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

I

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

II

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.

III

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

IV

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

1

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

2

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.

3

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

4

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

5

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).

6

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.

7

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

8

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)

9

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).

10

Figure 8 The Bioturbo system: a power plant with pressurised combustion system

2.1.4 Advantages of Condensing Boilers

Latent Heat Recovery

80

85

90

95

100

105

110

115

120

30 50 70 90 110 130 150

Flue gas temperature, oC

Eff

icie

ncy

, %

λ=1.25

λ=1.6

λ=2.0

λ=2.4

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.

11

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)

12

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

13

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).

14

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 P in

SO3

( )2 3 2 3

1000

2.276 0.0294ln 0.0858ln 0.0062lnH O SO H O SOP P P P− − + ×

atm

SO2

( )2 2 2 2

1000

3.9526 0.1863ln 0.000867 ln 0.00091lnH O SO H O SOP P P P− + − ×

mmHg

HCl

( )2 2

1000

3.7368 0.1591ln 0.0326ln 0.00269lnH O HCl H O HClP P P P− − + ×

mmHg

NO2

2 2 2 26 6

1000273

% %3.664 0.1446 ln 0.0827 ln 0.00756 ln ln

100 760 760 10 100 760 760 10

V H O VppmNO V H O VppmNO−

− − +× × × ×

15

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

16

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

17

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

Alloy Type Alloy UNS No. Density

lb/in3

Conductivity

Btu/hr·ft·F

Thermal Expansion

in/in ×10-6/F

Sea-Cure® S44660 0.28 9.5 5.4

AL29-4C® S44735 0.28 9.5 5.2 27-29% Cr

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

18

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

19

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).

20

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.

21

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.

22

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.

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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 Flue gas condenser with a heat pump

Figure 24 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).

24

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).

25

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

26

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

Exit flue gas temperature, °C 20 30 35 40 50 60 100 140

Heat recovered, kJ/Nm3 6115 5896 5626 5291 4873 3665 1802 1210

Heating surface increase, m2 144.8 108.8 95.6 83.2 57.2 36.3 19.0 7.8

Material cost (RMB)

Carbon steel 11595 8713 7655 6663 4582 2909 1524 626

Stainless steel 46382 34850 30621 26653 18327 11636 6094 2502

PTFE 105615 79357 69727 60690 41731 26497 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.

27

Table 4 Savings in fuel cost

Exit flue gas temperature 20 25 30 35 40 50 60 100 140

Heat reclaimed (kJ/Nm3) 6103 5881 5610 5274 4855 3652 1802 1210 609

Saved gas (Nm3/h) 55.8 53.8 51.3 48.3 44.4 33.4 16.4 11.0 5.6

Saved cost (RMB/h)* 83.7 80.7 77.0 72.4 66.7 50.1 24.7 16.6 8.3

*The price of the natural gas is taken as 1.5 yuan RMB/Nm3

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

28

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

(AEC), $/a

1184 1179

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

System Type

Installed

Cost of

System

Annual Energy

Consumption

(Therm Eq.)

Energy

Cost

Maintenance

Cost

Net Present

Cost

Net Present

Cost

Compared to

Conventional

Gas Boiler

Condensing

Gas Boiler $304,015 197,586 $1,446,342 $102,947 $1,853,304 ($397,026)

Conventional

Gas Boiler $246,450 262,670 $1,935,248 $68,631 $2,250,329 -

BCHP** $695,950 404,489 ($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

29

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 Grate

Construction pressure (bar) 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. In the condenser after the boiler,

30

the flue gas temperature decreases to 45°C. The heat output from the flue gas

condenser is 2MWth, as shown in Figure 29 .

(a) (b)

Figure 29 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.

Table 8 Emissions from the plant

Emission Emission amount in 2002 Emission factor

kg/year mg/MJ

NOx 11 605 90

CO 6 776 100

Dust 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.

31

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).

32

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

Flue gas flow Nm3/hr 38000

Emissions

HCl mg/Nm3 dg 4

HF mg/Nm3 dg 0.2

SO2 mg/Nm3 dg 100

Cd mg/Nm3 dg 0.007

Pb mg/Nm3 dg 0.3

Hg mg/Nm3 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.

33

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.

34

Figure 32 Process diagram of the combined heat and power plant

Table 11 Emissions from the combined heat and power plant

Pollutant Emission limits (24-hour average) units

Dust 5 mg/Nm3

HCl 5 mg/Nm3

HF 1 mg/Nm3

SOx 25 mg/Nm3

NH3 5 mg/Nm3

Cd+Tl 0.05 mg/Nm3

Hg 0.03 mg/Nm3

Dioxin 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.

35

Figure 33 Process diagram of the waste to energy plant

Table 12 Performance data of the plant

Waste throughput 26 ton/h

Thermal capacity 74MWth

Flue gas flow 150000Nm3/h (w.g.)

Max extended energy recovery 17MW

Figure 34 Condensing scrubber in the plant

36

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).

37

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

38

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.

30303030

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

39

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 Properties of the fuel input for the heating plant

Dry basis As received

C 51.2 25.6

H 6.1 3.05

N 0.3 0.15

S - -

O 40.9 20.45

Ash 1.5 0.75

Moisture - 50.0

GCV, MJ/kg 20.1 10.05

NCV, MJ/kg - 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

Temperature of the flue gas exit the fluidised bed boiler, °C 150

Temperature of the flue gas to the stack, °C 35

Hot water pressure, bar 16

Hot water temperature, °C 140

Heat capacity of the fluidised bed boiler, MW 40

Temperature of the feed water to the fluidised bed boiler, °C 55

Reference states (Pref, Tref) 1 atm, 25°C

40

Figure 39 Process diagram for the heating plant

Table 15 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

41

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

42

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

43

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.

44

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%

45

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),

( )outss

t

s

intt hhG

Ghh ,, −+=

(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,

( ) ( ) ( )[ ]swrefswprefsgps TiTTcxTTch +−×+−= .0,

(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,

( ) ( ) ( )[ ] ( ) ( )15.2730,, −−++−×+−= swsswrefswpsrefsgps TcxxTiTTcxTTch

(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,

( )15.273−= twt Tch

(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

46

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.

0

20

40

60

80

100

120

140

160

0 50 100 150 200 250 300 350 400 450

Specific enthalpy difference (h -h outlet) on the shell side, kJ/kg

Tem

per

atu

re,

oC

Flue gas (shell side)

Return water (tube side)

Zone IZone IIZone IIIIV

A

B

CD

E

Figure 42 Condensation curve for the flue gas

Table 18 Boundaries of the four zones in Figure 42

Zone Boundaries Shell-side Tube-side

A 150.0 55.0

B 64.3 49.3

C 50.0 37.0

D 40.0 32.0

E 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,

47

(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,

48

(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)

49

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 for each zone

LMTD F Heat transfer rate, MW

Zone I 43.3 0.971 2.652

Zone II 14.0 0.865 5.757

Zone III 10.3 0.935 2.296

Zone IV 6.2 0.966 0.865

3.3.4 Heat Transfer Coefficients

50

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

51

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

52

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),

53

(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

i

cb

cd

Nuh

λ=

(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 D E

Temperature, °C 55 49.3 37.0 32.0 30.2

Velocity, m/s 0.268 0.267 0.267 0.266 0.265

Density, kg/m3 986.3 989.0 991.8 994.9 996.0

Heat capacity, kJ/kgK 4.18 4.18 4.18 4.18 4.18

Viscosity, PaS 5.04×10-4 5.54×10-4 6.09×10-4 7.26×10-4 7.79×10-4

Thermal conductivity, W/mK 0.650 0.643 0.627 0.623 0.618

Pr 3.24 3.59 4.06 4.86 5.26

Re 1.2×104 1.1×104 9.9×103 8.3×103 7.8×103

f 0.0075 0.0077 0.0079 0.0083 0.0085

Nu 69.4 66.6 64.1 58.5 56.4

Tube-side coefficient, W/m2K 1968.4 1870.5 1754.2 1591.6 1522.0

54

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),

34.063.0 PrRe27.0 gDe

g

eg DhNu ==

λ

(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.

55

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.

56

Table 22 Shell-side heat transfer coefficients for the gas stream

A B C D E

Temperature, °C 150.0 64.3 50.0 40.0 35.0

Flow rate, Nm3/s 21.3 21.3 18.5 17.5 17.2

Density, kg/m3 0.796 0.998 1.097 1.157 1.185

Heat capacity, kJ/kgK 1.176 1.141 1.069 1.039 1.028

Viscosity, PaS 2.08×10-5 1.73×10-5 1.76×10-5 1.75×10-5 1.74×10-5

Thermal conductivity, W/mK 0.032 0.026 0.025 0.025 0.025

Equivalent diameter (De), m 0.074 0.074 0.074 0.074 0.074

Crossflow area (As), m2 1.59 1.59 1.59 1.59 1.59

Gas velocity, m/s 13.4 13.4 11.7 11.0 10.8

Pr 0.77 0.77 0.74 0.73 0.73

Re 3.8×104 5.7×104 5.4×104 5.4×104 5.4×104

Nu 188.1 244.0 231.5 230.2 230.9

Coefficient (hg) , W/m2K 81.1 84.7 79.4 77.7 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 Overall shell-side heat transfer coefficients

B C D E

ilg, MJ/kg 2.35 2.38 2.41 2.42

57

Density, kg/m3 981.1 988.0 992.2 994.0

Viscosity, PaS 4.40×10-4 5.47×10-4 6.53×10-4 7.20×10-4

Thermal conductivity, W/mK 0.658 0.644 0.631 0.623

Tsat-Tw 15.01 13.03 7.95 4.81

Nu1 391.5 388.9 425.0 472.9

h1, W/m2K 10143 9855 10553 11604

hN, W/m2K 4264.6 4143.4 4437.0 4879.0

Z 0.083 0.106 0.133 0.150

Shell-side heat transfer

coefficient, heff, W/m2K

821.0 636.7 516.0 466.0

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.

58

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

Thermal resistance, m2K/W 6.55×10-5 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

Tube-side fluid 5.08×10-4 5.35×10-4 5.70×10-4 6.28×10-4 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.

Table 26 Thermal resistances in the carbon steel condensing boiler (unit: m2K/W)

A B C D E

Tube-side fluid 5.08×10-4 5.35×10-4 5.70×10-4 6.28×10-4 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

59

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

Table 29 The overall heat transfer coefficients for each zone

Stainless Steel Condenser Carbon steel condenser

Zone I 68.5 64.7

Zone II 253.7 207.6

Zone III 229.9 191.0

Zone IV 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,

, ,

j

j

m j LM j

QA

U θ=

(30)

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 Carbon steel condenser

Tube length in Zone I 12.17 12.90

Tube length in Zone II 27.87 34.06

Tube length in Zone III 14.31 17.22

Tube length in Zone IV 8.94 10.61

60

Total tube length 63.30 74.79

Surface area, m2 4918.9 5811.5

Q/S (kW/m2) 2.34 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,

2

42

mt

i

uLp f

d

ρ∆ =

(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 0.14

( 1)

2s b s w

s

s e b

G N Dp f

A D

µ

ρ µ

+∆ =

(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

ρη

pGP

∆=

(35)

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.

61

Table 31 Pressure drop and pump/fan powers for the condensing boiler

Stainless Steel Condenser Carbon steel condenser

Tube-side

Fluid mean temperature, °C 42.6 42.6

Fluid density, kg/m3 991.9 991.9

Mean velocity, m/s 0.267 0.267

Re 9729 9729

Frictional factor 0.0080 0.0080

Tube length, m 63.30 74.79

Pressure drop, Pa 3239.7 3802.0

Power of the pump, kW 0.43 0.50

Shell-side

Fluid mass flow rate, kg/s 26.3 26.3

Fluid density, kg/m3 1.061 1.061

As, m2 1.59 1.59

Shell inner diameter, m 2.09 2.09

Equivalent diameter, m 0.074 0.074

Re 57112 57112

Frictional factor 0.222 0.222

Baffle number 34 41

Pressure drop, Pa 28407.2 34088.7

Power of the fan, kW 828.3 993.4

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.

62

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.

63

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

64

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). In this case, carbon steel is

chosen as the fan material.

Table 38 Material cost factor

� Fp is the pressure factor, as shown in Table 39.

65

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.

66

Table 40 Costs and pay back period for the condensing boiler

Stainless steel condenser Carbon steel condenser

Capital costs

Boiler cost ($) 2,562,000 852,000+50,000(PP)

Installed factor 1.9 2.2

Installed boiler cost ($) 4,868,000 1,984,000

Fan cost ($) 40,000 44,000

In total ($) 4,908,000 2,028,000

O&M costs

Electricity rate ($/kWh) 0.1 0.1

Electricity ($/year) 580,300 695,800

Chemical treatment expense ($/m3) 0.45 0.45

Condensate treatment cost ($/year) 37,600 37,600

Maintenance cost factor (%) 6 6

Maintenance cost ($/year) 294,480 121,680

Benefit

Wood chips saving (t/h) 5 5

Wood chips cost ($/tonne) 60 60

Fuel cost saving ($/year) 2,100,000 2,100,000

Payback period (years) 4.1 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.

67

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.

68

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.

69

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