EP0876569A2 - Continuous steam generator - Google Patents

Abstract

The invention concerns a continuous steam generator (2) having a combustion chamber (4) with vertically extending pipes (12) which have a surface structure (26) on the interior. A flow medium (S) flows upwards through the pipes (12). According to the invention, a particularly advantageous mass flow density m in the pipes (12), at a load at which critical pressure (pcrit) prevails therein, corresponds, according to the invention, to the relation (7).

Description

description

Continuous steam generator

The invention relates to a once-through steam generator with a combustion chamber surrounded by a surrounding wall of pipes which are connected to one another in a gas-tight manner, wherein the pipes which run vertically and have a surface structure on the inside thereof can be cut by a flow medium from bottom to top.

Such a steam generator is known from the article "Evaporator Concepts for Benson Steam Generators" by J. Franke, W. Köhler and E. Wittchow, published in VGB Kraftwerkstechnik 73 (1993), Issue 4, pp. 352 to 360 Such a continuous steam generator leads to the heating of the combustion chamber forming evaporator tubes, in contrast to a natural circulation or forced circulation steam generator with only partial evaporation of the water / water / steam mixture circulated, to a complete evaporation of the

Flow medium in the evaporator tubes in one pass. While in the natural circulation steam generator the evaporator tubes are in principle arranged vertically, the evaporator tubes of the continuous steam generator can be arranged both vertically and spirally - and thus inclined.

A continuous steam generator, the combustion chamber walls of which is constructed from vertically arranged evaporator tubes, is less expensive to manufacture than a continuous steam generator having a spiral-shaped tube. Continuous-flow steam generators with vertical pipes also have lower water / steam-side pressure losses than those with inclined or spirally rising evaporator pipes. Furthermore, in contrast to a natural circulation steam generator, a continuous steam generator is not subject to any pressure limitation, so that live steam pressures far above the critical pressure of water (pk _n . = 221 bar) - where there is only one ring density difference between liquid-like and vapor-like medium - are possible. High live steam pressures are required in order to achieve high thermal efficiencies and thus low CO ₂ emissions.

A particular problem is the design of the combustion chamber or peripheral wall of the once-through steam generator with regard to the pipe wall or material temperatures that occur there. In the subcritical pressure range up to about 200 bar, the temperature of the combustion chamber wall is essentially determined by the level of the saturation temperature of the water if wetting of the heating surface in the evaporation area can be ensured. This will e.g. achieved by using inner finned tubes. Such pipes and their use in steam generators are e.g. B. is known from European patent application 0 503 116. These so-called finned tubes, d. H. Pipes with a ribbed inner surface have a particularly good heat transfer from the inner wall to the flow medium.

In the pressure range of approximately 200 to 221 bar, the heat transfer from the inner tube wall to the flow medium drops sharply, so that the flow velocity - as a measure of which the mass flow density is usually used - must be increased accordingly in order to ensure adequate cooling of the tubes. Therefore, in the evaporator tubes of continuous steam generators that are operated at pressures of approx. 200 bar and above, the mass flow density and thus the friction pressure loss must be selected higher than in the case of continuous steam generators which are operated at pressures below 200 bar. As a result of the higher frictional pressure loss, the advantageous property of the vertical pipe is lost, particularly in the case of small inner pipe diameters, that the throughput also increases when individual pipes are heated. However, since high vapor pressures above 200 bar are required in order to achieve high thermal efficiencies and thus low CO ₂ emissions, it is necessary to have one in this pressure range as well ensure good heat transfer. Continuous steam generators with a vertically touched combustion chamber wall are therefore usually operated with relatively high mass flow densities in the pipes in order to always have a sufficiently high heat transfer from the pipe wall to the flow medium, ie to the water / water, in the unfavorable pressure range of approximately 200 to 221 bar -Vapor mixture to achieve. For this purpose, in the publication "Thermal Engineering" IE Semenovker, Vol. 41, No. 8, 1994, pages 655 to 661, both for gas and for coal-fired steam generators, a mass flow density at 100% load is uniform with about 2000 kg / m ² s specified.

The invention is based on the object of specifying a design criterion suitable for a particularly favorable mass flow density in the pipes for pipes of a peripheral wall of a once-through steam generator.

This object is achieved in that the steam generator is designed in such a way that the mass flow density m in the tubes of the surrounding walls at the load at which critical pressure Pkrit prevails in the tubes, the relationship:

m = - (kg / m's)

C (Tmax - Tkrit - ΔTw) ^{v 6} '

corresponds to, where qi (kW / m ² ) the heat flow density on the inside of the pipe, Tmax (° C) the maximum permissible material temperature of the pipe, Tkrit (° C) the temperature of the flow medium at critical pressure picnt,

Δ T _w (K) is the temperature difference between the outer and inner wall of the tube, and C ≥ 7.3 * 10 ^"3 kWs / kgK is a constant.

The invention is based on the consideration that for the fluidic design of the internally finned pipes because the mass flow density had to meet two fundamentally contradictory conditions. On the one hand, the average mass flow density in the pipes should be chosen to be as low as possible. This is to ensure that a higher mass flow flows through individual pipes, to which more heat is supplied than other pipes due to unavoidable differences in heating, than pipes which are heated on average. This natural circulation characteristic known from the drum boiler leads to an equalization of the steam temperature and thus the pipe wall temperatures at the outlet of the evaporator heating surface.

On the other hand, the mass flow density in the pipes must be chosen so high that reliable cooling of the pipe wall is ensured and permissible material temperatures are not exceeded. In this way, high local overheating of the pipe material and the associated damage (pipe ripper) are avoided. In addition to the temperature of the flow medium, the main influencing variables for the material temperature are the external heating of the pipe wall and the heat transfer from the inner pipe wall to the flow medium (fluid). There is thus a connection between the internal heat transfer, which is influenced by the mass flow density, and the external heating of the tube wall.

The invention is now based on the knowledge that the relationship between the inner minimum heat transfer coefficient αmin and the mass flow density m is in an admissibly simplified form by the relationship: can be described, where αmm (kW / m ² K) the heat transfer coefficient, m (kg / m ² s) the mass flow density in the finned tubes, and C a constant with the average value C = 7.3 * IO ^"3 kWs / kgK for commercially available Depending on the structure of the inner surface of the pipes, this constant is C. also to be selected in the range between 7.3 * IO ^"3 kWs / kgK and 12 ^• 10 ³ kWs / kgK.

The relationship mentioned gives an optimal mass flow density in the pipes, which results both in a favorable flow characteristic (natural circulation characteristic) and also ensures reliable cooling of the pipe wall and thus compliance with the permissible material temperatures.

A fundamental consideration when deriving the above-mentioned relationship for the mass flow density in the pipes is that with a given external heating of the pipe wall - in the following the so-called heat flow density (kW / m ² ), ie the heating, is used for this per unit area used - the material temperature of the pipe wall is only slight, but is definitely below the permissible value. The physical appearance must be taken into account that the heat transfer from the inner tube wall to the flow medium is the most unfavorable in the critical pressure range of approximately 200 to 221 bar.

Extensive investigations show that the highest material stress is reached when a relatively low mass flow density is combined with the greatest heat flow density in the evaporation area at about 200 to 221 bar. This is e.g. in the area of the combustion chamber in which the burners are arranged. When the evaporation has ended and the steam overheating begins, the material stress on the pipes of a combustion chamber wall drops again. The reason for this is that the heat flow density also decreases with the usual burner arrangement and the usual combustion process.

Furthermore, it was found that there are no heat transfer problems in other pressure ranges if sufficient cooling of the tube wall is ensured in the pressure range of 200 to 221 bar when using finned tubes. is steady. Thus, at low pressures, ie below approx. 200 bar, the internal ribbing of the pipes means that the sieve crisis only begins at the end of the evaporation zone, ie in a region with a reduced heat flow density. There is no longer a boiling crisis in the supercritical printing area. The

Heat transfer is now so intense that adequate cooling of the tube wall is ensured.

To determine the optimal mass flow density m in the tubes of the tube wall, which on the one hand ensures an advantageous flow characteristic and on the other hand ensures reliable cooling of the tube wall, the following procedure can be followed:

Step 1:

Determination of the heat flow density g _a on the outside of the pipe on the basis of the thermal calculation for the load at which there is a pressure of 210 bar in the pipes of the pipe wall. This heat flow density determined in this way must be increased by a factor between 1.1 and 1.5 in order to take account of local irregularities in the heat transfer.

Step 2 :

Calculation of the maximum permissible material temperature T _max at the top of the pipe on the heated side of the pipe wall. If one assumes that the enclosure or combustion chamber wall has an average temperature that corresponds to the average of T, ™ * and T _κr i _t , the maximum thermal stress is calculated as:

Omax = ^Tma * - ^Tkrit _B , _E ( _{N / mm} ,) ₍₍₂ )

with σ _max maximum thermal stress (N / mm ² )

T _max Maximum material temperature (° C) T _{kri t} Temperature of the fluid at the critical point (° C) ß Thermal expansion coefficient (1 / K) E modulus of elasticity (N / mm ² )

Since the relevant stresses here are thermal stresses, they can be secured as secondary stresses in accordance with the ASME code with three times the permissible stresses σ _2u ι. This results in the temperature T _max

Tmax = Tkrit + ^{6 'Gml} (° C) (3)

The permissible voltage can be found in the pipe manufacturer's specifications.

Step 3: Conversion of the specified heat flow density q _a (related to the outside of the pipe wall) to a heat flow density q _{i (} related to the inner wall of the pipes:

The determination of the heat redistribution factor K is based on temperature field calculations and can be determined with sufficient accuracy as follows:

K = A (da ² - qa) + B (5) with A = 0.45 and B = 0.625 for (da ² - qa) <0.5 kW and A = 0.25 and B = 0.725 for (da ² - qa)> 0.5 and <1.1 kW and A = 0 and B = 1 for (da ² - qa)> 1.1 kW, with d _a = pipe outer diameter (m) di = pipe -Inner diameter (m) q _a = heat flow density on the outside (kW / m ² ) qi = heat flow density on the inside (kW / m ² ) Step 4:

Determination of the temperature difference Δ T "between the outer tube wall and the inner tube wall. The temperature difference Δ T _w is determined using the thermal conductivity equation:

with λ = thermal conductivity of the pipe material (kW / mK).

Step 5:

Determining the required mass flow density rh using the relationship:

An embodiment of the invention is explained in more detail with reference to a drawing. In it show:

1 shows a simplified representation of a continuous steam generator with vertically arranged evaporator tubes,

FIG. 2 shows a single evaporator tube in cross section,

FIG. 3 shows curves E, F, G and H for the mass flow density in the case of various geometries of an evaporator tube made of the material 13 Cr Mo 44, and

Figure 4 is a graphical representation of the dependence of the maximum permissible material temperature of 13 CrMo 44 on the permissible stress (N / mm ² ).

Corresponding parts are provided with the same reference symbols in all figures. In Figure 1, a continuous steam generator 2 is shown schematically with a rectangular cross section, the vertical gas train is formed from a surrounding wall 4, which merges into a funnel-shaped bottom 6 at the lower end. The bottom 6 comprises a discharge opening 8 for ashes, not shown.

In the lower area A of the gas flue, a number of burners 10, only one of which is visible, are attached for a fossil fuel in the surrounding wall or combustion chamber 4 formed from vertically arranged evaporator tubes 12. The vertically arranged evaporator tubes 12 are welded together in this area A via tube fins or tube webs 14 to form gas-tight combustion chamber or peripheral walls. The evaporator tubes 12 flowed through from bottom to top during operation of the continuous-flow steam generator 2 form an evaporator heating surface 16 in this area A.

In the combustion chamber 4, when the continuous steam generator 2 is operating, there is a flame body 17 which arises when a fossil fuel is burned, so that this region A of the continuous steam generator 2 is distinguished by a very high heat flow density. The flame body 17 has a temperature profile which, starting from approximately the center of the combustion chamber 4, decreases both in the vertical direction upwards and downwards and in the horizontal direction to the sides, ie to the corners of the combustion chamber 4. Above the lower area A of the throttle cable there is a second area B remote from the flame, above which a third upper area C of the throttle cable is provided. Convection heating surfaces 18, 20 and 22 are arranged in areas B and C of the gas flue. Above area C of the gas flue there is a flue gas outlet channel 24, via which the flue gas RG generated by the combustion of the fossil fuel leaves the vertical gas flue. FIG. 2 shows an evaporator tube 12 provided on the inside with ribs 26, which during operation of the continuous steam generator 2 on the outside inside the combustion chamber 4 is exposed to heating with the heat flow density q _a and through which the flow medium S flows on the inside. At the critical point, ie at a critical pressure p _K rit of 221 bar, the temperature of the flow medium or fluid in the tube 12 is designated T _k πt. For the calculation of the maximum thermal stress Omax, the maximum permissible material temperature T ^ x at the pipe apex 28 of the heated side of the pipe wall is used. The inner diameter and the outer diameter of the evaporator tube 12 are denoted by d _x and d _a , respectively. In the case of internally finned tubes, the equivalent inner diameter must be used, which takes into account the influence of the fin heights and valleys. The pipe wall thickness is denoted by d _r .

FIG. 3 shows four curves E, F, G and H in a coordinate system for different outside diameters d _a (mm) and tube wall thicknesses d _r (mm). For this purpose, the heat flow density q _a (kW / m ² ) is plotted on the outside of the pipe and the preferred or optimal mass flow density rh (kg / m ² s) is plotted on the ordinate. Curve E shows the course for a pipe outer diameter d _a of 30 mm with a pipe wall thickness d _r of 7 mm. Curve F shows the course for a pipe outer diameter d _a of 40 mm with a pipe wall thickness d _r of 7 mm. Curve G shows the course of the mass flow density m as a function of the heat flow density q _a for a pipe 12 with an outer diameter d _a of 30 mm and a pipe wall thickness d _r of 6 mm. Curve H shows the course of a tube 12 with an outer diameter d _a of 40 mm and a tube wall thickness d _r of 6 mm. The mass flow densities m are calculated for heat flow densities q _a of 250, 300, 350 and 400 kW / m ² at critical pressure of the flow medium S for the tube material 13 CrMo 44. An example for the determination of the optimal mass flow density th is shown below. The following conditions are required:

q _a = 250 kW / m ² ; Heat flow density on the outside of the pipe at a pressure of 210 bar.

1.4 as an increase factor to take into account local irregularities in the heat transfer to the pipes 12,

d _a = 40 mm tube outer diameter, d _r = 7 mm tube wall thickness, and tube material: 13 CrMo 44.

From d _a and d _{r it} follows: di = 26 mm inner tube diameter.

Step 1: Calculation of the heat flow density

The heat flow density based on the thermal calculation is multiplied by the increase factor. It follows:

q _a = 350 kW / m ²

Step 2: Determining the maximum permissible material temperature

According to equation (3) this temperature is calculated with T _κrιt = 374 ° C (temperature of the fluid at critical pressure p _kr it) _/ with ß = 16.3 * IO ^"6 (1 / K) (thermal expansion coefficient of 13 CrMo 44 ), E = 178 * IO ³ (N / mm ² ) (elastic modulus of 13 CrMo 44) and σ _to ι = 68.5 (N / mm ² ) (permissible stress of 13 CrMo 44 at the maximum permissible material temperature) to:

T _max = 515 ° C.

This determination of T _[nax , which has to be carried out iteratively, shows the dependence of the permissible stress C _to χ on the material temperature. This dependence between the allowable stress σ _2u ι of the maximum material _temperature T _max for the material 13 Cr Mo 44 is shown graphically.

Step 3: heat flow density on the inside of the pipe

With equations (4) and (5) it follows for A = 0.25 and

B = 0.725 for the heat flow density qi on the inside of the

Pipes 12: qi = 466 kW / m ² .

4th step: Determination of the temperature difference Δ T _w between

Pipe outer and inner pipe wall

According to equation (6) with the thermal conductivity of 13 Cr Mo 44 of λ = 38.5 * IO ^"3 kW / m K:

Δ T _w = 73 K.

5th step: Determination of the required mass flow density

According to equation (7) with C = 7.3 ^* IO ^"3 kWs / kgK:

rh = 939 kg / mJ s.

With the available values for the heat flow density q _a on the outside of the pipe and the maximum permissible material temperature T _m a _x , the optimal mass flow density rh can thus be determined. This value is represented by the dashed lines in FIG. 3 for the specified conditions. It can be seen that for the assumed heat flow density q _{a of} the pipe outer side of 350 kW / m ² for pipes 12 with outer diameters d _a between 30 and 40 mm and wall thicknesses d _r between 6 and 7 mm there are optimal mass flow densities rh between 740 and 1060 kg / m ² s result.

The mass flow rate determined in this way can be used for the fluidic design of the tubes 12 of the tube or peripheral wall 4. dense rh can still be converted to the conditions at 100% load. For this purpose, the operating pressure at the inlet of the tubes 12 is calculated at 100 Z. The mass flow densities mentioned above are then converted in proportion to the operating pressure at 100% load. If, for example, the operating pressure at 100% load is p _B = 270 bar, the mass flow density rh increases from 740 to 951 kg / m ² or from 1060 to 1363 kg / m ² s.

It can be expedient to take into account uncertainties in the determination of the heat flow density q _a by increasing the mass flow density rh from + 15% to + 20% compared to the calculated value.

Claims

claims 1. continuous steam generator with a combustion chamber (4) surrounding pipes (12) that are gas-tightly connected to one another by a surrounding wall, the pipes (12), which run vertically and have a surface structure (26) on their inside, from below in the flow medium (S) can be flowed through upwards, characterized in that the mass flow density rh in the tubes (12) at the load at which critical pressure pk _rit prevails in the tubes (12), the relationship: m = -; - (kg / m ² s) C (Tmax - Tkrit - ΔTw) ^V ' corresponds, where qi (kW / m ² ) the heat flow density on the inside of the tube (12), T _max (° C) the maximum permissible material temperature of the tube (12), T _kr it (° C) the temperature of the flow medium (p ) at critical pressure (pkrit), Δ T _w (K) the temperature difference between the outer and inner wall of the pipe (12), and C> 7.3 ^• IO ^"3 kWs / kgK is a constant. 2. Continuous steam generator according to claim 1, d a d u r c h g e k e n n z e i c h n e t that the heat flow density qi related to the inner wall of the relationship: with K = A (d _a ² .q _a ) + B, where: A = 0.45 and B = 0.625 for (d _a ² .q _a ) <0.5 kW, A = 0.25 and B = 0.725 for (d _a ² .q _a )> 0.5 and <1.1 kW, A = 0 and B = l for (d _a ² .q _a )> 1.1 kW , and where q _{a is} the heat flow density on the outside of the pipe (kW / m ² ) and d _{a is} the outside diameter of the pipe (m). 3. continuous steam generator-r according to claim 1 or 2, characterized in that the maximum permissible material _temperature T _{raax of} the relationship: Tm-x = Tkπt + _n (° C) corresponds to ß £ where σ _to ι is the permissible thermal stress (N / mm ² ), ß is the thermal expansion coefficient (1 / K) and E is the elastic modulus (N / mm ² ) of the pipe material. 4. Continuous steam generator according to one of claims 1 to 3, characterized in that the temperature difference ΔT _W between the outer tube wall and the inner tube wall of the relationship: m with K = A (d _a ² .q _a ) + B, where: A = 0.45 and B = 0.625 for (d _a ² .q _a ) <0.5 kW, A = 0.25 and B = 0.725 for (d _a ² .q _a )> 0.5 and <1.1 kW, A = 0 and B = l for (d _a ² .q _a )> 1.1 kW, and where q _{a is} the heat flow density on the outside of the pipe (kW / m ² ), d _{a is} the outside diameter of the pipe (m), di the inside diameter of the pipe (m) and λ is the thermal conductivity of the pipe material (kW / mK). 5. continuous steam generator according to one of claims 1 to 4, characterized in that for a tube (12) made of the material 13 CrMo 44 by value pairs of the heat flow density rh (kg / m ² ) certain points in a coordinate system on a curve E, for a pipe outer diameter d _a of 30 mm and a pipe wall density d _r of 7 mm is defined, by which the value pairs: q _a = 250 kW / m ² , m = 526 kg / m ² s, q _a = 300 KW / m ² , m = 750 kg / m ² s, q _a = 350 kW / mJs, rh = 1063 kg / mJs, and q _a = 400 kW / m ² , m = 1526 kg / m ² s certain points. 6. continuous steam generator according to one of claims 1 to 4, d a d u r c h g e k e n n z e i c h n e t that for a Pipe (12) made of the material 13 Cr Mo 44 by means of pairs of values for the heat flow density rh (kg / m ² ) in a coordinate system lie on a curve F which for a pipe outer diameter d _a of 40 mm and a pipe wall density d _r of 7 mm is defined by the value pairs: q _a = 250 kW / m ² , m = 471 kg / m ² s, q _a = 300 KW / m ² , m = 670 kg / m ² s, q _a = 350 kw / m ² s, m = 940 kg / mJs, and q _a = 400 kW / m ² , m = 1322 kg / m ² s certain points. 7. continuous steam generator according to one of claims 1 to 4, characterized in that for a tube (12) made of the material 13 Cr Mo 44 by value pairs of the heat flow density rh (kg / m ² ) certain points in a coordinate system on a curve G. , which is defined for a tube outer diameter d _a of 30 mm and a tube wall density d _r of 6 mm, by means of the value pairs: q _a = 250 kW / m ² , m = 420 kg / m ² s, q _a = 300 KW / m ² , ril = 576 kg / m ² s, q _a = 350 kW / m ² s, m = 775 kg / mJs, and q _a = 400 kW / m ² , rh = 1037 kg / m ² s certain points. 8. continuous steam generator according to one of claims 1 to 4, characterized in that for a tube (12) made of the material 13 Cr Mo 44 by pairs of values Heat flow density m (kg / m ² ) certain points in a coordinate system lie on a curve H, which is defined for a pipe outside diameter d _a of 40 mm and a pipe wall density d _r of 6 mm, by means of which the value pairs: q _a = 250 kW / m ² , rh = 399 kg / m ² s, q _a = 300 KW / m ² , th = 549 kg / m ² s, q _a = 350 kW / m ² s, m = 737 kg / m ² s, and q _a = 400 kW / m ² , th = 977 kg / m ² s certain points. 9. continuous steam generator according to one of claims 1 to 8, d a d u r c h g e k e n n z e i c h n e t that a mass flow density <1.2. m is permissible.