Steam Power Stations for Electricity and Heat Generation
4.4 Possibilities for Efficiency Increases in the Development of a Steam Power Plant
4.4.1 Increases in Thermal Efficiencies
4.4.1.3 Lower Heat Dissipation Temperatures – Optimisation of the “Cold End”
4.4 Possibilities for Efficiency Increases in the Development of a Steam Power Plant 151 The temperature differences between the heat-dissipating and the heat-absorbing flows in a preheating stage are characterised by the so-called terminal temperature difference (TTD), which is defined as the smallest temperature difference between the two mediums. At the transition to small TTDs, larger heating surfaces and hence heavy and expensive plant components are required. A compact construction is the result when counterflow heat exchangers are used.
Preheaters are usually designed as shell-and-tube heat exchangers. The extracted steam and the feed water are segregated from each other by a heat exchanger sur- face, which consists of tube bundles. The sensible heat of the steam can be utilised in so-called desuperheaters. The sensible heat of the condensate can be utilised in condensate coolers, which can be mounted either by integration into the preheaters or separately. The desuperheater, with respect to the feed water, is mounted after the preheater(s). This way, the feed water can be heated to a higher temperature than is possible with the condensing preheater. The condensate cooler, with respect to the feed water, is mounted before the preheater.
The most reasonable solution in terms of thermodynamics is to mix, without cooling, the condensate in the preheater with the feed water. This method is not used for HP preheaters because the high feed water pressure requires a complex system of pumps, pipes and fittings. Thermodynamically, it is therefore a compromise to subcool the condensates and to let them flow into the next lowest preheat stage. In configurations with multistage LP preheaters, it is usually economical to pump the condensates of one or several preheat stages back into the condensate flow.
In a direct-contact heater, the heat of the extracted steam is transferred to the feed water by mixing and condensation of steam in water. Given its low terminal temperature difference, the direct-contact heater has thermodynamic advantages.
However, because the container is under the pressure of extraction, the entire condensate flow has to be pumped to reach the corresponding pressure level.
Because of the necessary pumps, direct-contact heaters are only used in the feed water tank for deaeration.
The common values for the terminal temperature differences of regenerative heaters of modern hard coal power plants are (STEAG 1988) as follows:
• Desuperheater 25 K
• Condensation equipment 2 K
• Condensate cooler 7 K
4.4.1.3 Lower Heat Dissipation Temperatures – Optimisation
152 4 Steam Power Stations for Electricity and Heat Generation Fig. 4.60 Impact of a heat
dissipation temperature reduction of 1 K
Besides the mean heat input temperature into the steam generator, the mean heat dissipation temperature is another factor which determines the thermal efficiency of the cycle. This temperature has to be chosen to be as low as possible in order to maximise the total efficiency. Low exhaust-steam temperatures and pressures in the condenser can be set by low temperatures of the cooling medium. The steam can be further expanded to the low exhaust steam pressure by the final LP blading. More heat is converted into mechanical work and thus the waste heat cut down by this heat fraction.
The heat dissipation temperature has an impact on the efficiency, which increases in strength when the heat input temperature is lower. This correlation is shown in Fig. 4.60 for the Carnot cycle, with a mean heat dissipation temperature of 30◦C, corresponding to the condensation temperature. These fundamental correlations also hold true for other thermal power processes. Therefore it is evident that in a pure steam process, in comparison to a combined-cycle (gas and steam turbine) pro- cess, a higher efficiency increase can be achieved by improvements at the cold end (Joh¨anntgen 1998). For the reference power plant, with a mean heat input temper- ature of 376◦C, a decrease in the condensation temperature of 1 K diminishes the heat rate by 0.29% in the ideal case.
In a given turbine unit, the steam outlet velocity rises with an increase in the specific volume, i.e. when the condenser pressure decreases. Compared to the isen- tropic expansion, changes in the condenser pressure cause less change in the heat rate. With the condenser pressure decreasing, the losses increase through the kinetic energy of the exhaust steam, due to the rising outlet velocity. If sonic velocity is reached at very low condenser pressures, a further decrease of the condenser pressure does not improve the efficiency (Adrian et al. 1986). The losses in the exhaust steam are taken into account by the internal efficiency of the turbine.
The optimisation of the cold end must therefore involve not only the design of the cooling circuit but also the choice of the low-pressure turbine. In order to make use of the efficiency advantage of low condenser pressures, it is necessary to enlarge the exhaust steam cross-section of the LP turbine. As well as developing and
4.4 Possibilities for Efficiency Increases in the Development of a Steam Power Plant 153
G
a b
G
Wet type cooling tower (evaporation cooling)
c
G
M
Dry cooling (direct condensation)
d
G
Dry cooling (indirect)
Fig. 4.61 Cooling systems in power plant technology (Baehr 1985)
utilising greater LP last-stage blade lengths – today, last-stage blades are manufac- tured with lengths up to 1,400 mm, at 3,000 r/min (revolutions per minute) (Neft and Franconville 1993) – a larger outlet cross-section can be obtained with two, three, or four LP turbine components mounted on a shaft. Increasing the number of LP turbines, however, entails step changes in the costs for turbines and turbine houses (STEAG 1988; Weber et al. 2005).
The temperature of heat dissipation is set by the cooling method. Schematics of the cooling systems are shown in Fig. 4.61 (Baehr 1985). Basically, there are three cooling systems to dissipate the waste heat arising in the condenser to the ambient air:
Once-Through Water Cooling
When fresh-water or once-through water cooling with river or seawater is used, the heat in the condenser is directly transferred to the cooling medium. Once-through cooling is simple and effective but can be utilised only at locations where there is fresh water available in sufficient quantities and the inevitable temperature rise eco- logically justifiable. In Germany, river water temperatures have an annual average of 12◦C; power plants at coastal locations in Denmark are based on a mean seawater temperature of 10◦C.
Back-Cooling of Cooling Water Through Evaporation
In Germany, new plant designs mostly incorporate closed-circuit cooling water sys- tems with natural-draught cooling towers. In such systems, the waste heat is initially
154 4 Steam Power Stations for Electricity and Heat Generation transferred to the cooling water in the condenser and then backcooled in a cooling tower by heat dissipation to the cooling air. In this process, water is lost through evaporation and has to be replaced. The cooling level theoretically achievable with a wet-type cooling tower is determined by the wet-bulb temperature.1This tempera- ture depends on the condition of the air and may lie below the cold inlet air, because of the extraction of evaporation heat (Berliner 1975; Schmidt et al. 1977). Given an annual average temperature of air of 8.5◦C (Germany) and a relative air humidity of 75%, the resulting theoretically possible cooling is 6.6◦C (STEAG 1988). Though this temperature is below the annual average of rivers, the cold water temperature that is economical, and therefore used, in back-cooling is around 15–20◦C.
Dry Cooling
In direct dry cooling, the condenser is directly cooled by ambient air. In indirect dry cooling, an additional water circuit is used, and the warmed cooling water is cooled again in an air/water heat exchanger. For dry cooling systems, it is the dry bulb temperature that sets the temperature difference between the saturated air and the (approach) cooling water, whereas the theoretical limit for wet cooling towers is set by the lower wet-bulb temperature. Depending on conditions at the location, the difference between the dry and wet-bulb temperatures can amount to 15◦C (at high temperatures and low air humidities). The poor heat transfer in air requires large heat exchange surfaces and therefore raises the economically achievable cold water temperature. Since, in contrast to wet cooling, dry cooling uses only convection, an air mass flow is necessary which is four times higher than the one in a wet cooling tower. These factors lead to higher exhaust steam temperatures and in consequence higher average heat dissipation temperatures, compared to evaporation cooling. Dry cooling is used only where the additional water required for wet cooling is not avail- able. Indirect dry cooling involves investment costs that are about three times as high as a wet cooling system (Henning 1985).
Hybrid-Type Cooling
In hybrid-type cooling towers, both wet cooling and dry cooling are used. This method combines the advantage of the high cooling efficiency of wet cooling with the advantage of dry cooling, i.e. the absence of water vapours (Sauer 1984). In the variant usually used, the air flow is divided. One part of the flow is used for dry cooling, the other for wet cooling. By mixing the partial air flows, one obtains a wet vapour-free cooling tower exhaust – i.e. the exhaust is not visible. The water to be cooled is first conducted through the dry section and then through the wet section.
1The wet bulb temperature is the temperature measured by a moist thermometer or psychrometer.
The thermometer is wrapped with a moist fabric. Water evaporates depending on the humidity and temperature of the air. The lower the air humidity and the higher the temperature, the higher the evaporation heat and hence the difference between dry and wet-bulb temperatures. The wet-bulb temperature is used in meteorology to determine the relative air humidity.
4.4 Possibilities for Efficiency Increases in the Development of a Steam Power Plant 155
Optimum values in Germany and Denmark - Once-through cooling - Seawater (10°C) 20–25 mbar - River water(12°C) 30–35 mbar - Evaporation 35–40 mbar
0.14 0.12 0.1 0.08 0.06 0.05 0.04 0.03
0.02 0.2
Condenser pressure [bar]
Evaporative cooling
0.035
Dry cooling indirect
0.07 0.15
Dry cooling direct
Once-through cooling
0.02 0.06
0.1
0.12 0.15
0.2
Fig. 4.62 Achievable condenser pressures in different cooling systems (Baehr 1985)
Wet cooling is usually exclusively used during summer operation, with the mixed use occurring in winter. The investment costs of this technique amount to three times as much the costs for a wet cooling tower, and its cooling characteristics resemble that of the wet-type cooling tower (Henning 1985).
Figure 4.62 shows the exhaust steam pressures achievable by the different cool- ing methods. It becomes evident that the chosen cooling technique has a substan- tial influence on the condensation temperatures and exhaust steam pressures. The ranges given in Fig. 4.62 are functions of the location-dependent air and water temperatures. Systems using once-through cooling thus offer favourable, systems with dry cooling unfavourable conditions for attaining a high thermal efficiency.
Evaporative cooling, in general, involves higher condensation temperatures than once-through cooling, though clearly lower temperatures than dry cooling sys- tems (Baehr 1985). In Denmark, condenser pressures between 20 and 25 mbar are achieved in advanced steam cycles with seawater cooling at an annual average of about 10◦C. Reports on power plants with wet cooling towers mention condenser pressures between 35 and 40 mbar (Meier 2004; Lambertz and Gasteiger 2003;
Tremmel et al. 2006; Mandel and Schettler 2007; Billotet and Joh¨anntgen 1995;
Eichholtz et al. 1994). The reference values for river water cooling in Germany range around 30 mbar.
The impacts of the condenser pressure on the net efficiency is shown in Fig. 4.63 for a power plant with conventional and with advanced steam conditions (Adrian et al. 1986; Kjaer 1993). Evaporative cooling, compared to seawater cooling, has a disadvantage in efficiency of about 1–1.5%, yet an advantage of greater than 1%
compared to dry cooling.
The seasonal fluctuations of water and/or air temperatures have a direct effect on the exhaust steam quality in the condenser and hence on the thermal efficiency too. Figure 4.64 shows the yearly trend of cold water temperatures for the cases of seawater cooling and evaporation cooling (Joh¨anntgen 1998).
156 4 Steam Power Stations for Electricity and Heat Generation
Fig. 4.63 Impact of the condenser pressure on the net efficiency (Adrian et al. 1986; Kjaer 1993)
Fig. 4.64 Yearly trend of cold water temperatures (Joh¨anntgen 1998)
Lower exhaust steam pressures in winter have less of an effect on the efficiency, however, than the rise of the condenser pressure in summer, because the outlet loss of the turbine increases with descending pressure. Exhaust steam qualities which are lower than those designed for can also be limited by the allowable exhaust moisture.
Wet cooling towers might also confer restrictions on the cold water temperature, for example that they should not fall below 12◦C, to prevent icing (Adrian et al. 1986).
A temperature rise of 22◦C to a level 30◦C above the design temperature of 8◦C of a
4.4 Possibilities for Efficiency Increases in the Development of a Steam Power Plant 157 Fig. 4.65 Influence of
ambient conditions on efficiency (Eichholtz et al.
1994)
natural-draught cooling tower deteriorates the efficiency of a modern power plant by 1.8%. Ambient temperatures of−10◦C yield an improvement of only about 0.2%
(see Fig. 4.65) (Eichholtz et al. 1994). The impact depends on the turbine design (Weber et al. 2005). The conditions on site and the legislation concerning water rights and urban planning and building laws set criteria which narrow the choice of thermodynamically reasonable cooling techniques at the cold end of the power plant. In Germany, natural-draught cooling towers for waste heat dissipation and back-cooling of the heat carrier, i.e. cooling water, have become standard. With all cooling systems, the difference between the condensation temperature and the cooling medium temperature for the heat dissipation has to be kept as small as possible.
Back-Cooling by Natural-Draught Cooling Tower
Figure 4.66 presents the schematic diagram and the design data of a closed-circuit system with a natural-draught cooling tower for a 720 MW hard coal power plant (Baehr 1985).
In the condenser, the cooling water gets heated from an inlet temperature tW1= 20◦C up to 34.5◦C. For back-cooling, the water is transported to the cooling tower where, over about 10–15 m height, it is sprayed through nozzles that are located around the cooling tower cross-section.
The cooling water falls and disperses, via distribution plates, onto the fill packing, which it flows through, then dropping down into the cooling tower basin. In coun- terflow to the rising cooling air, the water cools to a temperature of tW1=20◦C both by convection and by evaporation, whereupon it is returned to the condenser. In the example shown in Fig. 4.66, 30% of the cooling efficiency is achieved by convective cooling with air and 70% by evaporation.
Ambient air at a temperature of tA1 = 8.5◦C and with a relative humidity of 76.3% flows into the cooling tower, where its temperature rises to tA2=27.1◦C. By warming and buoyancy of the air, a convective flow forms in the cooling tower – this
158 4 Steam Power Stations for Electricity and Heat Generation
Power plant cooling tower circuit
Condenser temperatures
Condenser
G~
Mist eliminators
Fill packing
Tower shell Cooling tower temperatures
C CT S V W L A dA MU B
= Condenser
= Cooling tower
= Steam
= Vapour
= Water
= Losses
= Air
= dry air
= Make-up water
= Blow down water tC=36°C
tW1=20°C
tW2=34.5°C
~720 MW
m.W2 =15555 kg/s
m.MW =m.V+m.B
m.B ~ 0.01 m.W1 mW =mW2–mV
. . .
DCT=96.5 m
ρA1=76.3 % xA1=5.2 g/kg m.dA =14847 kg/s HCT=128 m WCT=5 m/s
xCT=23 g/kg
m.A=15194 kg/s; v.A=13098 m3/s tW1 =20°C tA1 =8.5°C tW2 =34.5°C
tA2 =27.1°C
m.V=0.024 m.W1 Boiler
Spray nozzles
dTS=54 m
Fig. 4.66 Wet tower cooling circuit with design data for a 720 MW hard coal fuelled power station (Baehr 1985)
defines a natural-draught cooling tower. With increasing humidity, the flowrate slows. Since the cooling power depends on the air mass flow, the air flowrate can be forced much higher by ventilators (a ventilator cooling tower), the driving power demand of which increases the auxiliary power requirement of the plant.
The wet-bulb temperature of the ambient air tWB, which in the example is 6.6◦C, is the physical limit for the mean condensation temperature tC and hence for the efficiency improvement at the cold end. For plants with wet cooling towers, the difference between the wet-bulb temperature of the ambient air and the economic temperature, or the mean condensation temperature (here 36◦C) results from (see Fig. 4.67) (Odenthal and Spangenmacher 1959)
• The cooling range (tW2−tW1): This is the temperature rise of the cooling water in the condenser from tW1to tW2, which is determined by the cooling water mass flow with a given heat dissipation. In the cooling tower, the cooling water is cooled back to its temperature tW2 before entry to the condenser. In the given example, the difference is tW2−tW1=14.5◦C.
• The terminal temperature difference (TTD) of the condenser: In the example, the difference is tC−tW2=1.5◦C.
• The approach tW1−tWBof the cooling tower: This is the temperature difference between the temperature of the backcooled water and the theoretically possible cold water temperature, which equals the wet-bulb temperature. In the example, the difference is tW1−tWB=13.4◦C.
4.4 Possibilities for Efficiency Increases in the Development of a Steam Power Plant 159 Fig. 4.67 Temperature
relations in circuit cooling systems by wet cooling tower
Diminishing the cooling range, the approach or the TTD of the condenser by 1 K results, for each of these parameters, in an equal lowering rate of the condensation temperature (STEAG 1988).
A smaller cooling range is achieved by a greater cooling-water mass flow. The lower water outlet temperature after the condenser then decreases the condensation temperatures correspondingly, with the same TTD maintained.
A greater cooling-water mass flow requires, for the heat and mass transfer, a greater surface for the cooling water to flow down, which is achieved by appropriate inserts, increasing the surface area. In the case of natural-draught cooling towers, enlarging the transfer surfaces as a rule involves the enlargement in height and diameter of the body as well. In the case of ventilator cooling towers, the power demand of the ventilators increases. In the condenser, narrowing the cooling range causes a reduction of the mean logarithmic temperature difference so that, at the same TTD, larger condenser surfaces are needed. The cooling range values common in Germany are between 16 and 10 K, the latter holding true for plants currently in planning.
Small approaches in the cooling tower can be achieved with larger transfer sur- faces. In the extreme case of the ideal cooling tower – which only exists theoreti- cally – the water is cooled down to the wet-bulb temperature, and the approach is then tW1−tWB=0. Such an ideal cooling tower has to function by counterflow and has an infinitely large transfer surface (Klenke 1966). Commonly, approaches are between 8 and 12 K.
Smaller terminal temperature differences (TTDs) in the condenser are achieved with larger condenser surfaces. Commonly, condenser TTDs are about 1–2 K.
160 4 Steam Power Stations for Electricity and Heat Generation Fig. 4.68 Thermodynamic
comparison between parallel- and series-connected partial condensers, both with the same condenser surface (STEAG 1988)
The above-described measures to reduce the condenser temperature necessitate additional investments in the LP turbine, the condenser and the cooling tower.
The reduction of the mean condenser temperature, and hence of the heat rate, while maintaining the same condenser surface area, is possible by implementing water-side series connections of partial condensers. Figure 4.68 demonstrates the advantage of such a configuration in comparison to the often-found parallel connec- tions of partial condensers (STEAG 1988).
Water losses arising through evaporation and blowdown have to be balanced out by additional water. Water loss through evaporation depends on the humidity in the air. Blowdown is necessary in order to prevent minerals contained in the cooling water from accumulating. The entire additional demand for water to account for these losses lies in the order of magnitude of 2–3% of the cooling water mass flow (Baehr 1985).
Pollution of the cooling air and residual contamination of the pre-treated cooling tower make-up water lead to foul deposits in the cooling circuit, which eventually settle at the bottom of the cooling tower basin. This cooling tower slurry is collected over long operating periods and cleaned up during an outage, after drainage of the basin. Until the slurry settles, it is carried along in the cooling cycle. In consequence, deposits form on the inside of the condenser tubes, which deteriorate to a consid- erable extent the heat transfer. An effective remedial action is constant condenser cleaning by a service system, which carries a number of calibrated cleaning bodies that pass through the tubes, such as sponge rubber balls.