perfect cooling of the flue gases either increase the
capacity of the incineration plant by 18–67% for the same fuel consumption or reduce
the fuel consumption by 15–40% for the same supplied power (Fagersta Energetics,
2009). The waste heat in flue gases can be extracted economically with the possible
exception of economic heat saving by additional insulation. One reason why little
effort has been made to recover the heat from flue gases is that it has not previously
been possible to counteract the corrosion which arises when sulphuric acid or nitrate
salts from flue gases condense in boilers and flue ducts. Another reason is that it can
be difficult to utilise the heat at the low temperatures which are normal for flue gases.
A considerable amount of the waste heat in flue gases is in the form of latent heat of
water vapour in these gases. This energy cannot be recovered until the flue gases are
cooled to temperatures under the dew point. Condensing boilers are designed to
recover the latent heat of water vapour in the flue gases.
Since the 1970s, condensing boilers have been developed and have found wide
applications in most countries (Comakli, 2008). The condensing boiler is a very
competitive technology in Europe due to somewhat higher energy prices, stricter
government regulations and a more favorable market interest in energy efficiency.
In addition, because central air conditioning is not generally provided in buildings in
Europe, the condensing boiler has only to compete with conventional boilers, which it
does successfully due to its low operating costs. Due to the attractiveness of this
technology, condensing boilers are very common in Europe (Figure 1), and even make
up over half of the total market for boilers in Holland (CEE 2001).
Figure 1
Market share of condensing boilers in annual residential gas boiler sales
(Weber et al, 2002)
Gas condensing boilers are a particularly suitable technology to increase energy
1
efficiency and to generate environmental benefits. It is one of the single end-use
technologies which offers the most important energy saving and emission reduction
potential. About 5% of the total energy use for residential space heating in the
European Union and 4% of the corresponding emissions could be saved by the
implementation of condensing boilers instead of improved efficiency boilers (Weber
et al, 2002).
In the literature, various schemes for reclaiming the latent heat in flue gas have so
far been put forward and general methods for designing condensing heat exchangers
have been proposed.
2
2. Literature Review: Condensing Boiler and its Application
2.1 Condensing Boilers
Combustion of hydrocarbon-rich fuels, such as natural gas, oil, coal and biomass, in
air yields two primary products, carbon dioxide and water vapor, entrained in the
relatively inert nitrogen of the air. Conventional boilers transfer most of the sensible
heat of this reaction to water as hot water or steam. Condensing boilers are designed
to capture a fraction of the latent heat, i.e., the energy released by condensing water
vapour in the flue gas. By extracting this latent heat in the condensing boiler, the
whole system can achieve higher efficiency levels. To capture this energy, the flue
gas requires a heat sink that is cool enough to allow water condensation. For heating
boilers that use the returning water from the system as a heat sink, this requires return
water temperature below 60°C.
2.1.1 Modes of Condensation
When condensation of the moisture in the flue gas occurs in a condensing boiler, the
condensate can accumulate on the cold surface in one of two ways (Marto, 1991). If
the water wets the cold surface, the condensate will form a continuous film (filmwise
condensation). If the water does not wet the surface, it will form into numerous
microscopic droplets (dropwise condensation). Dropwise condensation leads to
much lower thermal resistances to heat transfer than filmwise condensation.
However, long-term dropwise condensation conditions are very difficult to sustain.
Filmwise condensation is normally encountered in industrial applications and
dropwise condensation can only be maintained under controlled conditions with
special surface coatings or additives to the vapour. All surface condensers today are
designed to operate in the filmwise mode.
The cooled surface in condensing boilers may be of any orientation, though vertical
and near-horizontal geometries are preferred (Chisholm, 1980).
In vertical
condensing boilers, a film of condensate (condensed from flue gases) will fall under
the influence of gravity thickening downstream with increasing load. Film flow will
be laminar near the top of the tube but may become turbulent at high loadings (Figure
2 (a)). The film surface is usually covered with ripples or waves which influence the
condensation process. The influence of the gas flow is important. If the gas flow is
concurrent with condensate, the surface shear causes film thinning. On the other
hand a small counter-current flow will cause film thickening (Figure 2 (b) (c)) while a
larger flow may cause flooding and eventually flow reversal (Figure 2 (d) (e)). A
flow of gas across tubes will lead to a non-axisymmetric liquid distribution with
thinning of the film on the upstream side and thickening in the wake.
The heat transfer resistance of the condensate film is directly proportional to its
3
thickness and is reduced by wave effects and turbulence. Cocurrent and cross
gaseous flows cause film thinning and reduce heat transfer resistance.
In horizontal condensing boilers with condensation outside tubes, the gas stream
may flow vertically upwards or downwards across the tubes, or horizontally in a
direction parallel to or perpendicular to the tubes. The boilers are usually vertically
baffled to produce a combination of horizontal cross flow and parallel flow. In
perfectly horizontal condensers, the condensate drips vertically from the uppermost
tubes leading to thicker films and higher heat transfer resistance of the lower tube
rows. This phenomenon is known as condensate inundation. If the tubes are
inclined at even modest angles to the horizontal, the liquid flows in the direction of
the slope, at least as far as any vertical baffle.
Figure 2
Film condensation on vertical surfaces
During film condensation in tube bundles, the conditions are significantly different
from a single tube (Kakac and Liu, 2002). The presence of neighbouring tubes
creates several added complexities, as shown in Figure 3. In the idealised case
(Figure 3a), the condensate from a given tube is assumed to drain by gravity to the
lower tubes in a continuous, laminar sheet. Actually, the condensate from one tube
may not fall on the tube directly below it. The inundation largely depends on the
spacing-to-diameter ratio of the tubes and on whether the tubes are arranged in a
staggered or in-line configuration. As shown in Figure 3b, the condensate may flow
sideways down the tube bundles. Experiments have shown that condensate does not
drain from a horizontal tube in a continuous sheet but in discrete droplets along the
tube axis. When these droplets strike the lower tube, considerable splashing can
4
occur (Figure 3c), causing ripples and turbulence in the condensate film. Moreover,
large gas velocity can also create significant shear forces on the condensate, stripping
it away from the film (Figure 3d).
Figure 3
Condensate flow in condensing boilers (Kakac and Liu, 2002)
2.1.2 Gas-fired Condensing Boilers
Typically, non-condensing boilers have atmospheric burners, cast iron heat
exchangers and metal or masonry chimneys (CEE, 2001). The products of
combustion (flue gases) are maintained at a sufficiently high temperature (resulting in
low heat transfer efficiency) to allow them to exit the system using natural convection.
If the flue gases do not contain enough heat to maintain proper stack buoyancy, the
combustion products will spill back into the building. In addition, if the internal flue
surface temperature is allowed to drop below the dew point, moisture in the
combustion products will condense on the internal walls of the heat exchanger and
flues. As the condensate is very acidic, it will corrode the heat exchanger walls and
damage metal and masonry chimneys. By not capturing any latent heat from flue
gases, non-condensing boilers operate at low efficiency. However, due to their
relatively low cost of fabrication, they dominate the market, and can use either natural
gas or distillate for fuel.
As the temperature of the flue gas at the exit of a conventional gas fired boiler is
usually high, a great amount of heat energy is lost to the environment. In the flue
gas, both sensible heat and latent heat can be recovered by adding a condensing heat
exchanger. Thus, the condensing boiler efficiency can be increased by as much as
10% (Comakli, 2008).
5
Figure 4
Schematic arrangement of a condensing boiler for house heating
Condensing boilers run at a positive pressure with forced-draft power burners or
pulse combustion instead of atmospheric draft to pull gases through the firebox and
heat exchanger.
These boilers are equipped with stainless steel or other
corrosion-resistant material since they are designed to tolerate the transient presence
of condensate in the boiler. Condensing boilers operate at high efficiency by
capturing some of the latent heat and most of the sensible heat of combustion. In
addition, these boilers operate at high efficiency even at part-load conditions when
return water temperatures from space heating equipment are low. Because of the
relatively low flue gas temperatures, condensing boilers require flue construction that
accommodates condensation downstream of the boiler (CEE, 2001).
The development of the condensation technique for heating applications presents
major opportunities in decreasing gas consumption in apartment houses, independent
houses, commercial building and official buildings. Figure 4 presents a schematic
arrangement of a condensing boiler for house heating. For condensing boilers, the
boiler efficiency can reach a theoretical maximum value over 110% based on the
lower heating value of fuels (Comakli, 2008). For natural gas, the boiler efficiency
is dependent upon the flue gas temperature, with air/fuel ratio λ as a parameter shown
in Figure 5.
6
Figure 5
Efficiency of a condensing boiler versus exit flue gas temperature under
different access air ratios
2.1.3 Flue Gas Condensers
The application of flue gas condensers to recover waste (latent) heat from flue gases
is much wider than the stand-alone gas-fired condensing boilers. Flue gas
condensers can be designed not only for power plants, but also for all commercial and
industrial facilities as well. Energy recovered by the flue gas condensers can be used
in district heating and cooling schemes or put back into an industrial process.
Moreover, with flue gas condensers, a large amount of water can also be recovered
from flue gases that otherwise is exhausted to the atmosphere.
There are two types of flue gas condensers developed for industrial application:
indirect and direct contact condensers, as shown in Figures 6 and 7.
An indirect contact condenser removes heat from hot flue gases by passing them
through one or more shell-and-tube or tubular heat exchangers (US DOE, 2007).
This condenser can heat fluids to a temperature of 90°C while achieving exit gas
temperatures as low as 25°C (depending on the temperature of the cooling fluid).
The indirect contact condenser is able to preheat water to a higher outlet or process
supply temperature than the direct contact condenser. However, the condenser must
be designed to withstand corrosion from condensed water vapour. The condensed
water is acidic and must be neutralized if it is to be discharged into the sewer system
or used as process water.
The indirect contact condensers can be further categorised into three types: pipe
condenser, lamella condenser, and combi condenser (Nederhoff, 2003). In a pipe
condenser, the flue gases flow through pipes that are surrounded by cold water. The
water flows along the pipes but in opposite direction of the gas. In this system, the
7
temperature of the flue gases can get as low as the temperature of the incoming water.
If the water is cold enough, water vapour in the flue gases condenses on the walls
inside the pipes (Nederhoff, 2003).
Figure 6
Indirect contact flue gas condenser (DOE 2007)
In the lamella condensers, the pipes contain cold water and are surrounded by flue
gases. The pipes have aluminium lamellas (fins) attached to them to enlarge the
contact surface with the gases. The flue gases are blown over and across the cold
pipes and lamellas. The lamellas are not as cold as the pipes and the gases cannot be
cooled as low as the temperature of the incoming water. This makes a lamella
condenser less effective than the pipe condenser.
Figure 7
Direct contact flue gas condenser (DOE 2007)
8
A combi condenser consists of two condensers in one system: one condenser cools
the flue gases down to 70°C, and the second takes care of further cooling from 70°C
down to 40°C. The second condenser takes the condensation energy out of the flue
gases. Generally, both condensers are lamella types. The first of the two
condensers operates at a high temperature level, and delivers the heat to the return
water of the normal pipe heating system. The second condenser operates at a much
lower temperature level. This condenser is usually connected to a separate heating
net that runs at a lower temperature. It is also possible that both are connected to the
normal heating system. In this case, the second condenser pre-heats the cold water,
and the first condenser then heats the water further to the required high temperature.
A combi condenser retrieves nearly all energy that is present in the flue gases, and
therefore achieves very high energy efficiency, but the investment costs are therefore
higher than for a single condenser.
Another heat recovery option is to use a direct contact condenser (Figure 7), which
consists of a vapor-conditioning chamber followed by a countercurrent spray chamber.
In the spray chamber, small droplets of cool liquid come into direct contact with the
hot flue gas, providing a non-fouling heat transfer surface. The liquid droplets cool
the stack gas, condense and remove the water vapour. The spray chamber may be
equipped with packing to improve contact between the water spray and hot gas. A
mist eliminator is required to prevent carryover of small droplets. The direct contact
design offers high heat transfer coupled with water recovery capability since heated
water can be collected for boiler feed water, space heating, or plant process needs.
Recovered water will be acidic and may require treatment prior to use, including
membrane technology, external heat exchangers, or pH control. Direct contact
condensers operate close to atmospheric pressure; altitude and flue gas temperature
limit the makeup water temperature to 40 to 60°C.
Condensers require site-specific engineering design and a thorough understanding
of the effects of their operation on the existing steam system and water chemistry.
If the pressure of the flue gases is increased, the dew point rises and there is greater
potential for abstracting heat by means of condensation. If the system is pressurized
to a pressure of 4bar, the condensation of moisture in the flue gases may occur at
temperatures between 60°C and 115°C. For example, Fagersta Energetic AB, a
Swedish company, developed an advanced project called the Bioturbo system, in
which wet peat was burned in a pressurised fluidised bed, as shown Figure 8. This
3MW pilot plant was tested for over 1000 hours and operated successfully with peat
containing water up to 78%. Very high overall efficiency was achieved during the
operation. Flue gas temperatures were between 10°C and 20°C, and emissions were
low (Fagersta Energetic, 2009a).
9
Figure 8
The Bioturbo system: a power plant with pressurised combustion system
2.1.4 Advantages of Condensing Boilers
Latent Heat Recovery
120
115
λ=1.25
λ=1.6
λ=2.0
λ=2.4
Efficiency, %
110
105
100
95
90
85
80
30
50
70
90
110
130
150
Flue gas temperature, oC
Figure 9
Theoretical efficiency of a wood chip boiler with a condenser versus exit
flue gas temperature under different excess air ratios
The most significant advantage of a condensing boiler is that the latent heat of water
vapour can be recovered from the flue gas. This greatly improves the overall
thermal efficiency of the system. Figure 9 shows the variations of the theoretical
thermal efficiency (with reference to net calorific value) of a wood chip boiler with a
condensing heat exchanger against exit flue gas temperatures under different excess
air ratios. The fuel (wood chips) used in the plant has 50% moisture content, 25.6%
C, 3.05% H, 20.45% O. With the increasing excess air ratio, the partial pressure of
the water vapour in the flue gas decreases. This lowers the dew point of the flue gas.
10
Meanwhile, more sensible heat is carried by non-condensable gas components under
higher excess air ratio conditions. Consequently, at the same exit temperature of the
flue gas, higher excess air ratio leads to lower thermal efficiency.
As shown in Figure 9, the condensing heat exchanger/condenser recovers the latent
heat of moisture when it is condensed. The recovery of the latent heat results in the
thermal efficiency exceeding 100% with reference to the lower heating value of the
input fuel (Neuenschwander et al. 1998).
Emission Abatement
Wood combustion generates fine particles, which contribute significantly to the
emissions from the energy sector (Grohn et al. 2009). The common aerosol size
distribution from wood combustion peaks at 50-400 nm with relatively high number
concentrations (Sipula et al. 2009). Condensing heat exchangers (condensers) can
be optimized for simultaneous particle collection and waste heat recovery. The
condensate forms a constant water film that can carry away any deposited particles.
Sipula et al. (2009) studied particle and gaseous emissions of four different wood
chip-fired district heating units in the size range of 5-15 MW. All of the units were
equipped with cyclones to remove coarse particles from the flue gas. In addition,
two of the rotating grate boilers were equipped with single field electrostatic
precipitators (ESP), and one with a condensing flue gas scrubber (as shown in Figure
10). It was found that the condensing flue gas scrubber removed on average 44% of
PM1 and 84% of total solid particles (TSP).
Figure 10
Schematic diagram of the condensing scrubber (Sipula et al. 2009)
11
Figure 11 Particle mass size distributions before and after the flue gas scrubber
Figure 11 shows the particle mass size distribution before and after the flue gas
scrubber. As shown, particles with diameter smaller than 300nm and between 1.0
and10µm were partially removed in the condensing scrubber. The average 44%
decrease in fine particles, which were clearly below 500 nm in size, probably resulted
from a combination of thermophoresis due to cool surfaces, and diffusiophoresis due
to steam condensation. In addition, the particle sizes were found to grow inside the
scrubber, as seen in the shift of the particle size distribution.
Recently, a commercially available wet scrubbing process called “FLUE-ACE” has
been developed (Keeth et al, 2005). It consists of a condensing reactive scrubber
that can be used for heat recovery and emissions control. The scrubber operates by
cooling flue gas substantially below the dew point temperature, thus forcing the
condensation of water vapor and other condensables. This results in greater removal
of condensables and fine particulates than can be achieved in a conventional wet
scrubber. The FLUE-ACE wet scrubber has been demonstrated to remove 96-99%
of flue gas SO2, NO2 and HCl. The High Performance (HP) FLUE-ACE model
additionally removes greater than 98% of SO3 mist and fine particulates greater than
0.3µm in diameter. Due to the condensing action used for pollutant removal, it is
expected that mercury removal in the scrubber can be greater than in a conventional
wet scrubber.
There are currently 13 commercial installations of the FLUE-ACE technology
operating in Canada, all installed in the past 16 years. The majority of these
installations provide acid gas control for smelting operations or paper mills, with the
largest operating commercial installation treating a 75MW equivalent stream of gas
(Keeth et al, 2005).
In addition to condensing scrubbers, condensing heat exchangers can also be used
for wet scrubbing. Keeth et al. (2005) reported a pilot process consisting of a
condensing heat exchanger with an FGD system. The condensing heat exchange
cooled the flue gas to 20 – 30°C. Large water droplets were formed around the
12
pollutants and then removed in the FGD system. The pilot process was tested in
1994 – 1995. Through this system, 58% Hg, 80–90% of PM and 95% SO2 were
removed.
In condensers, the particle separation mechanisms include inertial impaction and
gravitational settling for larger particles and diffusion for the smallest particles. In
addition to Brownian diffusion, important factors in the removal of fine particles
include thermophoresis, induced by the temperature gradient between the flue gas and
the cool surface, and diffusiophoresis, caused by the steam condensation on cool
surfaces. Furthermore, in condensing scrubbers/heat exchangers, particle growth by
water condensation can affect particle size distributions in the emission (Sipula et al.
2009).
2.1.5 Technical barriers
Corrosion
Because the products of combustion include materials that are highly corrosive,
corrosion arising from condensing gases has been a problem in industry for many
years (Huijbregts and Leferink. 2004). Corrosion-derived cracks were mostly found
in the low temperature heat exchangers (typically operating at temperatures between
70 and 90°C). In general, the exchangers had been fabricated from steel St35.8, a
standard low-carbon steel for construction purposes. Most cracking occurred where
mechanical stresses were relatively high. Microscopic analysis of samples revealed
that inter granular corrosion had occurred and it was frequently reported that complete
grains of material had become detached. In the case shown in Figure 12, nitrate
stress corrosion cracking was identified as the cause of the failures.
To avoid corrosion due to condensing gases, it is of vital importance to well
understand the composition and amount of condensed liquid that could be formed in
the condensing boilers. Descriptions and calculation methods of condensation have
been extensively studied during the past few decades (Kiang, 1981; Huijbregts and
Leferink. 2004). In clean air, the dew point can be directly obtained from the water
vapour pressure table, as shown in Figure 13.
When other gaseous species are present, such as SO3, SO2, HCl or NO2 in particular,
the dew point will deviate from the ideal dew point line (Figure 13). Under the
atmospheric pressure, dew point of the flue gas in the presence of these species can be
calculated by means of the equations listed in Table 1 (Huijbregts and Leferink.
2004).
13
Figure 12
A typical micrograph of stress corrosion cracking on the cross section of
tube material
Figures 14 – 17 show the examples of calculated dew points for gases with SO3,
SO2, HCl and NO2 respectively. In the cases of very low HCl and NO2 levels, the
calculated dew points are lower than the water dew point. This practically does not
occur, and the water dew point should be preferred. As can be seen in Figures 16
and 17, straight water dew point lines are used for low HCl and NO2 levels.
Table 1 Equations for dew point calculation
Species Dew point, Tdew
SO3
1000
(
2.276 − 0.0294 ln PH 2O − 0.0858ln PSO3 + 0.0062 ln PH 2O × PSO3
SO2
P in
atm
)
mmHg
1000
(
3.9526 − 0.1863ln PH 2O + 0.000867 ln PSO2 − 0.00091ln PH 2O × PSO2
HCl
mmHg
1000
(
3.7368 − 0.1591ln PH 2O − 0.0326 ln PHCl + 0.00269 ln PH 2O × PHCl
NO2
)
1000
− 273
V % H 2O
VppmNO2
V % H 2O VppmNO2
3.664 − 0.1446 ln
− 0.0827 ln
+
0.00756
ln
ln
100 × 760
760 ×106
100 × 760 760 ×106
14
)
Figure 13
Dew point of clean air versus water vapour pressure
Figure 14
Dew point of the flue gas versus SO3 content under different water vapour
pressures
Figure 15
Dew point of the flue gas versus SO2 content under different water vapour
pressures
15
Figure 16
Dew point of the flue gas versus HCl content under different water
vapour pressures
Figure 17
Dew point of the flue gas versus NO2 content under different water
vapour pressures
Return water temperature
As described above, heat sinks are required in condensing boilers and flue gas
condensers to capture the latent heat of condensation of water vapour in the exhaust
stream. Generally, return water from heating systems serves as a heat sink. Thus,
the return water temperature is the critical factor to the operation of a condensing
boiler. The return water temperature determines whether the boiler operates in
condensing mode (CEE 2001). Meanwhile, the efficiency of a condensing boiler
depends largely on the return water temperature. At higher temperatures, less water
vapour is condensed from the flue gas and the efficiency is decreased. This situation
is often encountered in practical applications (Doherty et al. 2006).
Technical requirements limit the suitability of condensing boilers in many
16
commercial applications. The need for low return water temperatures and 2-pipe
(minimum) hydronic distribution systems severely limits the penetration of
condensing boilers into the large retrofit market. Competitive alternatives such as
unitary roof-top packaged air conditioning and heating units and combination space
conditioning-water heating systems severely limit the applicability of condensing
boilers in all market segments. These alternatives not only provide zoning and
reasonably precise temperature control but also allow for individual billing of energy
costs. (CEE 2001)
2.1.6 Potential solutions
Corrosion-resistant materials
To capture as much latent heat as possible, and because the products of combustion
include materials that are highly corrosive, condensing boilers require specialized
materials for fabrication. To withstand these corrosive conditions, condensing
boilers are made of stainless steel and other corrosion resistant (and sometimes costly)
materials. They can require more sophisticated controls, and more careful
installation, to achieve their potential. In addition, the terminal units (radiators,
convectors, and fan-coils) connected to the condensing boiler tend to be more
expensive due to the greater heat exchanger surface required to operate at lower water
temperatures.
Condensing boilers thus require specialized corrosive-resistant
materials and sophisticated controls resulting in installed costs that are up to 3 times
higher than that for a conventional boiler (CEE 2001).
Table 2 High performance stainless steel materials
Density Conductivity Thermal Expansion
Alloy Type
Alloy
UNS No.
lb/in3
Btu/hr·ft·F in/in ×10-6/F
Sea-Cure®
S44660 0.28
9.5
5.4
27-29% Cr
®
AL29-4C
S44735 0.28
9.5
5.2
Ferritic
FS 10
S44800 0.28
9.5
5.4
®
25% Cr Duplex SAF2507
S32750 0.28
8.2
7.2
®
AL6XN
N08367 0.29
7.9
8.5
6% Mo Austenitic
®
254SMO
S31254 0.29
7.5
8.9
®
7% Mo Austenitic 654SMO
S32654 0.29
7.5
8.5
One of the typical materials used for condensers are high performance Stainless
Steel materials, which are characterised by high chromium contents together with
molybdenum and nitrogen. They include both austenitic and ferritic material.
They were developed by companies in the US, Europe and Japan. The properties of
some of the materials are listed in Table 2. These are seawater corrosion resistant
17
materials. Of all these properties, the most important is the thermal conductivity.
The thermal conductivity affects the heat transfer capability of these alloys. The
higher the thermal conductivity the higher the heat transfer capability. As shown in
Table 2, the ferritic stainless steels have higher thermal conductivity than the
austenitic alloys (Burns and Tsou, 2009).
Reducing the return water temperature: large radiators and under-floor heaters
In a local/district heating system, the return water serves as a heat sink in a
condensing boiler. As stated previously, the temperature of the return water is the
critical factor to the operation of a condensing boiler. The condensation of moisture
from the flue gases requires that the gases be cooled below their dew point. In the
case of natural gas combustion products at atmospheric pressure, this temperature is
about 55°C to 65°C and it follows that the temperature of the water returned from a
central heating system, (or from the hot water heating system), must be about 30°C,
i.e. well below this temperature. This requires an under-floor heating system or a
high surface area of the radiators in the building. Similar considerations apply to
district heating schemes if the latent heat of the moisture is to be recovered from the
flue gases. Most district heating schemes in the UK presently use delivery and
return water temperatures of about 120°C and 70°C respectively, although some
Scandinavian district heating schemes do utilize the latent heat (Paappanen and
Leinonen, 2005).
Figure 18
Supply water temperatures and outdoor air temperatures
In this section, a Finish heating scheme is introduced to illustrate the application of
radiators and floor heaters. The space heating load is decreasing in modern Finnish
apartments due to lower U-values of the construction, tight envelopes and heat
recovery from exhaust ventilation air. This makes it possible to develop a new
combined low temperature water heating system with nominal supply/return water
18
temperatures of 45°C/35°C. Such a system includes radiators in rooms and floor
heating in bathrooms (Hasan et al, 2009).
A common heating system in Finnish apartment buildings is a water radiator system
that operates by district heating. The supply and return water temperatures are 70
and 40°C, respectively, at an outdoor air temperature of -26°C, which are the design
temperatures for the southern parts of Finland. The supply water temperature is
outdoor air temperature compensated, i.e., the supply water temperature increases as
the outdoor air temperature decreases, as shown in Figure 18. Typically, in modern
Finnish buildings, floor heating is expected in bathrooms and toilets. As the supply
water temperature is too high for direct use in floor heating, a secondary circuit with a
mixing valve is one design used to lower the operating temperature. Another option
is to use electric floor heating. Major disadvantages of this latter method are its high
consumption of primary energy and its ON/OFF switching.
Figure 19
Low temperature water heating system with radiators and floor heating
The decreasing load for space heating in Finland has led to the development of a
new combined low temperature water heating system that includes radiators in rooms
and floor heating in bathrooms. The nominal supply and return water temperatures
for such a system are 45 and 35°C, respectively, at an outdoor air temperature of
-26°C for the southern parts of Finland. This new system is simple, easy to install
and expected to perform well compared with conventional systems. The application
of such a system is mainly in apartment buildings, but it would also be possible in
detached houses. This system can include the air handling unit heating coils as well.
This system can be connected to low temperature heat production units, e.g. heat
pumps, or conventional high temperature systems, e.g. district heating. The basic
arrangement of the system is presented in Figure 19 (Hasan et al, 2009).
19
2.2 Heat Pumps
A heat pump is a machine or device that moves heat from one site (the “source”) to
another (the “sink” or “heat sink”) using mechanical work. Most heat pump
technology moves heat from a low temperature heat source to a higher temperature
heat sink. Heat pumps can be regarded as a heat engine which is operating in
reverse and can be categorised as two main types: compression heat pumps and
absorption heat pumps. Compression heat pumps always operate on mechanical
energy (using electricity), while absorption heat pumps may also run on heat as an
energy source (Heat Pump Centre, 2009).
2.2.1 Compression heat pump
The great majority of heat pumps work on the principle of the vapour compression
cycle. The main components in such a heat pump system are the compressor, the
expansion valve and two heat exchangers referred to as the evaporator and condenser.
The components are connected to form a closed circuit, as shown in Figure 20. A
volatile liquid, known as the working fluid or refrigerant, circulates through the four
components.
Figure 20
Compression heat pump (electricity driven)
In the evaporator, the temperature of the liquid working fluid is kept lower than the
temperature of the heat source (the return water to the condensing boiler), causing
heat to flow from the heat source to the liquid, and the working fluid evaporates.
Vapour from the evaporator is compressed to a higher pressure and temperature. The
hot vapour then enters the condenser, where it condenses and gives off useful heat.
Finally, the high-pressure working fluid is expanded to the evaporator pressure and
temperature in the expansion valve. The working fluid is returned to its original
state and once again enters the evaporator.
20
The compressor is usually driven by an electric motor and sometimes by a
combustion engine. Electric motors drive the compressor with very low energy
losses. The overall energy efficiency of the heat pump strongly depends on the
efficiency by which the electricity is generated. When the compressor is driven by a
gas or diesel engine (Figure 21), heat from the cooling water and exhaust gas is used
in addition to the condenser heat.
Figure 21
Compression heat pump (engine driven)
Industrial vapour compression heat pumps often use the process fluid itself as
working fluid in an open or semi-open cycle. These heat pumps are generally
referred to as mechanical vapour re-compressors, or MVRs. Generally, MVRs can
be classified as open and semi-open heat pumps (Heat Pump Centre, 2009).
In open systems, vapour from an industrial process is compressed to a higher
pressure and thus a higher temperature, and condensed in the same process giving off
heat. In semi-open systems, heat from the recompressed vapour is transferred to the
process via a heat exchanger. Because one or two heat exchangers are eliminated
(evaporator and/or condenser) and the temperature lift is generally small, the
performance of MVR systems is high, with typical coefficients of performance (COP:
the ratio of heat delivered by the heat pump and the electricity supplied to the
compressor) of 10 to 30. Current MVR systems work with heat-source temperatures
from 70-80ºC, and deliver heat between 110 and 150ºC, in some cases up to 200ºC.
Water is the most common “working fluid” (i.e. recompressed process vapour),
although other process vapours are also used, notably in the (petro-) chemical
industry.
2.2.2 Absorption heat pump
Absorption heat pumps are thermally driven by heat rather than mechanical energy.
21
Absorption heat pumps for space conditioning are often gas-fired, while industrial
installations are usually driven by high-pressure steam or waste heat.
Absorption systems utilise the ability of liquids or salts to absorb the vapour of the
working fluid. The most common working pairs for absorption systems are:
water (working fluid) and lithium bromide (absorbent); and
ammonia (working fluid) and water (absorbent).
Figure 22 Absorption heat pump
In absorption systems, compression of the working fluid is achieved thermally in a
solution circuit which consists of an absorber, a solution pump, a generator and an
expansion valve as shown in Figure 22. Low-pressure vapour from the evaporator is
absorbed in the absorber. This process generates heat. The solution is pumped to
high pressure and then enters the generator, where the working fluid is boiled off with
an external heat supply at a high temperature. The working fluid (vapour) is
condensed in the condenser while the absorbent is returned to the absorber via the
expansion valve.
Heat is extracted from the heat source in the evaporator. Useful heat is given off at
medium temperature in the condenser and in the absorber. In the generator
high-temperature heat is supplied to run the process. A small amount of electricity
may be needed to operate the solution pump.
Heat transformers share the same main components and working principle
(absorption processes) as absorption heat pumps. Heat transformers can upgrade
waste heat virtually without an external heat source. Waste heat of a medium
temperature (i.e. between the demand level and the environmental level) is supplied to
the evaporator and generator. Useful heat of a higher temperature is given off in the
absorber. All current systems use water and lithium bromide as the working pair.
These heat transformers can achieve a delivery temperature up to 150ºC, typically
with a lift of 50ºC. COPs under these conditions range from 0.45 to 0.48.
22
2.2.3 Application of heat pumps with a condensing boiler
It should be pointed out that, in many application cases of condensing boilers, direct
condensation works well. A system of this type provides a simple and reliable
method of increasing boiler output and, at the same time, provides reasonable gas
cleaning (Fagersta Energetics, 2009a).
Figure 23
Figure 24
Flue gas condenser with a heat pump
Flue gas condenser with an absorption heat pump
If direct cooling cannot reduce the flue gas temperature sufficiently, a mechanical
heat pump is the most obvious way of providing further temperature reduction. It
enables the flue gases to be cooled to a low temperature, while at the same time
providing output heat at 80–90ºC. The system is essentially simple and uses only
conventional, tried-and-tested components. These condensing flue gas heat recovery
systems incorporating heat pumps are robust and easily-operated, as shown in Figure
23. The main drawbacks are economic: capital cost is high and the energy
consumption, normally electricity, imposes a heavy burden on the cost calculations
(Fagersta Energetics, 2009a).
23
As an alternative to the conventional electrically driven compression heat pump the
flue gas condenser can be combined with an absorption heat pump, as shown in
Figure 24. An absorption heat pump costs at least as much as a mechanical heat
pump, at any rate in terms of cost per unit of cooling power. However, it is
particularly suitable for use in process applications where there is a considerable
quantity of steam or hot water available, the heat level of which is to be reduced from
about 150ºC to 80–90ºC. The process requires about 50% more input drive energy,
in the form of heat, than the waste heat to be recovered from the flue gases. In other
words, the total heat output is about 2.5 times greater than the quantity of heat that
would be supplied from a direct-condensing cooler. Waste heat can be accepted at a
temperature of 25–30ºC and raised to 80–90ºC. If the district heating network is
capable of accepting large quantities of heat at 80–90ºC, an absorption heat pump is
an excellent solution (Fagersta Energetics, 2009a).
Figure 25 illustrates the energy saving for an example heating system firing wood
chips containing 50% water, which is the normal water content for green chips. The
flue gas temperature from the boiler is 175°C and the air to fuel ratio is 1.2. The
cooling water which can partly come from the mains as cooling water and partly from
preheating of the tap water has a temperature of 50°C. Thus, while the boiler prior
to the installation gave 10MW, it now gives 11.8MW.
Figure 25
Energy balance for the heating system with a condensing boiler
Systems involving heat pumps give considerably greater increases in efficiency than
simple flue gas coolers. This is partly due to the fact that one normally cools the
exhaust gases more and partly due to the fact that electricity is fed to the heat pump
compressor (Fagersta Energetics, 2009b).
24
Figure 26
Energy balance for the heating system with a condensing boiler and a heat
pump
Figure 26 illustrates the situation at the same plant as used above for calculation
with a simple flue gas condenser, but in this case with a heat pump installed. Thus
we have a 10MW chip fired boiler (50% moisture content), 175°C flue gas
temperature and an air to fuel ratio of 1.2. This is cooled by a heat pump with an
evaporator temperature of 25°C. 3.4MW heat is obtained from the flue gases and
1.4 MW of electric power is supplied to the heat pump compressor and also passed to
the hot water. The heat factor (COP) of the heat pump is 3.5. (Using the most
recently developed heat pumps, heat factors of 5–6 can be obtained). The electric
power consumption can be lowered further in two ways. Firstly by raising the
temperature of the evaporator and, secondly, by cooling the flue gases directly using
the cooling water (Fagersta Energetics, 2009b).
Theoretically, heat pumping can be achieved by many more thermodynamic cycles
and processes. These include Stirling and Vuilleumier cycles, single-phase cycles
(e.g. with air, CO2 or noble gases), solid-vapour sorption systems, hybrid systems
(notably combining the vapour compression and absorption cycle) and
electromagnetic and acoustic processes. Some of these are entering the market or
have reached technical maturity, and could become significant in the future.
2.3 Cost and Economical Issues
Generally, condensing boilers require specialised corrosive-resistant materials and
sophisticated controls. The addition of a condensing heat exchanger can lead to
improvement of the boiler efficiency and the conservation of fuel gas but also can
25
cause an increase of the investment cost of the equipment, which is due to the
exchanger material, valves, piping, installation, extra maintenance and resistance
increase.
Figure 27
Schematic arrangement of a heating system with a condensing heat
exchanger
A Chinese study (Che et al. 2004) analysed the feasibility of retrofitting a
conventional boiler in a heating system into a condensing boiler.
The
WNS2.8-1.0/95/70-QT boiler is a gas fired shell type boiler with output of 2.8MWth,
rated pressure of 1.0 MPa and supply water temperature of 95°C. Figure 27 shows
the schematic arrangement when the condensing heat exchanger is used to heat
domestic hot water. The excess air ratio for the boiler is 1.05.
Table 3 Cost of increased material for the condensing boiler
30
35
40
50
60
Exit flue gas temperature, °C 20
3
Heat recovered, kJ/Nm
6115
5896
5626 5291 4873 3665
2
Heating surface increase, m 144.8
108.8 95.6
83.2
57.2
36.3
Material cost (RMB)
Carbon steel
11595 8713
7655 6663 4582 2909
Stainless steel
46382 34850 30621 26653 18327 11636
PTFE
105615 79357 69727 60690 41731 26497
100
1802
19.0
140
1210
7.8
1524 626
6094 2502
13877 5699
Table 3 lists the cost of different condenser materials at various exit flue gas
temperatures. For lower exit flue gas temperatures, a larger condensing heat
exchanger is required in order to recover more heat from the flue gases.
This leads to higher costs for the condenser. On the other hand, if there is an
increase in the amount of heat recovered by the condenser, then less gas is consumed
to provide the nominal thermal output. Thus, there will be some savings in the fuel
costs, as shown in Table 4.
26
Table 4 Savings in fuel cost
Exit flue gas temperature 20
25
30
35
40
3
Heat reclaimed (kJ/Nm ) 6103 5881 5610 5274 4855
Saved gas (Nm3/h)
55.8 53.8 51.3 48.3 44.4
Saved cost (RMB/h)*
83.7 80.7 77.0 72.4 66.7
*The price of the natural gas is taken as 1.5 yuan RMB/Nm3
50
3652
33.4
50.1
60
1802
16.4
24.7
100
1210
11.0
16.6
140
609
5.6
8.3
Figure 28 presents the estimated payback period for different exit flue gas
temperatures. It can be seen that the carbon steel condenser has the shortest payback
period, and the PTFE condenser has the longest payback period, which implies that
the price of material has a significant impact on the payback period (Che et al. 2004).
Figure 28
Payback period versus different exit flue gas temperature
As the energy savings increase due to the lower exit flue gas temperature, the
payback period is greatly shortened. As shown in figure 28, when the exit flue gas
temperature is reduced to approximately 90°C, the payback period increases. This is
due to the rapid increase in the material cost. When the exit flue gas temperature
approaches the dew point of the water vapour in the flue gas, the payback period is
sharply reduced, which is due to the recovery of the latent heat in great quantities.
For the carbon steel heat exchanger, the shortest payback period is only 320 h at the
exit flue gas temperature of 55°C. For the stainless steel heat exchanger, the shortest
payback period is 850 h at the exit flue gas temperature of some 50°C. For the
PTFE heat exchanger, the shortest payback period is 1800 h (Che et al. 2004).
This estimation is based on the assumption that all three types of condensing heat
exchangers have identical lifetime. However, it is often the case that the carbon steel
27
condensers have a shorter lifetime because of their poor corrosion resistance. The
PTFE material is very corrosion resistant, but it is also very expensive. The results
show that the optimum exhaust gas temperatures for different plant lifetimes stay
unchanged while the payback periods vary very slightly (Che et al. 2004).
Table 5
Costs of the conventional combi boiler and the condensing combi boiler.
Conventional combi boiler Condensing combi boiler
Investment cost (IC), $
1150
1790
Annual fuel cost (FC), $
1043
960
Annual equivalent cost 1184
1179
(AEC), $/a
Unit fuel cost (CF), $/m3
0.398
Life time (N), year
10
Interest rate (i), %
3.85
Table 6 Benefits of the condensing gas boiler
Installed
System Type Cost of
System
Annual Energy
Energy
Consumption
Cost
(Therm Eq.)
Condensing
$304,015 197,586
Gas Boiler
Conventional
$246,450 262,670
Gas Boiler
BCHP**
$695,950 404,489
Maintenance Net Present
Cost
Cost
Net Present
Cost
Compared to
Conventional
Gas Boiler
$1,446,342 $102,947
$1,853,304
($397,026)
$1,935,248 $68,631
$2,250,329
-
($155,929) $641,701
$1,181,722
($1,068,607)
*Results are based on 20-year system life
**Building Combined Heat and Power system
Comakli (2002) employed life cycle cost analysis to evaluate the cost of a
condensing boiler over the life cycle. The initial boiler cost can be converted into a
series of equal annual costs. Thus, the annual equivalent cost (AEC) for interest rate
i and N years can be defined as (Comakli, 2002),
where IC denotes the investment cost and FC is the annual fuel cost. Table 5
compares the costs of the conventional combi boiler with the condensing combi
boiler.
The US Consortium for Energy Efficiency (CEE, 2001) retrofitted conventional gas
boilers to the condensing gas hot water boilers to obtain actual installation cost and
28
energy cost data for use in developing a screening tool (CEE, 2001). As shown in
Table 6, the total installed cost of five new condensing boilers (AERCO Brand) rated
at 2.0MMBtuh (approx. 600kW) was $304,015 .
2.4 Application of Condensing Boilers and Heat Pumps in
Heating Systems
2.4.1 Sodra Nas Vimmerby Energi AB Biomass District Heating
Plant, Sweden
This district heating plant is located close to the municipality of Vimmerby (OEPT,
2004). The plant consists of seven boilers: four oil-fired boilers, two briquette-fired
boilers and one wood chip fired boiler. The biomass-fired (wood chip) boiler was
built in 1999 and taken into operation in January 2000. The biomass-fired boiler is a
grate boiler with an output of 8MWth (Table 7), as shown in Figure 29. This boiler
is also equipped with a flue gas condenser.
In this plant, four different fuels are used. These are: gas from a sewage treatment
works, briquettes and biomass such as bark, sawdust and wood chips. The main fuel
for the biomass-fired boiler is bark and saw dust.
Fuel is delivered from sawmills
located in Vimmerby. Moisture content in fuel is 50%. The lower heating value at
50% moisture content is approximately 8000 kJ/kg.
Table 7 Specification of the biomass-fired boiler
Thermal output of the boiler (MW)
8
Efficiency according to DIN 1942. (%)
85
Thermal output of the flue gas condenser (MW)
2
Efficiency including the heat from the flue gas condenser (%) 110
Combustion equipment
Construction pressure (bar)
Grate
16
Fuel
Bark and wood chip
The fuel is delivered by lorries and dumped in a bin house. The fuel is then
transported from the delivery bin to a second bin with a scoop. From the second bin,
the fuel is transported with a scrap conveyer and finally into the furnace with a screw
conveyer. The total capacity for the two bins is 3000 m3, corresponding to four days
of operation.
The outlet temperature from the boiler is 200°C.
29
In the condenser after the boiler,
the flue gas temperature decreases to 45°C.
condenser is 2MWth, as shown in Figure 29 .
(a)
Figure 29
The heat output from the flue gas
(b)
Process diagram of the district heating plant
Approximately 27,000,000 Swedish kronor (approx. €3,200,000) were invested in
the plant in 1999. These investments included the purchase and installation of the
biomass fired boiler with all the associated gas production equipment.
By firing biomass the CO2 emissions can be reduced. Flue gas condensation
improved the plant efficiency and reduced the emission of SOx. There is an
estimated 50 tonnes/year reduction in SOx emission from plant when biomass is used
as the fuel. The plant has replaced its oil-fired boilers with biomass fired boilers
which has resulted in the replacement of approx 7000 tonnes of oil annually. This
corresponds to a decrease of 21,000 tonnes of CO2 emissions annually. The emissions
from the biomass fired boiler in 2002 are listed in Table 8.
Emission
NOx
CO
Dust
Table 8 Emissions from the plant
Emission amount in 2002
Emission factor
kg/year
mg/MJ
11 605
90
6 776
100
7 800
25
Energy produced in the plant is delivered to the district-heating grid of Vimmerby.
Approximately 90% of 1900 flats and about 40% of 1700 detached houses located in
the village of Vimmerby are connected to the district heating grid.
30
2.4.2 Kraftvarmeværk Waste Incineration Plant in Thisted Denmark
Figure 30
Process diagram of the waste incineration plant
The Kraftvarmeværk combined heating and power (CHP) plant is driven by waste
incineration to provide power and district heating for the citizens of Thisted (Climate
Solutions, 2009). A wet method of flue gas cleaning is applied in the plant by
cooling flue gases by spraying water inside scrubbers (flue gas condenser).
Meanwhile, the heat recovery takes place by condensing the water-saturated gases.
The gases are cooled by recirculated water from the district heating network, as
shown in Figure 30. The performance data of the plant is shown in Table 9.
A geothermic plant is also installed adjacent to the power plant where water with a
temperature of 45°C is pumped at the rate of 130m3/h from a bore hole 1250m deep.
The heat from the water is transferred to an absorption heat pump and an electrically
driven heat pump (Figure 30). The water then passes to an injection bore hole. The
heat pumps release an additional quantity of heat which is transferred to the district
heating system. A gas fired boiler is installed adjacent to the geothermic plant so
that the temperature in the district heating facility can be further increased. The
absorption heat pumps provide a heat production totalling 17000MWh annually
(Gotaverken Miljo AB, 1991).
31
Table 9 Performance data of the waste incineration plant
Waste incineration plant
Steam produced
tons/hr
17
Electricity output
MW
3.3
Heat output
MW
10.6
Annual production
Waste incinerated
tons/year
45000
Electricity
MWh/year
22000
Heat
MWh/year
65000
Flue gas cleaning facility
Condenser output
MW
approx. 1
3
Flue gas flow
Nm /hr
38000
Emissions
4
HCl
mg/Nm3 dg
3
HF
mg/Nm dg
0.2
3
SO2
mg/Nm dg
100
3
Cd
mg/Nm dg
0.007
3
0.3
Pb
mg/Nm dg
3
Hg
mg/Nm dg
0.01
2.4.3 The Hedenverket Waste-to-Energy Plant at Karlstad, Sweden
As shown in Figure 31, the 17MWth waste incineration boiler plant has a fabric bag
house filter with prior additive injection for acid removal. The efficiency of the filter
dictated the need for additional cleaning system and a scrubber-based system was
installed at the plant. This cleaning system is now employed to reduce emissions of
HCl, SO2, HF, NH3 and heavy metals from the plant.
Figure 31
Process diagram of the Waste-to-Energy Plant
After the bag house filter, the flue gases enter an open-type Ca(OH)2 scrubber.
32
Most acid gas components are removed from the gases in the scrubber. The second
stage of the scrubber consists of condensation tower packing with the ADIOX
material, which provides additional dioxin capture.
This condensation system recovers energy from flue gas through an absorption heat
pump system. Up to 5 MW of heating power can be recovered. The second section
also serves as a final polishing stage to meet final emission limits (Gotaverken Miljo
AB, 2004).
2.4.4 Davamyran Heat and Power Plant
The plant incinerates 175000tons/year of municipal waste and bio-fuels (20
tons/hour). In the extensive energy recovery system, the latent heat in the flue gas,
mainly in water vapour, is recovered in a condenser connected to a heat pump system.
This energy is then transferred into the district heating system of Umea city. The
Dava heating and power plant has a total heat production of 350GWh/year, of which
20% originates from the flue gas condenser.
In addition, approximately
80GWh/year of electricity is produced (Gotaverken Miljo AB, 2001), as shown in
Table 10.
Table 10 Specifications for the heat and power plant
Plant design data
Furnace type
Water-cooled grate type for waste and bio-fuel
Boiler output
55MW heat for district heating
Flue gas cleaning process Bag house filter, acid scrubber, SO2-scrubber and water
treatment
Energy output, MW
Flue gas condenser
11
Heat pumps
2×5.7
Turbine (Electricity)
15
Turbine condenser
40
The flue gas cleaning takes place in a fabric filter followed by an acid scrubber, an
SO2-scrubber, and a gas condenser. Water is also recovered from the gas. Thus,
the cleaning process is self-sufficient with regard to water. The typical emissions
from this plant are shown in Table 11.
33
Figure 32
Table 11
Pollutant
Dust
HCl
HF
SOx
NH3
Cd+Tl
Hg
Dioxin
Process diagram of the combined heat and power plant
Emissions from the combined heat and power plant
Emission limits (24-hour average)
units
5
mg/Nm3
5
mg/Nm3
1
mg/Nm3
25
mg/Nm3
5
mg/Nm3
0.05
mg/Nm3
0.03
mg/Nm3
0.1
ng/Nm3
2.4.5 The Vestforbranding Waste to Energy Plant in Copenhagen,
Denmark
The plant is the largest waste-to-energy plant in Denmark. It produces 140GWh of
electricity and 400GWh of district heating every year. The flue gas condenser and
integrated absorption heat pumps were installed in February 2006, as shown in Figure
33 (Gotaverken Miljo AB, 2007a).
The incineration line was operated using conventional wet scrubbing technology
including an HCl and SO2 scrubber. The plant is being expanded to allow a
maximum of energy to be recovered from flue gases through the installation of a
condensing scrubber (Figure 34) and absorption heat pumps (Figure 35).
Flue gases are cooled by a circulating cooling water system, which allows a
substantial amount of energy to be recovered (nominal output 13MWth, maximum
17MWth, Table 12). The temperature of the heat recovered from the flue gases is
lower than the district heating return temperature. Low value energy is raised to
high value energy by two steam-driven heat pumps in series which increase the
district heating temperature from 60°C to 80°C.
34
Figure 33
Process diagram of the waste to energy plant
Table 12
Waste throughput
Thermal capacity
Flue gas flow
Max extended energy recovery
Figure 34
Performance data of the plant
26 ton/h
74MWth
150000Nm3/h (w.g.)
17MW
Condensing scrubber in the plant
35
Figure 35
Heat pumps in the plant
2.4.6 Sonderborg Waste to Energy Plant, Denmark
Sonderborg waste-to-energy plant realises a large potential to recover energy using
flue gas condensation. Conventional wet scrubbing technology with an HCl and a
SO2 dioxin scrubber is used to clean the flue gases. The plant is now upgraded with
a condensing scrubber and condensate treatment, as shown in Figure 36.
Figure 36
Process diagram of the waste to energy plant
Flue gases are cooled by a circulating cooling water system (indirect district heating
water) which allows a substantial amount of energy to be recovered (nominal output
4.5MWth). The condensate water produced is fed back upstream to the flue gas
cleaning scrubbers. In normal operation, this water will replace all the fresh water
used in the gas treatment (Gotaverken Miljo AB, 2007b).
36
3. Case Study: Condensing Boiler Design for a Biomass
Heating Plant
In general the decision (based on economic feasibility) to integrate a condensing
boiler into a heat system is mainly case-dependent.
In this case study, a series of calculations were carried out using an existing large
scale biomass heating plant. The aim was to determine the thermal design of the
condensing boiler. Further work was also carried out in order to investigate various
technical and economic issues in relation to the condensing boiler application.
3.1 Plant Description
Oriketo heating station is the largest biofuel-fired heating station in Finland (tekes,
2008), as shown in Figure 37. Located in the industrial area of Oriketo, this station
was commissioned in November 2001. The heat generated replaces district heat
energy generated from fossil fuels. The main fuel is logging residue delivered
mainly from final felling of spruce-dominant forests, plus other forestry residues and
by-products from sawmills, such as sawdust, bark, wood chips and cutter chips.
Figure 37
The Oriketo heating station
Wood is burned in a fluidised-bed boiler. The output of the boiler is 40MW at full
fuel feed, and the temperature of flue gases after the boiler is about 150oC. The flue
gases are then led to an electric precipitator for removing particles from the flue gases
37
with efficiency > 99%. The ash separated is comprised of clean wood ash and can
be used as fertilizer in the forests. The amount of ash is about 800 t/a. After the
electric precipitator the flue gases are channelled into a flue gas scrubber and
condensing plant.
The average moisture content of the wood fuels is 50%. Therefore, the flue gases
contain an abundance of water as steam. In the flue gas condensing plant, the
temperature of flue gases is decreased to 35oC and the most of water vapour is
condensed to water. About 12MW of district heat capacity is produced at the
condensing plant. This increases the derived efficiency to as high as 118%, when
calculated from the effective heat value of the fuel prior to combustion.
30
Figure 38
Flow chart of the Oriketo heating station
The condensed water is used for heating buildings and yard prior to leading the
water into the sewage. Finally, the flue gases are led through a 60m high stack to the
open air. As the fuels do not contain sulphur, no sulphur oxides are formed in
combustion. The emissions of nitrogen oxides (NOx) are about 140t/a, and particles
emissions are 5t/a, as shown in Figure 38.
The total heat output of the plant is 52MWth, and the yield of energy is about 300
GWh/a for an annual operating time of 7000 hours. The cost of construction
amounted to €14.3 million.
3.2 Conditions of the Heating Plant
3.2.1 Fuel input
38
In the heating plant, the main fuel is logging residue delivered mainly from final
felling of spruce-dominant forests. Table 13 presents the ultimate analyses and
calorific values of Spruce wood obtained from literature (Demirbas, 2009). These
values were used in our case study calculations.
As the moisture content in the wood chips is as high as 50%, the low heating value
of the fuel reduces to 8.16MJ/kg. Hence the fuel input for this plant is
approximately 19.4tonnes/hr.
Table 13
C
H
N
S
O
Ash
Moisture
GCV, MJ/kg
NCV, MJ/kg
Properties of the fuel input for the heating plant
Dry basis As received
51.2
25.6
6.1
3.05
0.3
0.15
40.9
20.45
1.5
0.75
50.0
20.1
10.05
8.16
3.2.2 Process Parameters
Table 14 presents a summary of some of the process parameters for the heating
plant. Using the data presented in Tables 13 and 14, the main properties of the
relevant streams in the process (as shown in Figure 39) were evaluated based on mass
and energy balances. In this simplified calculation, the heat losses due to heat
dissipation and incomplete combustion of fuel are not considered. The heat loss of
the flue gas in the ESP is also neglected.
Table 15 lists the conditions for each stream in the system. Consequently, the
conditions of the inlet and outlet of the condensing boiler are obtained.
Table 14 Known parameters for the heating plant
Operating pressure, bar
1
150
Temperature of the flue gas exit the fluidised bed boiler, °C
35
Temperature of the flue gas to the stack, °C
Hot water pressure, bar
16
140
Hot water temperature, °C
Heat capacity of the fluidised bed boiler, MW
40
55
Temperature of the feed water to the fluidised bed boiler, °C
Reference states (Pref, Tref)
1 atm, 25°C
39
Figure 39
Table 15
Process diagram for the heating plant
Calculated process parameters for the heating plant
Stream No. Pressure, bar Temperature, °C Mass flow rate, kg/s Enthalpy, kJ/kg
1
1
20
5.4
-17.2
2
1
20
20.9
-5.15
3
1
150
26.3
146.7 (+388.4)*
4
1
150
26.3
146.7 (+388.4)
5
1
35
23.0
10.6 (+92.1)
6
16
140
111.6
590.0
7
16
30
111.6
128.0
8
16
55
111.6
231.7
9
1
35
3.32
42.4
* Data in the parentheses are the potential latent heat of the moisture in the flue gas
3.2.3 Flue Gas Composition
Table 16 Composition of the flue gas before entering the condensing boiler
mol/s
kg/s
mol fraction mass fraction
CO2
115.1
5.1
12.1
19.3
H2O
232.4
4.2
24.4
15.9
O2
30.6
1.0
3.2
3.7
N2
573.3
16.1
60.3
61.1
In the condensing boiler, the amount of heat recovered from the flue gas largely
depends on the gas composition. The important parameter is the water vapour
40
content. It determines the dew point of the flue gas and thus the potential amount of
latent heat that could be recovered. Table 16 shows the composition of the flue gas
prior to the condensing boiler. Note that the vapour pressure in the flue gas is about
0.244bar. The dew point can thus be calculated to be 64.3°C.
3.3 Condensing Boiler Design
3.3.1 Heat Exchanger Selection
A condensing boiler mainly consists of heat exchangers containing heat transfer
elements and fluid distribution elements. Based on the construction features, heat
exchangers can be categorised into four major types: tubular, plate-type, extended
surface, and regenerative exchangers (Shah and Sekulic, 2003).
Tubular heat exchangers are generally built of circular tubes, although elliptical,
rectangular, or round/flat twisted tubes have also been used in some applications.
There is considerable flexibility in the design because the core geometry can be varied
easily by changing the tube diameter, length, and arrangement. Tubular exchangers
can be designed for high pressures relative to the environment and high-pressure
differences between the fluids.
Tubular exchangers are used primarily for
liquid-to-liquid and liquid-to-phase change (condensing or evaporating) heat transfer
applications.
They are used for gas-to-liquid and gas-to-gas heat transfer
applications primarily when the operating temperature and/or pressure is very high or
fouling is a severe problem on at least one fluid side and no other types of exchangers
would work. These exchangers may be classified as shell-and-tube, double-pipe,
and spiral tube exchangers (Shah and Sekulic, 2003).
Shell-and-tube exchanger is generally built of a bundle of round tubes mounted in a
cylindrical shell with the tube axis parallel to that of the shell. One fluid flows
inside the tubes, the other flows across and along the tubes. They are the most
versatile exchangers, made from a variety of metal and nonmetal materials (such as
graphite, glass, and Teflon) and range in size from small (0.1m2) to supergiant (over
105m2) surface area. The major components of this exchanger consist of tubes (or
tube bundle), shell, frontend head, rear-end head, baffles, and tubesheets.
A variety of different internal constructions are used in shell-and-tube exchangers,
depending on the desired heat transfer and pressure drop performance and the
methods employed to reduce thermal stresses, to prevent leakages, to provide for ease
of cleaning, to contain operating pressures and temperatures, to control corrosion, to
accommodate highly asymmetric flows, and so on. Shell-and-tube exchangers are
classified and constructed in accordance with the widely used TEMA (Tubular
Exchanger Manufacturers Association) standards, DIN and other standards in Europe
and elsewhere, and ASME (American Society of Mechanical Engineers) boiler and
41
pressure vessel codes. TEMA has developed a notation system to designate major
types of shell-and-tube exchangers. In this system, each exchanger is designated by
a three-letter combination, the first letter indicating the front-end head type, the
second the shell type, and the third the rear-end head type. These are identified in
Figure 40 (TEMA, 2003). Some common shell-and-tube exchangers are AES, BEM,
AEP, CFU, AKT, and AJW.
Figure 40
TEMA classification of heat exchangers
Depending on the application, a specific combination of geometrical variables or
types associated with each component can be selected. Since the desired heat
42
transfer in the exchanger takes place across the tube surface, the selection of tube
geometrical variables is important from a performance point of view. In most
applications, plain tubes are used. However, when additional surface area is required
to compensate for low heat transfer coefficients on the shell side, low finned tubing
with 250 to 1200fins/m and a fin height of up to 6.35 mm is used. While
maintaining reasonably high fin efficiency, low-height fins increase surface area by
two to three times over plain tubes and decrease fouling on the fin side based on the
data reported (Shah and Sekulic, 2003).
The most common plain tube sizes have 15.88, 19.05, and 25.40 mm (5/8, 3/4, and
1in.) tube outside diameters (do). From the heat transfer viewpoint, smaller-diameter
tubes yield higher heat transfer coefficients and result in a more compact exchanger.
However, larger-diameter tubes are easier to clean and more rugged. The foregoing
common sizes represent a compromise. For mechanical cleaning, the smallest
practical size is 19.05 mm (3/4 in.). For chemical cleaning, smaller sizes can be
used provided that the tubes never plug completely.
The selection of tube pitch (pt) is a compromise between a close pitch (small values
of pt/do) for increased shell-side heat transfer and surface compactness, and an open
pitch (large values of pt/do) for decreased shell-side plugging and ease in shell-side
cleaning. In most shell-and-tube exchangers, the ratio of the tube pitch to tube
outside diameter varies from 1.25 to 2.00. The minimum value is restricted to 1.25
because the tubesheet ligaments may become too weak for proper rolling of the tubes
and cause leaky joints.
Figure 41
Layouts of the tubes
Two standard types of tube layouts are the square and the equilateral triangle,
shown in Figure 41. The equilateral pitch can be oriented at 30° or 60° angle to the
flow direction, and the square pitch at 45° and 90°. Note that the 30°, 45° and 60°
arrangements are staggered, and 90° is inline. For the identical tube pitch and flow
rates, the tube layouts in decreasing order of shell-side heat transfer coefficient and
pressure drop are: 30°, 45°, 60°, and 90°. Thus, the 90° layout will have the lowest
heat transfer coefficient and the lowest pressure drop.
43
The E shell (as shown in Figure 40), the most common due to its low cost and
relative simplicity, is used for single-phase shell fluid applications and for small
condensers with low vapour volumes. Multiple passes on the tube side increase the
heat transfer coefficient. However, a multipass tube arrangement can reduce the
exchanger effectiveness compared to that for a single-pass arrangement (due to some
tube passes being in parallel flow) if the increased heat transfer coefficient does not
compensate for the parallel flow effect. Two E shells in series (in overall
counter-flow configuration) may be used to increase the exchanger effectiveness
(Shah and Sekulic, 2003).
The function of the cross baffle is to direct the flow across the tube field as well as
to mechanically support the tubes against sagging and possible vibration (Taborek,
1983). The most common type is the segmental baffle, with a baffle cut resulting in
a baffle window. Baffle spacing is subject to minimum and maximum limitations for
good thermo-hydraulic performance and tube support. The practical range of single
segmental baffle spacing is 1/5 to 1 shell diameter, although the optimum could be 2/5
to 1/2. The ratio of baffle spacing to baffle cut is a crucial design parameter for
efficient conversion of pressure drop to heat transfer. If very low pressure drops
have to be accommodated, so-called double-segmental or disk-and-doughnut baffles
will reduce the pressure drop by about 60%. Other types include triple-segmental
and no-tubes-in-window, for particularly low pressure drops and prevention of tube
vibration.
In this case study, the condensing boiler is assumed to consist of a single-pass
shell-and-tube heat exchanger. Some of the assumptions about the tube dimensions
and pattern used in the calculation are listed in Table 17.
Table 17 Dimensions of the condensing boiler
Heat exchanger type
Single tube pass, counter-current shell-and-tube
exchanger (E type shell)
Tube outside diameter, do (mm)
25.4
Tube inner diameter, di (mm)
22.9
Tube thickness, δt (mm)
1.25
Pitch, pt/do
1.75
Total tube number, N
1024 (32×32)
Tube layout
Rotated square as shown in Figure 41
Shell inner diameter, Do (mm)
2090
Shell thickness, δs (mm)
14
Baffle type
Single-segmental
Baffle spacing, B (mm)
1776
Baffle cut
25%
44
3.3.2 Condensation Curve
In the counter-current condensing boiler, the shell-side stream (the flue gas) enters
with specific enthalpy hs,in and leaves with specific enthalpy hs,out. The tube side
(water) specific enthalpy changes from ht,in to ht,out. If the shell-side and tube-side
mass flows are Gs and Gt, respectively, then, the heat balance over the condensing
boiler gives (Butterworth, 1991),
ht = ht ,in +
Gs
(hs − hs,out )
Gt
(1)
where ht and hs are the tube-side and shell-side specific enthalpy in the condensing
boiler, respectively. In the shell-side, when condensation does not occur, the specific
enthalpy of the flue gas can be calculated as,
[
]
hs = c p , g (Ts − Tref ) + x0 × c p.w (Ts − Tref ) + i w (Ts )
(2)
where cp,g is the specific heat capacity of non-condensable gases, cp,w the specific heat
capacity of the water vapour, iw, the latent heat of water, and x0, the initial molar
fraction of water vapour in the flue gases. When condensation happens, the specific
enthalpy in the shell-side can be obtained from,
[
]
hs = c p , g (Ts − Tref ) + x s × c p , w (Ts − Tref ) + i w (Ts ) + ( x0 − x s )c w (Ts − 273.15)
(3)
where xs is the molar fraction of water vapour in the saturated flue gases, cw, the heat
capacity of liquid water. In the tube-side, the specific enthalpy of the coolant (water)
is expressed as,
ht = c w (Tt − 273.15)
(4)
Based on Eqs. (1) – (4), the corresponding temperatures can be plotted as shown in
Figure 42. This figure shows the equilibrium condensation curve for the flue gases,
where the equilibrium vapour temperature is plotted versus the difference of the
specific enthalpy of the mixture from the outlet, assuming a constant pressure
throughout. The curve clearly indicates that along the path of condensation, as the
water vapour condenses out, the equilibrium condensing temperature drops. As a
result, the temperature difference between the gas mixture and the coolant is reduced,
leading to a lower heat transfer rate. The real condensing curve may not follow this
equilibrium curve closely since condensation is a non-equilibrium process.
Nevertheless, this curve shows the correct trend and the implications for design
(Marto, 1991).
As shown in Figure 42, the tube-side temperature changes linearly whereas the
temperature variations in the shell-side show a de-superheating zone together with
condensation occurring in the presence of non-condensable gases. According to the
shell-side temperature variations, the diagram can be divided into zones where the
45
temperature curves on both sides are almost linear. In Figure 42, the condensation
curve is divided into four zones, as shown by the vertical dashed lines as zone
boundaries. Zone I represents the de-superheating of the flue gases in which the flue
gas temperature decreases linearly. At the boundary B, the water vapour in the flue
gas begins to condense. Over each zone, the temperature difference (Ts-Tt) varies
linearly with hs, i.e., with the amount of heat transferred from the shell-side to the
tube-side. Table 18 presents the temperatures in the shell-side and the tube-side for
each zone.
160
A
Flue gas (shell side)
Return water (tube side)
140
o
Temperature, C
120
100
80
B
60
C
E
40
D
IV
20
Zone III
Zone II
Zone I
0
0
50
100
150
200
250
300
350
400
450
Specific enthalpy difference (h -h outlet) on the shell side, kJ/kg
Figure 42
Table 18
Zone Boundaries
A
B
C
D
E
Condensation curve for the flue gas
Boundaries of the four zones in Figure 42
Shell-side
Tube-side
150.0
55.0
64.3
49.3
50.0
37.0
40.0
32.0
35.0
30.0
At the boundaries of each zone, the overall heat transfer coefficients (Ua and Ub)
can be calculated. A mean overall heat transfer coefficient (Um) for each zone can
thus be obtained through the following three equations, whichever is most appropriate
(Butterworth, 1991).
i) if the heat transfer coefficient U varies linearly with A, then,
46
(5)
ii) if both U and the temperature difference (θ=Ts-Tt) vary linearly with the amount
of heat transferred,
(6)
iii) if both 1/U and θ vary linearly with the amount of heat transferred,
(7)
These equations will not usually be valid over the whole of the condenser but may
apply to small portions of it. If Ua and Ub vary only by a small amount, Eq. (5) is
preferred because of its simplicity. There is a long tradition in the use of Eq. (6) but
with little justification. Eq. (7) seems more in line with the variations observed in
condensers and is hence recommended in those situations when Eq. (5) cannot be
used due to the large difference between Ua and Ub. Of course, any question about
which equation is more accurate can always be avoided by dividing the exchanger
into a large number of sections (Butterworth, 1991).
3.3.3 Thermal Design Methodology
In general there are four methods for heat exchanger design: ε-NTU, LMTD,
P-NTU, and ψ-P methods, among which ε-NTU and LMTD methods are most
commonly used (Shah and Sekulic, 2003).
ε-NTU Method
In the ε-NTU method, the total heat transfer rate from the hot fluid to the cold fluid
in the heat exchanger is expressed as,
(8)
where Cmin is the minimum of the heat capacity rate of the hot and cold fluids
(Ch= m& ch and Cc= m& cc), ∆Tmax= (Th,i – Tc,i), the fluid inlet temperature difference
(ITD). ε is the heat exchanger effectiveness, a measure of thermal performance of a
heat exchanger. It is defined as the ratio of the actual heat-transfer rate, q, to the
thermodynamically possible maximum heat-transfer rate (qmax) by the second law of
thermodynamics,
47
(9)
It is non-dimensional and dependent on NTU (number of heat transfer units), C* (heat
capacity rate ratio), and the flow arrangement, as follows,
(10)
The heat capacity rate ratio, C*, is simply the ratio of the smaller to larger heat
capacity rate for the two fluid streams so that C* <1,
(11)
NTU designates the non-dimensional “heat-transfer size” or “thermal size” of the
exchanger. It is defined as a ratio of the overall conductance to the smaller heat
capacity rate.
(12)
LMTD Method
In a heat exchanger, the maximum driving force for heat transfer is generally the log
mean temperature difference (LMTD) when two fluid streams are in countercurrent
flow (Kuppan, 2000). The log-mean temperature difference (LMTD or ∆Tlm) is
defined as,
(13)
where ∆TI and ∆TII are temperature differences between two fluids at each end of a
counter-flow or parallel-flow exchanger. However, the overriding importance of
other design factors causes most heat exchangers to be designed in flow patterns
different from true counter-current flow. The true mean temperature difference of
such flow arrangements will differ from the logarithmic mean temperature difference
by a certain factor dependent on the flow pattern and the terminal temperatures.
This factor is usually designated as the log mean temperature difference correction
factor, F. The factor F may be defined as the ratio of the true mean temperature
difference (MTD) to the logarithmic mean temperature difference and the heat
transfer rate equation incorporating F is given by,
(14)
48
Generally, F is dependent upon the thermal effectiveness P, the heat capacity rate
ratio R, and the flow arrangement. The thermal effectiveness P is the ratio of the
heat actually transferred, to the heat which would be transferred, if the same
cold-fluid temperature was raised to the hot-fluid inlet temperature, i.e.
(15)
The heat capacity rate ratio, R, is defined as the ratio of the capacity rate ( m& cp) of
the cold fluid to that of the hot fluid, as follows,
(16)
The value of R ranges from zero to infinity, zero being for pure vapor condensation
and infinity being for pure liquid evaporation.
For a single-pass cross-flow shell-and-tube heat exchanger, the dependent function
for F is as follows,
(17)
Generally, the ε-NTU method is used for the design of compact heat exchangers.
The LMTD method is used for the design of shell-and-tube heat exchangers. It
should be emphasized that either method will yield identical results within the
convergence tolerances specified (Kuppan, 2000).
In this case study, the LMTD method is employed for the thermal design of the
condensing boiler. Based on the temperatures listed in Table 18, the log-mean
temperature differences and F values were calculated, as presented in Table 19.
Table 19
LMTD
Zone I
Zone II
Zone III
Zone IV
43.3
14.0
10.3
6.2
LMTD for each zone
F
Heat transfer rate, MW
0.971
0.865
0.935
0.966
3.3.4 Heat Transfer Coefficients
49
2.652
5.757
2.296
0.865
Mechanisms of Condensation
A clear understanding of the underlying heat and mass transfer processes is
necessary for reliable selection and design of condensing boilers. However, these
processes are very complex (Chisholm, 1983). At the present state of development,
the film model has been developed for a somewhat crude approximation to the
physical reality. The transfer of heat and mass is considered to be impeded by a
number of resistances, localised in a series of real or hypothetical layers or films, as
shown in Figure 43. There are as many as three resistances to heat and mass transfer
which may be considered to arise on the gaseous side of the wall: the condensate
resistance, the interfacial resistance and the resistance of the vapour film.
Figure 43
Resistances of the heat and mass transfer during the condensation
A condensate resistance is always present in condensation, though it may arise from
the presence of a continuous film, droplets or a combination of these. The heat flux
in the condensate film (which may be real or hypothetical) will vary and the
temperature profile will be non-linear due to sub-cooling of the condensate.
However, the sensible heat of sub-cooling is always small compared to the latent heat
and may be neglected or taken into account by assuming it to be constant across the
liquid film when corrected for sub-cooling.
With the presence of a non-condensable gas, the heat and mass transfer in the
condensation process becomes far more complex than for a pure vapour condensation.
The process involves mass transfer effects that create additional thermal resistances,
thus lowering the overall heat transfer coefficient. As shown in Figure 44, toward
the interface between the condensate and the gas, the local temperatures and pressures
vary from the bulk conditions. The presence of the gas decreases the resulting local
heat transfer rate in two ways. First, in the presence of a non-condensable gas, the
50
water vapour exists at a partial pressure Pgb causing the bulk vapour temperature Tg to
be less than the saturation temperature. In addition, as the vapour molecules migrate
toward the cold wall, they sweep non-condensable gas molecules with them. Since
the non-condensable gas does not condense at the prevailing operating conditions in
the condenser, these gas molecules accumulate near the liquid-vapour interface. The
concentration profile of these gas molecules reaches an equilibrium condition due to a
local balance of vapour momentum effects in one direction and back-diffusion effects
in the other. As a result, the local partial pressure of the non-condensable gas
increases to a maximum at the interface. The vapour molecules must travel through
this gas-rich layer and, since the total pressure of the mixture is constant, the vapour
partial pressure decreases from Pgb to Pgi. This lower vapour pressure at the
interface corresponds to a lower vapour temperature Tl, which creates a reduced
effective temperature difference across the condensate film (Butterworth, 1991).
Figure 44
Temperature and pressure profiles around the condensate film
Due to the complexity and the important role of mass diffusion during condensation
of flue gases, two kinds of analytical methods have been developed for analysing the
heat transfer process, namely “equilibrium methods” and “non-equilibrium (or
differential) methods” (Marto, 1991).
The equilibrium methods assume that there is local equilibrium between the
gaseous phase and the condensate throughout the condenser (Marto, 1991). Thus the
gas temperature follows the equilibrium condensation curve (Figure 42). These
methods are particularly well suited to the situation where vapour and condensate do
not become separated, since in this case the overall local composition is the same as
the vapour feed composition (Chisholm, 1983).
The advanced non-equilibrium/differential methods include film, penetration, and
51
boundary layer models. These models provide physically realistic formulations of
the problem, yielding more accurate local coefficients at the expense of considerable
complexity. In these methods, the calculation of local heat and mass transfer rates
are combined with differential mass and energy balances, which describe the
downstream development of the independent vapour and coolant temperatures and
vapour composition though the condenser.
The equilibrium methods have the advantages of simplicity and speed. As the
Silver equilibrium method is very widely applied in engineering design practice, this
case study employs this method for the condenser design.
The local overall heat transfer coefficient (U) from the bulk vapour mixture to the
coolant is written as
(18)
where hc is the heat transfer coefficient on the tube side (the coolant), R is the thermal
resistance due to the tube wall (and any fouling), and hef is an effective
condensing-side heat transfer coefficient, which includes the thermal resistance across
the condensate film, as well as the sensible cooling of the gas. This effective
coefficient is obtained by writing the overall temperature difference from the bulk gas
to the wall as,
(19)
Since each temperature difference may be written in terms of a heat flux divided by a
heat transfer coefficient, this equation can be expressed as,
(20)
Therefore,
(21)
Tube-side Heat Transfer Coefficient
On the tube-side, the flow rate (Gt) of water is approximately 111.6kg/s and the total
cross-sectional area (St) of the tube bundles is 0.42m2. Thus the mean velocity (vt)
of the water in a tube is about 0.27m/s and the Reynolds number (Re) ranges from
7×103 to 1.2×104. Consequently, the correlation (Nusselt number) obtained under
fully developed turbulent flow in smooth tubes can be used to calculate the tube-side
heat transfer coefficient (Kakac and Liu, 2002),
52
(22)
where Pr is the Prandtl number and f can be expressed as,
(23)
The tube-side heat transfer coefficient (hc) can thus be obtained through
hc =
Nu b λc
di
(24)
where λc is the thermal conductivity of the water and di is the inner diameter of the
tubes. Table 20 presents the tube-side heat transfer coefficients at the boundaries of
the four zones.
Table 20 Tube-side heat transfer coefficients
A
B
C
55
49.3
37.0
Temperature, °C
Velocity, m/s
0.268
0.267
0.267
3
Density, kg/m
986.3
989.0
991.8
Heat capacity, kJ/kgK
4.18
4.18
4.18
-4
-4
Viscosity, PaS
5.04×10
5.54×10
6.09×10-4
Thermal conductivity, W/mK
0.650
0.643
0.627
Pr
3.24
3.59
4.06
4
4
Re
1.2×10
1.1×10
9.9×103
f
0.0075
0.0077
0.0079
Nu
69.4
66.6
64.1
2
Tube-side coefficient, W/m K 1968.4
1870.5
1754.2
53
D
32.0
0.266
994.9
4.18
7.26×10-4
0.623
4.86
8.3×103
0.0083
58.5
1591.6
E
30.2
0.265
996.0
4.18
7.79×10-4
0.618
5.26
7.8×103
0.0085
56.4
1522.0
Table 21
Fouling resistances for different fluids
It is certain that fouling may occur inside and/or outside the tubes in the condensing
boiler. Although fouling is time dependent, only a fixed value can be prescribed
during the design stage. Inside the tubes, the feed water to the boiler should be
chemically treated. However, outside the tubes, the flue gas contains ultrafine
particles and trace acid gases. Thus the condensate on the shell-side may contain
some amounts of solid and liquid contaminants. Consequently, the fouling
resistances inside and outside the tubes were chosen from the TEMA tables as shown
in Table 21: Rf,i=0.000176 m2K/W, Rf,o=0.00176 m2K/W (coal flue gas) (Kakac and
Liu, 2002).
Shell-side Heat Transfer Coefficient
On the shell-side, the volumetric flow rate of the flue gas decreases from 21.3
Nm3/s at the inlet to 17.2Nm3/s at the outlet. Part of the water vapour condenses to
liquid water. As stated previously, the heat resistances in the shell-side consist of
those of the condensate film and the cooling of the sensible heat of the flue gases.
For the heat transfer coefficient of the gas stream (hg) in the shell side, the following
correlation can be used (Chisholm, 1983),
Nu =
h g De
λg
= 0.27 Re 0De.63 Prg0.34
(25)
where λg is the thermal conductivity of the gas mixture and De is the equivalent
diameter calculated along (instead of across) the long axes of the shell.
54
Figure 45
Equivalent diameter and the single-segmental baffles
The equivalent diameter of the shell is taken as four times the net flow area as
layout on the tube sheet divided by the wetted perimeter. For the square-pitch layout
shown in Figure 45, the equivalent diameter can be calculated by (Kakac and Liu,
2002),
(26)
In the baffled heat exchanger, the variables that affect the gas velocity are shell
diameter (Ds), the clearance (C) between adjacent tubes, the pitch size (PT), and the
baffle spacing (B). The width of the flow area at the tubes located at the centre of
the shell is (Ds/PT)×C and the length of the flow area is taken as the baffle spacing (B).
Thus the bundle cross-flow area (As) is,
(27)
From this equation, the gas velocity on the shell-side can be calculated. The heat
transfer coefficients of the flue gases are calculated and presented in Table 22.
55
Table 22
Shell-side heat transfer coefficients for the gas stream
A
B
C
D
150.0
64.3
50.0
40.0
Temperature, °C
Flow rate, Nm3/s
21.3
21.3
18.5
17.5
3
Density, kg/m
0.796
0.998
1.097
1.157
Heat capacity, kJ/kgK
1.176
1.141
1.069
1.039
-5
-5
-5
Viscosity, PaS
2.08×10
1.73×10
1.76×10
1.75×10-5
Thermal conductivity, W/mK
0.032
0.026
0.025
0.025
0.074
0.074
0.074
0.074
Equivalent diameter (De), m
2
Crossflow area (As), m
1.59
1.59
1.59
1.59
Gas velocity, m/s
13.4
13.4
11.7
11.0
Pr
0.77
0.77
0.74
0.73
4
4
4
Re
3.8×10
5.7×10
5.4×10
5.4×104
Nu
188.1
244.0
231.5
230.2
2
Coefficient (hg) , W/m K
81.1
84.7
79.4
77.7
E
35.0
17.2
1.185
1.028
1.74×10-5
0.025
0.074
1.59
10.8
0.73
5.4×104
230.9
77.2
In addition to the thermal resistance of the gas stream (1/hg), there exist the
resistance of the condensate film in Zones II – IV. Nusselt treated the case of
laminar film condensation of a quiescent vapour on an isothermal horizontal tube.
The analysis yields the average heat transfer coefficient (hm) outside the top tube
upstream as follows (Kakac and Liu, 2002),
(28)
where ρl, kl, and µl are the density, thermal conductivity and viscosity of the
condensate, respectively, ilg is the latent heat, Tsat and Tw are the saturation
temperature and the tube wall temperature, respectively. Including the inundation
effect in the tube bundles, the average coefficient for a vertical column of N tubes
(hm,N) compared to the coefficient for the first tube (i.e., the top tube in the row) is
(29)
The effective heat transfer coefficients for the condensate film can then be
calculated according to Eq. (21), as listed in Table 23.
Table 23
ilg, MJ/kg
Overall shell-side heat transfer coefficients
B
C
D
2.35
2.38
2.41
56
E
2.42
Density, kg/m3
Viscosity, PaS
Thermal conductivity, W/mK
Tsat-Tw
Nu1
h1, W/m2K
hN, W/m2K
Z
Shell-side
heat
transfer
2
coefficient, heff, W/m K
981.1
4.40×10-4
0.658
15.01
391.5
10143
4264.6
0.083
821.0
988.0
5.47×10-4
0.644
13.03
388.9
9855
4143.4
0.106
636.7
992.2
6.53×10-4
0.631
7.95
425.0
10553
4437.0
0.133
516.0
994.0
7.20×10-4
0.623
4.81
472.9
11604
4879.0
0.150
466.0
Thermal Resistance of the Tube Wall
In general, heat exchangers can be made from a variety of metals (aluminium,
copper, steel alloys, etc.) and non-metal materials (such as graphite, glass, and Teflon,
etc). Different materials lead to different thermal resistance in condensing boilers
and eventually different cost of the equipment.
Generally, carbon steel and stainless steel are two of the most common materials for
industrial heat exchangers. In this case study, two materials (Stainless steel 316 and
carbon steel) were chosen for calculation purposes. The aim was to investigate their
impacts on the thermal design of a condensing boiler and their associated cost
implications. It should be noted that the flue gas from wood fuel combustion
contains nitric oxides, chloride, and sulphate/sulphite. So the material used to make
the shell and tubes must be corrosion resistant. Stainless steel is a good
corrosion-resistant material. It differs from carbon steel by the amount of chromium
present in it. By contrast, carbon steel rusts quickly when exposed to the air and
moisture. Thus in order to use carbon steel in the condensing boiler, the outer
surface of the tubes and the inner surface of the shell must be coated or lined with a
corrosion resistant material for protection purposes.
Polypropylene (PP) is produced by the polymerization of propylene, a relatively
inexpensive olefin derived from petroleum. The use of polypropylene has expanded
through the years due to its high strength to weight ratio, excellent resistance to
corrosion, ease of fabrication, and low cost. Polypropylene's main characteristics
include its resistance to strong acids, even at elevated temperatures. The melting of
polypropylene occurs over a specific range. Isotactic PP has a melting point of
171°C. Commercial isotactic PP has a melting point that ranges from 160 to 166°C,
depending on atactic material and crystallinity (Maier and Calafut, 1998).
Table 24 compares the thermal resistance of the stainless steel tubes with the carbon
steel tubes coated with PP.
57
Table 24
Thermal resistances of the tube wall
Stainless Steel
Carbon steel + PP
Metal thickness, mm
1.245
1.245
Coating thickness, mm
0.125
Thermal conductivity of metal*, W/mK 19
43
Thermal conductivity of PP, W/mK
0.12
2
-5
Thermal resistance, m K/W
6.55×10
9.23×10-4
*(Green and Perry, 2008)
Overall Heat Transfer Coefficient
The values for thermal resistances in condensing boilers made from stainless steel
and carbon steel are shown in Tables 25 and 26, respectively. Tables 27 and 28
compare the results for each type of the resistances.
Table 25 Thermal resistances in the stainless steel condensing boiler (unit: m2K/W)
A
B
C
D
E
-4
-4
-4
-4
Tube-side fluid
5.08×10
5.35×10
5.70×10
6.28×10
6.57×10-4
Tube-side fouling
1.76×10-4 1.76×10-4 1.76×10-4 1.76×10-4 1.76×10-4
Tube wall
6.55×10-5 6.55×10-5 6.55×10-5 6.55×10-5 6.55×10-5
Shell-side fouling
1.76×10-3 1.76×10-3 1.76×10-3 1.76×10-3 1.76×10-3
Shell-side fluid
1.23×10-2 1.22×10-3 1.57×10-3 1.94×10-3 2.15×10-3
In both condensing boilers, the thermal resistances in the tube-side (including the
fluid and fouling) are relatively small, whereas the shell-side resistances contribute
approximately 60 – 90% of the total resistance. As can be seen, in Zone I where no
condensation occurs, the thermal resistance of the shell-side flue gas flow is the
dominant factor. When the condensate film forms around the tubes, the thermal
resistance of the shell-side fluid decreases and the shell-side fouling becomes a major
contributor to the total thermal resistance. Due to the low thermal conductivity of PP,
the thermal resistance of the tube wall in the carbon steel condenser is much higher
than for the stainless steel.
Using Eqs. (5) and (18), the mean overall heat transfer coefficients for each zone
are calculated and listed in Table 29.
Thermal resistances in the carbon steel condensing boiler (unit: m2K/W)
A
B
C
D
E
-4
-4
-4
-4
Tube-side fluid
5.08×10
5.35×10
5.70×10
6.28×10
6.57×10-4
Tube-side fouling
1.76×10-4 1.76×10-4 1.76×10-4 1.76×10-4 1.76×10-4
Tube wall
9.23×10-4 9.23×10-4 9.23×10-4 9.23×10-4 9.23×10-4
Shell-side fouling
1.76×10-3 1.76×10-3 1.76×10-3 1.76×10-3 1.76×10-3
Shell-side fluid
1.23×10-2 1.22×10-3 1.57×10-3 1.94×10-3 2.15×10-3
Table 26
58
Table 27
Distributions of the resistances for the stainless steel condensing boiler (%)
A
B
C
D
E
Tube-side fluid
3.4
14.2
13.8
13.8
13.7
Tube-side fouling
1.2
4.7
4.2
3.9
3.7
Tube wall
0.4
1.7
1.6
1.4
1.4
Shell-side fouling
11.9
46.9
42.5
38.5
36.6
Shell-side fluid
83.1
32.4
37.9
42.4
44.7
Table 28 Distributions of the resistances for the carbon steel condensing boiler (%)
A
B
C
D
E
Tube-side fluid
3.2
11.6
11.4
11.6
11.6
Tube-side fouling
1.1
3.8
3.5
3.2
3.1
Tube wall
5.9
20.0
18.5
17.0
16.3
Shell-side fouling
11.2
38.2
35.2
32.4
31.1
Shell-side fluid
78.6
26.4
31.4
35.7
37.9
Zone I
Zone II
Zone III
Zone IV
Table 29 The overall heat transfer coefficients for each zone
Stainless Steel Condenser
Carbon steel condenser
68.5
64.7
253.7
207.6
229.9
191.0
213.4
180.0
3.4 Size of the Condenser and the Pressure Drops
Once the heat transfer rates (Q), LMTD (θLM) and the mean overall heat transfer
coefficients (Um) for each zone have been obtained (Tables 19 and 29), the heat
transfer area (A) for each zone can be calculated from,
Aj =
Qj
(30)
U m, jθ LM , j
where the subscript j refers to the zone number. Table 30 lists the lengths of the
tubes and the total heat transfer areas (based on the outer diameter of the tubes).
Table 30 Size of the condensing boiler
Stainless Steel Condenser
Tube length in Zone I
12.17
Tube length in Zone II
27.87
Tube length in Zone III
14.31
Tube length in Zone IV
8.94
59
Carbon steel condenser
12.90
34.06
17.22
10.61
Total tube length
Surface area, m2
Q/S (kW/m2)
63.30
4918.9
2.34
74.79
5811.5
1.98
The thermal design of the condensing boiler includes the calculation of adequate
surface area to handle the thermal duty for the given specification. Fluid friction
effects in the boiler are also equally important since they determine the pressure drop
of the fluids flowing in the system and, consequently, the pumping power or fan work
input which is necessary to maintain the flow. Provision of pumps or fans adds to
the capital cost and hence it is a major part of the operating costs for the condensing
boiler (Kakac and Liu 2002).
On the tube-side, the frictional pressure drop can be expressed as,
L ρ um2
∆pt = 4 f
di 2
(31)
where L is the tube length, um is the mean fluid velocity and the friction factor f can be
obtained by,
(32)
On the shell-side, the pressure drop depends on the number of tubes the fluid is
passing through in the tube bundle between the baffles as well as the length of each
crossing. The pressure drop can be calculated by,
2
 G  ( N b + 1) Ds
∆ps = f  s 
2 ρ De
 As 
 µw 


 µb 
0.14
(33)
where Nb=L/B-1 is the number of baffles, and the friction factor f on the shell-side is,
(34)
The power (P) of the feed water pump (the tube-side) or fan (the shell-side) is
proportional to the pressure drop (∆p) and can be calculated from
P=
G∆p
(35)
ρη
where G and ρ are the mass flow rate and density of the fluid, respectively, η is the
efficiency of the pump/fan.
Table 31 presents the condensing boiler sizes, pressure drops and pump/fan powers
calculated for the two different construction materials.
60
Table 31
Pressure drop and pump/fan powers for the condensing boiler
Stainless Steel Condenser Carbon steel condenser
Tube-side
Fluid mean temperature, °C
Fluid density, kg/m3
Mean velocity, m/s
Re
Frictional factor
Tube length, m
Pressure drop, Pa
Power of the pump, kW
Shell-side
Fluid mass flow rate, kg/s
Fluid density, kg/m3
As, m2
Shell inner diameter, m
Equivalent diameter, m
Re
Frictional factor
Baffle number
Pressure drop, Pa
Power of the fan, kW
42.6
991.9
0.267
9729
0.0080
63.30
3239.7
0.43
42.6
991.9
0.267
9729
0.0080
74.79
3802.0
0.50
26.3
1.061
1.59
2.09
0.074
57112
0.222
34
28407.2
828.3
26.3
1.061
1.59
2.09
0.074
57112
0.222
41
34088.7
993.4
3.5 Cost Estimation
The cost of the condensing boiler was estimated as part of our case study
calculations. The overall total cost (also known as lifetime costs), associated with a
heat exchanger consists of the capital, installation, operating, and sometimes also
disposal costs. The capital cost includes the costs associated with design, materials,
manufacturing (machinery, labour, and overhead), testing, shipping, installation, and
depreciation. As pointed out above, installation of the heat exchanger at the site can
be as expensive as the capital cost for some types of shell-and-tube heat exchangers.
The operating cost consists of the costs associated with fluid pumping power,
warranty, insurance, maintenance, repair, cleaning, lost production/downtime due to
failure, energy cost associated with the utility (steam, fuel, water) in conjunction with
the exchanger in the network, and decommissioning costs.
It should be noted that it is very difficult to find reliable and accurate cost
estimation data for industrial plants because of confidential nature of these data and
reluctance by the company to release any information (Fraas, 1989). Few significant
cost data have been published in the open literature in recent years (Couper, 2003).
It should also be noted that costs are very sensitive to special requirements.
The following section presents the results obtained from our cost calculations.
61
3.5.1 Capital Costs
Equipment cost data are generally correlated as a function of equipment parameters.
For the industrial condensing boiler, which is actually a condensing heat exchanger,
the capital cost can be correlated to typical capacity parameters such as surface area,
number of passes, etc. A simple correlation of cost data is obtained by the
“six-tenths rule” (Couper, 2003),
(36)
where C1 is the equipment cost for capacity of S1, C2 is the cost for equipment
capacity of S2, n is an exponent varying between 0.3 and 1.2 depending on the type of
equipment. For heat exchangers, n usually is assumed to be 0.68.
For a shell-and-tube heat exchanger, the cost (CE) may be estimated from the
following equation when the pressure, materials of construction or equipment design
type is changed (Couper, 2003),
(37)
where
CB is the base cost of a carbon steel, floating-heads exchanger (10.5MPa design
pressure),
(38)
where A is the heat transfer area between 150 and 12,000ft2
FD is the design-type cost factor if different from that in CB, as shown in Table
32,
Table 32
Design-type cost factor
FMC is the material of construction cost factor,
(39)
where g1 and g2 can be obtained from Table 33.
62
Table 33
Construction material cost factor
Otherwise, if the materials for the shell and tubes are different then the cost factor
can be chosen from Table 34.
Table 34
Material cost factor
FP is the design pressure (psig) cost factor, as shown Table 35.
Table 35
Pressure cost factor
If the pressure is low, then the pressure cost factor can be chosen from Table 36.
Table 36 Pressure cost factor
Couper (2003) stated that to update the costs obtained from the above relationships
to late 2002, it was necessary to multiply the values by a factor of 1.25.
Based on the above relationships, the equipment cost for a condensing boiler (made
from carbon steel) with a working pressure of 16bar (231psig) is approximately
$852,000 (i.e. £568,000). This value may be slightly overestimated. Seamonds et
al. (2009) reported that the cost of a 0.5mmBtu/hr condensing heat exchanger was
$10,000. If this equipment is scaled up to the capacity of the condensing boiler
63
which is being considered in this case study, then the estimated cost is $240,000. In
this case study, polypropylene is used as a liner material. The additional cost for this
corrosion-resistant coating is about $50,000 (from Polymer Plastics Corporation).
For the condensing boiler made from stainless steel, the equipment cost is $2,562,000,
because the material cost for stainless steel 316 is about 3.
The installation cost of the equipment is generally estimated by multiplying the
equipment cost with a multiplier. For the stainless steel heat exchanger, the
multiplier is 1.9 whereas for the carbon steel exchanger it is 2.2.
The installation cost (C, k$) of the induced-draft fan (ID fan) in the shell-side can
be calculated from the following expression,
(40)
where
Q is the gas flow rate in KSCFM;
Coefficients a, b, c can be obtained from Table 37
Table 37
Coefficients of the installed fan cost
fm is the installed factor (as shown in Table 38).
chosen as the fan material.
Table 38
In this case, carbon steel is
Material cost factor
Fp is the pressure factor, as shown in Table 39.
64
Table 39
Pressure cost factor
In this case study a radial centrifugal fan is used to overcome 28 – 34 kPa pressure
drop in the shell-side. The installation cost for this fan is approximately $40,000.
As the pump power in the tube-side is only 0.5kW, the pump cost can be neglected in
the calculations.
3.5.2 Operating and Maintenance Costs
The O&M costs for the condensing boiler in this case study mainly include the
electricity consumption for the pump and fan, chemical treatment for the condensate,
fouling removal in the condensing boiler, etc.
The total power required for the pump and fan are 829kW for the stainless steel
condensing boiler and 994kW for the condensing boiler made from carbon steel.,
Assuming the availability of the system is 7000hrs/year and the non-domestic
electricity tariff is approximately $0.1/kWh (DECC 2010), the annual costs for
electricity consumption are $580,300 for the stainless steel boiler and $695,800 for
the carbon steel boiler.
The condensate from the condensing boiler generally contains nitric acid, halides,
and PMs. The waste needs to be chemically treated before reuse or being disposed
of in an environmentally acceptable manner. The expense for the chemical treatment
is about $0.45/m3 (Spirax Sarco, 2007).
Little useful information is found in the literature about the maintenance costs with
regards to fouling treatment in the heat exchangers. In our case study, an estimated
value of 6% of the fixed capital cost per year was used for fouling treatment costs.
3.5.3 Profitability
The use of a condensing boiler enables the plant to recover an additional 11.5MW
of heat from the flue gas. This recovery can save the plant approximately 5t/h of
wood chips (50%MC and 8.16MJ/kg NCV). Based on the local delivery cost of
£40/tonne (or $60/tonne) for the wood chips, the financial benefit from this fuel
saving is approximately $2,100,000 per year.
65
Table 40
Costs and pay back period for the condensing boiler
Stainless steel condenser
Carbon steel condenser
Capital costs
Boiler cost ($)
Installed factor
Installed boiler cost ($)
Fan cost ($)
In total ($)
O&M costs
Electricity rate ($/kWh)
Electricity ($/year)
Chemical treatment expense ($/m3)
Condensate treatment cost ($/year)
Maintenance cost factor (%)
Maintenance cost ($/year)
Benefit
Wood chips saving (t/h)
Wood chips cost ($/tonne)
Fuel cost saving ($/year)
Payback period (years)
2,562,000
1.9
4,868,000
40,000
4,908,000
852,000+50,000(PP)
2.2
1,984,000
44,000
2,028,000
0.1
580,300
0.45
37,600
6
294,480
0.1
695,800
0.45
37,600
6
121,680
5
60
2,100,000
4.1
5
60
2,100,000
1.7
Table 40 summaries the costs and financial benefits for both types of condensing
boilers. As the stainless steel boiler is more expensive than the carbon steel boiler,
the capital cost for the stainless steel condenser is about 2.5 times higher than the
carbon steel condenser. Due to the higher operating costs (mainly the electricity
consumption for the fan) in the carbon steel condenser, the total operating and
maintenance costs for both condensers are relatively the same. Consequently, the
payback period for the carbon steel condenser (1.7 years) is shorter than the stainless
steel condenser (4.1 years), but due account must be taken of the shorter life of the
carbon steel boiler.
66
4. Conclusions
In this report, the technology and application of industrial condensing boilers in
various heating systems were reviewed. As the ccondensers require site-specific
engineering design, a case study was carried out to investigate the feasibility
(technically and economically) of applying condensing boilers in a large scale district
heating system (40 MW). The main conclusions are as follows:
1.
By recovering the latent heat of water vapour in the flue gas through
condensing boilers, the whole heating system can achieve significantly higher
efficiency levels than conventional boilers.
2.
In addition to waste heat recovery, condensing boilers can also be optimised for
emission abatement, especially for particle removal. The particle separation
mechanisms include inertial impaction and gravitational settling for larger
particles and diffusion for the smallest particles. In addition to Brownian
diffusion, important factors in the removal of fine particles include
thermophoresis, induced by the temperature gradient between the flue gas and
the cool surface, and diffusiophoresis, caused by the steam condensation on
cool surfaces. Furthermore, in condensing scrubbers/heat exchangers, particle
growth by water condensation can affect particle size distributions in the
emission.
3.
Two technical barriers for the condensing boiler application are corrosion and
return water temperatures. Highly corrosion-resistant material is required for
condensing boiler manufacture.
In order to lower the return water
temperature, an under-floor heating system or a high surface area of the
radiators is needed to combine with a condensing boiler in the heating system.
In some cases, heat pumps may be installed with condensing boilers. All
these factors will increase the complexity and the costs of the heating system.
4.
The thermal design of a ‘Case Study’ single pass shell-and-tube condensing
heat exchanger/condenser shows that a considerable amount of thermal
resistance is on the shell-side. This includes fouling, gas phase convective
resistance and vapour film interface resistance. Approximately 4919m2 of
total heat transfer area is required, if stainless steel is used as a construction
material. If the heat transfer area is made of carbon steel, then polypropylene
could be used as the corrosion-resistant coating material outside the tubes.
The addition of polypropylene coating increases the tube wall thermal
resistance, hence the required heat transfer area will be approximately 5812m2
5.
The estimated total capital cost for the condensing boiler ranges from
$2,028,000 (carbon steel) to $4,908,000 (stainless steel). The application of
the condensing boiler increases the energy efficiency, leading to fuel savings of
up to 20%. The payback period is about 2 years for the carbon steel
condenser or 4 years for the stainless steel condenser.
67
6.
The condensing boiler requires a lower water return temperature and should be
used in conjunction with a heat pump or with an under-floor system or larger
radiators for building heating.
68
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