DE1919122A1 - Control system for steam turbines - Google Patents

Description

B.EOLZEB

89 AUGSBURG

-WEIiSEB-STBJLSSB IA id B1S7S

f. 436

Augsburg, April 13, 1969

Westinghouse Electric Corporation, 3 Gateway Center, Pittsburgh, Pennsylvania, United States of America

Control system for steam turbines

The invention relates to control systems for steam turbines. The field of application of the invention is the operational monitoring of Steam turbines.

In the operation of the different types of steam turbines, certain limits for the change in the steam flow rate and / or the input enthalpy for the purpose of changing the speed, the load or other variable parameters must normally be adhered to, so that the thermal and mechanical response n dar turbine is taken into account. The turbine is largely responsible for the, working steam pressure and the working

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temperature constructed; the rotor design can handle the centrifugal loads record at the maximum speed and the nominal torque, taking further design parameters into account accordingly including bending stiffness and critical speeds. The turbine housing and the rotor must however, also in the transition behavior and in the case of periodic changes in the steam state variables, if this causes a change the turbine speed and / or the turbine load takes place. The rotor must also withstand the centrifugal forces in the transition area can record, for example during a cold start and the like Operating states. Power or dynamic limits for turbine operation require careful consideration of the thermal expansion and contraction load due to changes in the state of the supply steam and also incoming Consideration of the centrifugal load in case of speed changes.

Below the steam state are the steam temperature, the Understand steam pressure and other parameters including steam flow.

The thermal load limits are particularly important because a change in the operating state of the turbine over time in most applications of the turbineÄ @ olmik is almost imra ©! ³ or at least Temperafc ^ Pgradienten b @ dingt _fl di © Aehr frequently to an expansion or contraction of the metal parts d © © r Turbia

result in exceeding the elastic limit. This forms in the turbine material and in particular in the Rotor material a temporal development of the processes and thus a "plastic deformation stress" with continued turbine operation the end. Within the turbine operating characteristics that meet the minimum requirements of the user the thermodynamic properties of the steam, the turbine size and the thermal and metallic properties of the turbine material together inevitable during turbine operation Pore change stresses of the turbine materials.

These deformation stresses in particular due to periodic Changes in temperature must be monitored when regulating turbine operation, since this causes fatigue damage and possibly fatigue fractures, which determine the service life of the turbine. Fatigue damage is particularly common in those areas that are the most severe and common Change in steam temperature. Such possible fatigue damage of the turbine material to the conscious Digits normally represent the strongest limitation for the dynamic characteristics in turbine operation »

In the case of large generator steam turbines, for example, changes in the steam temperature within the pulse chamber result

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Temperature gradients in the rotor, which in turn are periodic Generate deformation stresses in the rotor, in particular on or near the rotor surface in the area of the. Nozzle chamber, preferably at the foot of the rotor blades, on labyrinth seals or other grooves in the area of the nozzle chamber, which change in shape Stresses due to the concentration of thermal stresses due to the shape of the groove arrangements in question are conditional. In the case of large power plant turbines, ■ Fatigue cracks due to deformation stresses in the rotor observed in large numbers after an operating period of seven years or more. The cracks mostly went to a depth of about 6 mm, with occasional cracks reaching a depth of 125 mm.

Normally, in large steam turbines, cracks form in the rotor only when the operating time is less than periodic Load ratios is five or more years. In particular, rotors with 1,800 revolutions per minute form as a result their larger diameter and the resulting larger temperature gradients and higher thermal stresses more numerous and deeper cracks. Casing fatigue damage can occur in large generator steam turbines, but these have Turbines nowadays several steam chambers and each individual separate valve chambers, so that housing cracks normally

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cannot be detected by the operating personnel. Rotor fatigue damage represent an increasingly important factor in turbine operation, as the general development progresses to larger turbine sizes.

Temporary centrifugal loads are in those rotor areas where the highest centrifugal load is in continuous operation occurs, for example in the medium pressure part of large steam turbines, of particularly great importance. In this case the centrifugal load in continuous operation may lead to creep cracks, which must be distinguished from cracks due to the deformation stress. Usually exist appropriate dynamic restrictions to limit these transient centrifugal loads, especially when one at the same time the thermal load due to rotor temperature changes also taken into account. When starting large steam turbines from the cold state, for example, the Total stress in the bore of the high pressure part or the Medium pressure part become excessively large, if not an operational one Limitation takes place. In some cases, this load can be so great during the turbine start-up that a Brittle fracture occurs and the rotor is shattered.

In conventional turbine systems, transitional or dynamic operation is usually used with regard to the operational

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safety and the extension of the service life through an appropriate regulation or with the help of a program control restricted. In both cases, the relationships between various changes in steam flow rate and / or enthalpy changes must be established as well as the proportion and type of damage due to a corresponding thermal and / or mechanical operating stress be known beforehand. The Betriebsstandsan.de are monitored and by the supervisor or by a fixed program control related to each other so that the dynamic performance in general taking this into account certain relationship is limited.

Thermal and stress studies are used in conjunction with certain turbine designs to determine the Relationships between the rate of change of steam state and the turbine damage. Such a one Study "Prevention of Cyclic Thermal-Stress Cracking in Steam Turbine Rotors" by W.R. Berry and I. Johnsson in "Journal of Engineering for Power, Transactions ASME", July 1964, gives precise values insofar as the thermal stress study and the resulting rotor fatigue damage / are taken into account. Another work "Electrohydraulic Control for Improved Availability and Operation of Large Steam Turbines ", by M. Birnbaum and E.G. Noyes," ASME-IEEE National

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Power Conference in Albany ", New York, September 19-2J5, I965, explains a digital reference system for limiting the turbine speed and rates of change of load by a corresponding fixed program control.

Other known measures are in the work "Automatic Electronic Control of Steam Turbines According to a Fixed Programs "published in the" Brown-Boveri-Mitteilungen ", Vol. 51 * No. 3 * March 1964 · In addition to one fixed program control, the turbine control described there requires a measurement of the housing temperature and a feedback circuit with regard to the dynamic load and / or speed change, so that impermissible housing voltages are exceeded will.

A disadvantage of the known regulations for the dynamic Turbine operation and perhaps the biggest disadvantage in terms of rotor protection requirements and economy and the efficiency in turbine operation under control conditions lies in the lack of precise knowledge of the actual thermal load on the rotor and the actual deformation stress in the rotor, as well as the accumulation of fatigue damage, on what basis restrictions on the dynamic

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Turbine operation with a view to reducing turbine damage including the accumulation of fatigue damage could be specified. The known operating regulations are based on the fact that only approximately dynamic parameters are available are that the usual and unrestricted operation under different Concern operating conditions. The turbine service life, the degree of efficiency and the profitability of turbine operation are adversely affected by the limitations of the known systems for the purpose of dynamic monitoring and control, particularly with regard to large generator turbines and turbines in general. In the case of electrical power plants, it has the dynamic restriction of turbine operation to only a limited utilization of the power plant with adaptation to itself changing load conditions.

The invention aims to achieve the object that Regulation and monitoring of the operation of steam turbines in the sense of an extension of the service life of variable ones Perform stress levels with a continuous record the rotor tension takes place.

This object is achieved according to the invention by the following Combination of features solved: presence of a measuring device for the steam temperature within a heat transfer area

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on a selected rotor part, presence of a measuring device for the thermal stress of at least one rotor section as a function of the steam temperature in the stated range, presence of an adjusting device for the steam state variables in the stated range and presence a 2-drag device for the actuating device to limit the amount of change in the steam temperature within the specified range depending on the thermal state of the relevant rotor section.

The invention will be described below on the basis of a preferred embodiment with reference to the accompanying drawings explained. They represent:

Fig. 1 is a block diagram of a generator

Steam turbine with one steam generation system and with various Sensing and control devices according to the basic idea of the invention,

2 shows a block diagram of a control system

with a digital computer for the arrangement according to FIg ₀ 1,

Pig _e 5 on an enlarged scale

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J

Longitudinal section through the high pressure part of the Pig steam turbine. I with ver s hierarchy which KiJhI e inr i _s chtungen

the pig. Kk and
Pig. 4b

Excerpts from Pig in a further enlarged measuring stick. J5 for explanation the rotor parts on which fatigue cracks can form,

Fig. 5

a graph of the relationship between the steam pressure and the steam temperature as well as the steam temperature in the nozzle chamber of such a steam turbine,

Fig. 6

Temperature transition curves for different areas of a large steam turbine operating at synchronous speed is brought and operated under constant load,

Pig, 7

a diagram of a
the

a generator

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Pig. 8 shows a graph for effectiveness

periodic thermal deformation stresses corresponding to a periodic Operation according to FIG. 7 for a turbine rotor which is connected to J56OO Revolutions per minute is operated and its geometric shape is geared towards high thermal loads during operation,

9 is a diagram for the control process

by means of the computer according to FIG. 2,

10 shows the diagram according to FIG. 9 in larger dimensions

Details and

11 is a diagram for explaining the

Controlled limitation of the steam enthalpy and / or throughput changes and thus the steam temperature in the nozzle chamber as a function of the rotor voltage.

Flg. 1 shows a large single-casing steam turbine 10

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with reheating built in a conventional manner and is operated and regulated according to the basic idea of the invention. The steam turbine is part of an electrical power plant 12 / that runs on fossil fuels. Other types of steam turbines, e.g. extraction turbines, nuclear reactor turbines, Back pressure turbines can also use the basic idea of the invention are regulated.

The turbine 10 has a drive shaft 14 which drives a conventional large AC generator 16, the three-phase alternating current supplies and its performance with ^a: is measured power meter 18th In general, the generator 16 is connected in phase to a line network via one or more disconnectors (not shown). In this switching state, the turbo generator works at synchronous speed. steady state. When the load changes, the network frequency can be influenced in the transition area, which would result in corresponding changes in the speed of the turbo generator. In the synchronous state, the power output of the generator 16 into the network is normally determined by the steam throughput of the turbine, the additional steam being fed into the turbine 10 at a substantially constant pressure.

The turbine 10 is multi-stage with axial flow.

forms and comprises a high pressure part 20, a medium pressure part 22 and a low pressure part 24. Each turbine part contains a plurality of stages with guide vanes and Rotor blades. For other purposes, an im According to the invention regulated turbine also has a different structure with more or less parts that work on one shaft or on several shafts.

The constant pressure steam to drive the turbine 10 is generated in a steam generator 26, which acts as a drum boiler is designed, which is heated with fossil fuels, for example with coal dust or natural gas. In generalization the invention can also be applied to steam turbines connected to other steam generators, for example with nuclear reactors or other heating systems.

In the present embodiment, the turbine is designed as a double-ended turbine. The steam is introduced into the turbine through a plurality of control valves 25. Generally require turbines with double-ended steam chambers and others Steam chambers and turbines with single-ended steam chambers different number and / or arrangement of the inlet valves.

The preferred tolai operation for the turbine includes first

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the increase of the turbine speed from the speed of the rotor rotation device of about 2 revolutions per minute to about 80 % of the synchronous speed with the help of the valve control and secondly the transition to the control valve control and the increase of the turbine speed to the synchronous value with simultaneous closing of the isolating switch for the network and with Adjustment of the system to the load requirement. When stopping, corresponding switching operations are required in reverse order. Other transition techniques can be used, but the transition is unlikely to occur at a load point above 40 % of rated load when efficiency is taken into account.

After the steam from the first stage of the guide vanes has passed to the last stage of the rotor blading in the high pressure part, it reaches a reheater 28 ^ which is coupled to the steam generator 26. The reheater 28 normally consists of two connected in parallel Superheater parts that are in heat exchange with the steam generator, as indicated by the dashed line 29 and which are connected to opposite sides of the turbine housing "

With a correspondingly increased enthalpy throw, "

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Superheated steam from the superheater 28 through the medium-pressure part and the low-pressure part 24. From the latter, the relaxed one arrives Steam in a condenser 32, from which the condensate is fed back into the steam generator 26 in a manner not shown, will. To control the superheated steam, superheater valves 33 are provided which are one or more normally open Check valves and one or more interception valves, which in the event of an overload of the turbine in the sense of a branching of the superheated steam flow work.

With a steam generator that heats with fossil fuel the control is carried out in such a way that the live steam pressure is kept essentially constant. A live steam pressure detection device 38 of conventional design measures the live steam pressure and thus enables the live steam pressure to be kept constant. If desired, a program-controlled computer can be provided as a protection system that controls the control process of Throttle control instead of load control changes over when the live steam pressure reaches the limits of a safety range or protection area. An action pressure detection device 40 gives signals for use in the program computer to regulate the turbine load and thus ultimately the electrical power output.

The throttle and control valves 25 are each hydraulic

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actuated valve control elements 42 are provided. Hydraulically operated switching elements 44 also belong to the superheater valves 33 · One controlled and monitored by a computer Pressure medium source 46 supplies a control pressure medium for actuation the switching elements for valves 25 and JJ. A not shown, is also a computer-monitored lubricant system provided separately for the lubrication circuit of the turbine system.

The respective switching elements 42 and 44 are commercially available and are established by corresponding regulations 48 and 50 operated to stable positions. These controls each contain a conventional analog control loop, not shown Position error feedback, which controls a servo valve, also not shown, in a known manner. the Regulation for the superheater intercept valve positions is normally carried out only if it is necessary to change the superheated steam flow rate. The actuation of the shut-off valve requires no feedback of the position control and takes place as a result by hand or by means of a computer that has a conventional trigger function or some other suitable shutdown function.

Since the turbine power is within the assumed control behavior with essentially constant live steam pressure pro-

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is proportional to the steam throughput, the valve position is in the sense a regulation of the steam throughput regulated as an intermediate variable, while the turbine speed and / or load the regulated Represent sizes. The switching element actuation results in the setting of the steam valves. The. respective detection devices for the valve positions PDIV and PDRV provide the required position feedback signals for deriving position Pehler signals, which create a respective position control loop 48 and 50, respectively. The position verification devices are designed in conventional form, For example, they use a linearly variable differential transformer that emits a negative feedback signal, which is algebraically summed up with the respective desired position signals Sp to form the position error signal.

The combination of position control and hydraulic Setting, the fault position detection device and others, not shown ex »Additional equipment forms a hydraulic-electric® Analog control loop for the valve position of each steam = throttle = bzi-j «Sfceuer inlet valve. the

Setpoints SP v-mraon determined by a computer and periodically. Setpoints SP entered into the respective control loops are ^{determined for the felt 0,} the catch valve controls,

A Dralisahlmeuser 52 determines the WelLendrehzalil for the

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Speed control for determining the centrifugal stresses and the stress ratios, for regulating the frequency influence and preferably for calculating the surface heat transfer in connection with the regulation of the thermal rotor stress. The tachometer 52 is designed, for example, as a magnetic pickup, not shown, which is coupled to a cam disk on the turbine shaft 14. The sensing device also includes a temperature gauge 54 for the steam in the steam chamber and case temperature gauges 56 which are used to calculate the load and thermal stress, as will be explained below. Analog and / or pulse signals from the tachometer 52, the power meter l8, the pressure detection devices 38 and 40 * the temperature meter 54 and 56 * the valve position sensors PDIV and PDRV as well as other sensors, not shown in detail and also not shown status contacts are all in one Digital controller 60 (Fig * 2) entered * the regulation of the turbine continuous operating state and the transition state on a real-time basis and also a monitoring »sequence control _s . Warning signals, alarm signals ₉ display = and recording functions provides «,

The program-controlled digital controller »60 allows operation the turbine 10 with improved dynamic characteristics and can be conventional components in Forsi a central Arithmetic unit 62 and associated input © = output exchange ~

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directions include those of the company, for example Westinghouse Electric Corporation, Pittsburgh, under the Trade name "Prodac 50" (P 50) can be supplied. In other cases, for example, when the turbine 10 as well as others Plant parts such as the steam generator 26 are under the control of the computer, a larger computer can be used, for example the "Prodac 250" computer or a separate P 50 computer can be used for the controlled parts of the system will. In the latter case, the interaction between the regulations is achieved by connecting the individual computers to one another reached via information channels and / or other elements.

The P 250 computer normally uses a magnetic core memory for 16,000 words with Vo digits each and the sign with an access time of 900 nanoseconds, and an external magnetic memory for 12,000 or more words with 16 digits each and the sign with an access time of 1.1 Microseconds and finally a disk storage for 375,000 words or more with random access. The P 50 calculator uses an internal magnetic core memory for 12,000 words (1-4 digits) with an access time of 4.5 microseconds.

The exchange device for the arithmetic unit 62 comprises

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an input part 64 with contact closure, the contacts or queries other signals representing variables of the various parts of the system. Such contacts 56 are normally designed as mercury switches, which are actuated by excitation circuits and for sensing the specified state variables for the various Parts of the plant are used. The contact information about the state is in a logical linking technique or according to other program plans for regulation, for a protection and Warning function, for automatic monitoring, for recording and for querying the recording and for puncturing a computer-operated one Hand monitoring control 68 used.

The input part 64 with closing contacts receives also a digital stress reference signal 70. The stress reference signal 70 can be entered manually or automatically be supplied by an operating computer. For the load control, the load reference signal 70 is the setpoint value of the load level fixed in megawatts and the digital controller 60 drives the turbine 10 in the sense of delivering the desired load.

An input exchange stage is through a conventional analog input device 72 realized the analog signals from the System 12 queries with a predetermined sampling frequency, for example in 15 points per second for each analog channel, and that

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converts these signals into digital values for input into the computer. The analog signals are generated in the power meter 18, the operating pressure detection device 40, the valve position proximity devices PDIV and PDRV, the temperature sensors 5 ^ f- and 56 and in various analog sensors IK , for example in the live steam pressure detection device 38, which is not shown in detail in FIG is, in various steam flow meters, in steam temperature gauges, in various other devices that work as temperature sensors, in a pressure gauge for a hydrogen cooling generator, in temperature gauges and the like for input into the calculator. The corresponding computer circuits for the analog and pulse input signals are used to control the program execution, the protection and alarm equipment, the programmed and queried recording and the like.

Information input and output devices allow Input and output of coded and recoded information for the computer. These devices include a conventional one Tape reader and printer 78 which serves various purposes including entering the program into the core memory of the Computer is used. A commercially available teletype 80 is

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also present and is used, for example, to write out the recorded values in a recording medium 82. Alphanumeric and / or other types of display devices 81, 8j5. » 85 serve for the respective transmission of the information about the respective rotor tension, the total rotor fatigue as well as others Information values.

A known interrupter unit 84 is with corresponding Switching stages for controlling the transfer of input and output information equipped with a low clock frequency between the arithmetic unit and the input-output devices. It lies to the arithmetic unit 62 an interrupt signal when an input information is ready to enter or when an output transfer is complete. Usually the arithmetic unit speaks 62 responds to interruptions in accordance with a conventional follow-up program. In some cases special interruption inheritors misses accepted and processed without the execution providing for a priority limit here.

The output is coupled for the arithmetic unit by means of a conventional output contact device, which is shown in Connection to an analog output section 88 and a valve position control section 90 works. A hand control 92 is coupled with the valve position control part and works with it

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together, so that a turbine control of Hand is possible.

Certain output channels of the digital computer are used directly to carry out certain program steps and those that are contact-controlled Control operations for the high pressure valve fluid and lubrication system 87, for alarm levels 94 such as buzzers and display devices, and for certain auxiliary equipment 96, such as. the hydrogen cooling system for the generator. Output information of the digital computer act in a similar manner directly to the tape printer and the telex 8θ and the display devices 8l, 8j and 85.

Other output signals from the digital computer are initially by means of the analog output part 80 and the valve position control part 90 converted into analog signals. The same are then on auxiliary devices 96 and on the pressure medium and Lubrication system 87 and the valve control stages 48 and 50 for execution certain control processes. The respective signals for the position control circuits 48 and 50 are setpoint signals SP for the Valve pitches that have already been explained above.

The temperature measurement is used to determine the plastic

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ORIGINAL INS ⁰ ECTED

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Deformation stress of the turbine material, which is based on the Fig. 5 shows in detail the most important parts of the turbine baus is explained in the high pressure part 20 with the preferred arrangement of the temperature sensor. The high pressure part 20 comprises a cylindrical housing 100 in which a rotor 102 is rotatably mounted. The case tension at certain parts of the case is used with thermocouples 104 and 106 to measure the temperature ah the inner and outer walls that determine the housing temperature meter 56 belong.

A corresponding steam temperature sensor (not included in the is shown individually, but belongs to the analog sensors 79) 'can also be present in the medium-pressure part 22, as well as in the steam inlet line, but is preferably located in the nozzle chamber IP, not shown. The steam temperature in the chamber serves as the basis for calculating the thermal Tension of the rotor shaft in the medium pressure part 22.

The steam flows into the turbine 10 via a plurality of inlets 108 distributed around the circumference, each with associated Nozzle groups 105. He is then within a speed step, which has two rows of rotor reaction blades 107 and 109 and a row of stator blades 111, in.eine Nozzle chamber 110 passed. As indicated by the flow arrows, the direction of flow of the steam is then reversed and

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it passes through reaction blades 112 into the subsequent ones Levels of the high pressure part. A thermocouple 114 sits on the housing 100 and is used to measure the temperature of the steam in the nozzle chamber. A blind pin 116 carries spring-loaded sealing rings 115 and sealing strips 113A (Fig. 4A), which provides a seal against impermissible axial steam outflows through the gap between the rotor shaft and the surrounding stator part.

In FIGS. 4A and 4B, the further enlarged sections 115 and 117 from FIG. 3 show circumferential labyrinth sealing grooves 118 and 120 and circumferential rotor blade slots 122. Furthermore, fatigue cracks 124, 126 and 128 are shown, which arise due to long-term, periodic deformation stresses Form at the bottom of the grooves 118, 120 and the slots 122 and go back to temperature changes in the nozzle chamber. In other cases, the vane slots can also run in the axial direction instead of in the circumferential direction. Wise in any case the rotor grooves and slots as well as similar shaped elements of the rotor within or in the vicinity of the rotor surface Effect in the sense of a concentration of the stresses, i.e. at these points there is a greater thermal stress an increased formation of fatigue cracks as a result of

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Pormäaderungsspannungen encountered.

In general, fatigue cracks develop in the groove direction because the stress concentration is mainly in the transverse direction the respective groove is effective.

Before the operation of the program-controlled computer is considered, a more detailed examination of the thermal Rotor stress and an analysis of the deformation stress as well as the relationships with the turbine characteristic are useful. the Thermal stress in the rotor surface and especially in the rotor slots and slots is significant because on such points of the rotor temperature gradient as a result of changes in the steam temperature a maximum of the thermal rotor stresses and cause deformation stresses. The thermal see Stresses within bushings are primarily during a cold start or a similar operating state large, in which cases the combined effect of centrifugal and thermal transient loads on the bore can become excessively large if not subject to restrictions the operating behavior of the turbines.

The thermal surface stresses and deformation stresses of the rotor are largest in the vicinity of the nozzle chamber part, because there, due to the strongest changes in the steam

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temperature these tensions are greatest. To determine the rotor surface tension, one must determine the surface temperature in the vicinity of the nozzle chamber. The surface temperature of the rotor is derived from the steam temperature Τ ,, in the nozzle chamber and the heat transfer coefficient in the rotor surface as described below. Since the heat transfer between the steam flowing through and the rotating rotor surface is very large at high turbine speeds, the rotor surface temperature T _a is essentially the same as the steam temperature T _T in the nozzle chamber, with the exception of the start-up and shutdown operation with comparatively low steam throughput and steam pressure as well a low speed. The steam temperature T _T in the nozzle chamber can for its part fluctuate to a large extent with changes in the steam throughput, even if the input enthalpy is kept constant.

Inside the medium pressure part 22 of the turbine 10 is due to the reduced steam density and the reduced steam pressure

the heat transfer into the rotor surface is lower than in the high-pressure part 20. Pur calculations of thermal bore stress the determination of the rotor surface temperature in the inlet section of the medium-pressure part 22 is based on the measured steam temperature IP in the inlet chamber and the different and

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ORIGINAL INSPECTED QO9839 / 1180

lower heat transfer coefficient in the rotor surface. The medium-pressure heat transfer coefficient is, similar to the high-pressure part, a predetermined point of the turbine speed. In addition, the mean pressure heat transfer coefficient is also preferably determined as a puncture of the mean pressure steam throughput and possibly the mean pressure vapor density or the mean pressure vapor pressure, namely with (Κ ™) _Ιρ = ^f ( ^w _S J ^SP * ^ρ τρ ^ where WS is the actual -Turbine speed, SP are the medium-pressure steam throughput and the medium-pressure steam pressure. Corresponding flow and pressure meters, which are not shown in detail, are used within the switching block 79 so that when calculating K _T "the values for the medium-pressure Receives throughput and pressure.

For the normal case, the relationship between the steam temperature T _T in the pulse chamber and the state variables of the Zudampfes in Fig's. 5 for a large steam turbine, which is designed for a steam pressure of I69 kg / cm and a steam temperature of 538 ⁰ C. As soon as the input enthalpy is determined from the input pressure and the input temperature, the steam temperature becomes T ₃ . of the nozzle chamber read from the intersection of the values for enthalpy and steam throughput »

The engine surface area can vary according to the above

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Work to be determined by Berry and Johnsson. The transition sections of the heating-up and cooling-down curves as well as the linear transitions of the heating-up and cooling-down curves when the temperature of the flowing steam changes are important, as these normally have to be taken into account alone and these basic curves can be superimposed to derive the individual transition behavior. In addition, the thermal surface tension is proportional to the difference between the rotor surface temperature Tg and the total rotor temperature T. As will be explained in detail below, the control with the computer requires the determination of the rotor surface tension E ", which is proportional to the difference between T _g and T.

A step-shaped heating or cooling transition curve normally results in the greatest rotor surface tension which is proportional to the size of the change in flow temperature, close to the starting point. The surface tension falls then from the peak values at the initial point in time until the internal temperatures of the rotor equate to the rotor surface temperature approximate with a tent constant that of the dimensions and shape of the overall cylindrical rotor and the thermal properties of the same depends. In the case of a linear heating or cooling transition, the thermal surface

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The voltage at the beginning is vanishingly small and increases as the difference between T "and T" increases, since the rate of change of the internal temperature lags behind the rate of change of the surface temperature. As the linear transition process continues, a state is eventually reached where the ambient temperature and the internal temperature change at the same rate. This represents a quasi-stationary state with a constant difference between T ₀ and T and a constant one

Rotor surface tension, which latter depends on the rate of change in the ambient temperature.

In a frequently occurring case of a change in the load on turbo-generators, the linear transition ends at the final value of the steady steam temperature as soon as the desired change in load has taken place. Usually the transition ends before the quasi-steady state has developed. As a result, the difference between T _g and T is usually greatest at the end point of a linear transition and sets the maximum thermal surface tension for the transition. The maximum voltage depends on both the size and the rate at which the steam temperature changes. At very high rates of change, one approaches the limit of a step-shaped transition.

ORIGINAL INSPECTED

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3 »

For a better understanding of the periodic behavior of the thermal stress and the thermal deformation stress in the rotor surface, FIG. 6 shows a time diagram for the. Throttle steam temperature, plotted for the steam temperature Ty in the nozzle chamber, for the rotor surface temperature T _g , for the rotor bore temperature Tg and the total rotor temperature T, namely for the start-up and load case of a generator steam turbine with a largest diameter of 575 mm and 5,600 Revolutions per minute. The rotor initially has a temperature of 205 C, with live steam at a temperature of ² I-OO ⁰ C being available at the start-up time. After throttling and expansion, the steam temperature in the nozzle chamber reaches a "value of 275 ° C. During the start-up period, the rotor undergoes a gradual heating transition with subsequent linear heating." During the start-up period, the rotor surface temperature follows curve I30 only slowly because of the low speed and the low pressure in the nozzle chamber, the heat transfer coefficient is small.

_{The temperature T 3} falls immediately before the synchronous speed is reached. in the Impulskaraner off because a switch is on

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a partial flow takes place, because the control is now from the throttle valves to the control valves. If the synchronous control during the initial load by a throttle bypass valve occurs, falls when switching to partial flow the steam temperature later decreases with a more pronounced influence on the rotor surface temperature.

After synchronization, the live steam temperature can rise steeply with increasing steam throughput and increasing load, as soon as a 5 percent load value of the turbine is reached. The rotor surface temperature T _g at this point in time due to the increased surface heat transfer coefficient due to the greater steam throughput and higher steam pressure and the higher rotor speed of the nozzle chamber temperature Tj is almost the same. In this case, the 5 % load remains until the rise in the live steam temperature has reached a steady-state value. The load is then applied with an approximately uniform rate of change as the throttle steam is brought to the setpoint temperature. The increase in load consequently results in a linear heating transition of the turbine rotor. In this case, the response of the rotor temperature reaches the quasi-stationary state during the load period.

The main effect of the periodic deformation

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The tension of the rotor in the event of changes in the turbine speed and / or load or, more generally, in the event of changes in the steam flow rate and / or the enthalpy, lies in the accumulation of fatigue damage to the rotor. In Fig. 7, a typical daily load _{fluctuation of the steam temperature T 1} in the nozzle chamber, the rotor surface temperature T ₀ and the total rotor temperature T due to a load change corresponding to full load operation during the day and part load operation at night is recorded for a large steam turbine.

Due to the periodic rotor surface temperature changes due to the heating and cooling transitions in the turbine rotor the material of the rotor surface passes through a stress-strain hyseresis curve "each depending on its properties, the rotor material being, for example, a Commercially available Cr-Mo-V steel alloy can be used. Thereby occur during the heating section 1 ^ 6 thermal Compression stresses within the deformation area. Thermal residual stresses remain within section 1 ^ 8. Thermal tensile stresses due to the change in shape arise during the cooling section 140 and ultimately remain thermal residual stresses again during section 142. The width of the stress-strain hysteresis loop gives

- 33 -

ORIGINAL INSPECTSD

009839/1180

•! 918122

the deformation stress within such a period. With small amounts of change in the steam throughput and / or the steam enthalpy, the elastic limit of the material becomes not exceeded, so that when the steam temperature changes, the deformation voltage in the rotor surface then presumably almost disappears. 'In most cases, however, the periodic change in the turbine load results in an influence the deformation stress.

Other periodic curves of the rotor surface temperature each cause similar hysteresis loops and corresponding ones periodic deformation stresses in the red surface. With turbo generators another typical one is periodic Temperature curve an essentially sinusoidal temperature curve, which is caused by the frequency control «,

In FIG. 8, a fatigue diagram for a flat change profile in the rotor temperature is similar to that in FIG. 7 shown. This diagram gives the number of deformation ' Periods which, under changing load conditions, lead to the generation of cracks in a rotor with one speed of 5,600 revolutions per minute and a diameter of 575mm is required if the rotor shape gives high thermal efficiency. Larger rotor diameters result a greater thermal inertia and thus require a

- 34 - ORSQWAL! NSPECTID

009839/1180

greater stretching of the time scale than in Fig. 8.

The flat periodic change of the surface temperature, which forms the basis of the fatigue diagram, is idealized in the upper part of the Pig. 8 specified. Normally, the linear heating takes place within a time period At from a first steady state of the surface temperature, up to a second steady state of the surface temperature, a state of equilibrium occurs in which the surface temperature is higher _{by Δ Τ σ at the beginning.} The linear cooling extends over a period of time Δ t until an equilibrium state is reached at the first steady-state temperature value. The heating and cooling amounts and speeds are for reasons of. "Simplicity is the same, even if these parameters can be different within the scope of the concept of the invention. When using the diagram according to Pig. 8, the change ΔΤ _{Ο of} the surface temperature and the heating and cooling time At are determined, the number of periods N required for generation rotor cracks are then read from the curves 144. Correspondingly, N can be determined from locus curves 146 if the rate of change of the surface temperature and the respective temperature interval Ut are known , - "

Derivation of the diagram in FIG. 8 from the strength

ORIGINAL INSPECTED

'\ QI Qi 9?

3ί

and other properties of the rotor material as well as from periodic voltage determinations is explained in the mentioned work by Berry and Johnsson. Corresponding diagrams can be set up for the periodic fatigue behavior due to other causes of fatigue, for example for other periodic temperature curves, such as sinusoidal curves.

The fatigue damage is determined from numerous diagrams of the fatigue behavior by first specifying the type of period, So flat course, sinusoidal course etc. are determined and then the period value N from a corresponding fatigue diagram based on the respective values of the rotor temperature change « and the rates of change as well as the transition duration for similar values are determined. The extent of damage per period l / N is then determined. All of the fatigue damage is equal to the sum of the values for l / N during the course of the turbine service life. If the total value is "l" reached, fatigue cracks are theoretically already in the rotor surface expected. The fatigue cracks can already occur with a comparatively few large periods, with a large one Number of small periods or any combination of different periods.

A known technique requires the determination of the

- "56 -

ORIGINAL IMSPECTID

009S39 / 1180

desired turbine life against fatigue Deformation stresses and the establishment of a fixed one Monitoring and control program restriction for the periodic Turbine operation so that the resulting total rotor fatigue damage according to the calculation from the The fatigue diagram and the planned period of operation corresponds to the desired turbine service life. As already has been said, this procedure could so far only be used approximately, with disadvantages with regard to Have shown accuracy, efficiency and economy of turbine operation.

In the program-controlled computer now to be considered, a control program is used to operate the arithmetic unit used. The same includes control and auxiliary program parts as well as certain conventional economy programs for the Internal regulation of the function of the arithmetic unit itself. The following individual levels are available:

1) Priority Execution Program

The same controls the use of the arithmetic unit. In general, this is done on the basis of a priority classification of all occurring regulatory and economic

00 · Μ · / 7ΐ · 0

1913 122 3 «

, economic programs and various types of Interruption conditions. The most significant program or the interrupt subroutine is determined and triggered when a change is to be made in the context of the executed program command sequence. Some breaker subroutines run outside of the priority relationship, as has already been pointed out, in particular, if security and / or extensive protection make it necessary.

2) Analog query "■

There is a periodic execution for the input of certain analog values, which are converted by the analog input part 72 and in the analog input buffer memory of memory are.

b 5) Contact status query ■

W ■

There is a periodic implementation for input certain input signals caused by a contact state are shown.

4) Input program for the programmer

On request, the operator is able to

"^ ⁸ ~ ORiGfNAL INSPEGTED

009Ö3Ö / 1180

Λ r

Q 1 O ₁ A 9?

Enter information into the computer memory. 5) Error subroutine

This is executed when the computer has an error condition occupies.

The control programs and utilities include the following Options:

1). Information recording

The. Execution occurs either periodically or on Requirement and includes printing out specified Data and parameters.

2) alarm

Periodic interruption of operation to operate the Alarm device 9 ^ and to operate other elements and for monitoring and / or switching off a valve position and other control programs.

3) display

Periodic and any alphanumeric execution

i ^: ORIGINAL! NSPFCTED

or graphical visual displays of certain parameter values and / or function profiles.

4} High pressure valve passage program Periodic execution with regulation monitoring.

5) Lubrication circuit program

Periodic execution for control monitoring.

6) Auxiliary facilities and program systems . .:. . .

Periodic execution for control monitoring ....

7) Steam inlet valve position control program

Periodic execution for monitoring purposes. ι

8) Turbine rotor loading and thermal deformation stress subroutine

serves as part of the steam inlet valve position control program to limit the rate of change

- ^4o - ORIGINAL INSPECTED

00983 ^ / 1180

of the inlet steam flow to ensure a safe Obtains turbine operation for a longer period of time and generally better performance.

9) Heating valve position control programs

Performed after and during overspeed alarm requirements.

The present invention primarily requires the turbine rotor loading and deformation stress subroutine to work properly as well as other sub-program parts, the description of which in the following refers to the valve position holder program and the rotor voltage subroutine included therein is constrained. Flow charts with corresponding Algorithms are given in FIGS. 9 and 10 which illustrate the represent the basic logical content of the valve position subroutine and the deformation stress subroutine 156. These programs are entered into the arithmetic unit 80 in the machine language of flow charts broken down into even more individual steps away, which in turn are derived from the flowchart shown.

Before starting, the turbine 10 is powered by a rotary motor

- 41 -

009839/1180 originally inspected

91912

driven at a speed of about two revolutions per minute to reduce the AnI to torque and the shaft keep in straight alignment. To start up the turbine 10, for example, the manual control 68 a start-up signal is input into the arithmetic unit 62. The start-up is carried out by a program sequence if the specified, permissible logical links are fulfilled, including the normal function of the steam generator, the setting of the Live steam pressure to the desired value, the opening of the Power take-off plate, the setting of the turbine steam valves to the start-up position and ensuring the normal Function of high pressure flow system, etc.

After the clearance for start-up, the steam valve position control program starts 145 performed periodically, "for example with a repetition frequency of 1 Hz, thus initially setting the steam valve in the sense of a run-up the turbine 10 takes place at synchronous speed, after which a regulation of the turbine load takes place.

The function block 147 represents a feedback correction d _{Q of} the turbine speed as the product of the gain g and the value Δ S, the latter representing the difference between a reference _{speed w R} and the actual speed w ". In

- 42 - CRlQfNALlMSPECTED

009839/1180

In this case, the reference _{speed w R} approximately gives the speed change within specified dynamic limits, which is formed from a start-up or deceleration curve of the turbine speed as a function of time stored in the computer. The gain g corresponds to the desired speed adjustment of the system. The speed adjustment g can be, for example, 3 % , ie an overspeed of 3 % at full turbine load leads to a complete closure of the inlet steam valves 25. The numerical form of the speed correction d_ is thus expressed in percent. This is advantageous in the case of a load control.

With the setting of the arithmetic unit 40 for the start-up or In the braking mode, function block 148 provides a transfer of the program to function block 150, which _{defines a valve setting variable D SM} for a maximum speed change that is dynamically given for the control system, the shape change voltage depending on the change in the inlet steam flow when the turbine speed control changes. By limiting the rate of change in turbine _speed, the limit D gM actually acts as a

"- 43 -

ORIGINAL INSPECTED

000839/1180

Negative coupling with reference to the speed edge w ", the a forward feed: makes necessary, but only with approximate dynamic restriction ...

If the turbine 10 'is already running at synchronous speed, the function block 148 transfers the program execution to the function block 152, where the speed-related load requirement is determined from the load reference variable Dy within the block 70. Function block 154- then supplies a valve setting value D _M for a maximum change in load, which is dynamically assigned to the control in the sense of limiting the rate of change with which the steam inlet can be changed in the case of turbine load control. When applying a limitation to the rate of change of the turbine _{load, the limitation variable D M} actually acts as a dynamic negative feedback limitation within the load control loop, which preferably works with feedforward control. Function blocks 150 and 154 each contain some common building stages and form the stages for the rotor load and stress constraint subroutine I56, which is shown in more detail in FIG.

When starting and braking control operation, the overall setting Dg of the steam valve setting is within a

- 44 - ORIGINAL INSPECTED

00983971180

determined closed control loop and made _{equal to a given puncture of d a} or restricted _{by D 0} .. under given load and thermal conditions. D -,., Can be a numeric variable; in this case, either an edge rise of w _R is preferably permissible or not permissible if the restriction of the speed change is to be met. The steam valve adjustment then takes place for a fixed value of the reference speed depending on a speed error signal until the restrictive effect is lifted. The determination of D _gM means an _{equalization of D g} according to the given function of d "with W _n kept constant.

ο η

Similarly, when driving with load control, the desired position D ^, of the steam valve within the function block 157 is determined from the variable D _R or with the aid of the variable Dj. limited under consideration of a given rotor load or a given voltage state, in particular if D _{R is} greater than D _{M lflt # D} is preferably determined using static ratios in the feedforward control loop. A calibration with D _pc then takes place on the basis of the pulse chamber error signal within the function block 158. Normally, the static determination within the function block 157 sets the overall steam valve position

- 45 - ORlGtMAL, IMSFECTED

009839/1180

fixed for the setpoint D _T or D., if there is a restriction

Jj JXI

of D _{T is} required. The puncture block. 158 leads due to

its balancing effect is a correction of every minor discrepancy and other errors.

The desired valve position value D ₃ or D " _c is distributed to the inlet steam valves 25 within the puncture block 16O according to a predetermined distribution rule.

The respective digital setpoints for the valve positions are then determined in puncture block 162 before the Program sequence comes to an end.

If the sub-program 156 for the dynamic restriction of the determination of the valve setpoints due to maximum load changes or speed changes is executed in function blocks I50 or 15 ^, the thermal surface tension E _g within function block 164 (FIG. 10} is first determined according to the equation given above The thermal stress Eg should preferably be a thermal state variable of the rotor for which a monitoring or regulating action must be carried out for the turbine control, because this is the basic variable on which the fatigue damage depends. If desired, other thermal state variables,

■ ■ ORIGINAL INSPECTED -

- 46 -

9333/1160

how the thermal tension is determined and in the arithmetic unit are processed.

The rotor surface temperature T _g can, under certain working conditions, be equated with the steam temperature Ty detected in the nozzle chamber, namely when the heat transfer coefficient Kjg in the rotor surface has a very high value. This is the case for most of the loading periods of turbo-generators according to FIG. 7 and for most of the other operating periods. '

For operating periods in which the heat transfer coefficient of the rotor is too small _{to justify equating Τ σ} and T _x , and during which high accuracy is required in the start-up and shutdown state of the turbine 10, or during normal operating periods under other working conditions which require a change in speed, the rotor surface temperature Τ _σ can be determined automatically as a function of variables such as the nozzle chamber steam temperature and the heat transfer coefficient K ™ for the steam flowing through with respect to the rotor. The heat transfer rate in the high-pressure part is preferably determined as a function of the rotor speed, that is to say (K _IS ) _H p = f (w _s ). The rotor surface temperature

- 47 _; ORiGiNAL INSPECTED

009839/1180

is calculated from the temperature T _{1 of} the steam flow and from the above-mentioned value K- _s in the following way: In the load "control operating state, the variable K- ™ from the determination ^: of f (Wc.) is normally essentially constant and large enough to set T ₀ = T ₁ - * Normally the heat transfer coefficient of the surface depends primarily on the turbine speed and the steam throughput, in the second place, on the steam pressure and / or the steam density as well as on other thermodynamic parameters The punctures (K ₁₃ ) _Hp and (K _IS ) _{Ip can be derived from a general relationship}

The puncture block 164 is used to determine the mean total temperature T of the rotor. This calculation is based on the usual analysis of transitions of the heat gradient in cylinder sections, for example according to the work by GM Dusinbene "Numerical Analysis of Heat Plow", 1949, McGraw-Hill. In general, for the calculation, the rotor is subdivided into a certain number of adjoining ring sections with the same radial expansion, which are numbered from the outside to the inside. The respective ring sections have heat capacities ^C l · ■ · " ^C N ^and J ^ewei l ^s exchange heat transfer coefficients Κ, ο * Κ, · * ··· K / _, \ ( _η \ · Equations for determining the ring section heat capacities for the heat flow between the steam stream

- 48 -

ORIGINAL INSPECTED

00 9839/1180

1913122

and the surface ring in question through the surface layer in question with the heat transfer coefficient K ™, then between the first and second and further inward pairs of rings, each with the corresponding heat transfer figures depending on the steam temperature and the ring temperature at time t _Q and on the ring temperatures after a time interval / \ t has elapsed or at the point in time (t _Q + t). In each such equation, the ring temperature at the point in time (t _Q + t) is expressed as a function of other variables. The calculated value T i is the value used in function block 164. The actual total rotor temperature T is calculated from the ring temperatures according to the following relationship:

C ₁ T ₁ + ... + -C ₁ T
rji 1 1η η

with C ₁ - surface ring heat capacity T ₁ = surface ring temperature = T = f (K _{T <3} , T _T )

1 5 JLo JL

C _n = heat capacity of the η-th ring section T _n = temperature of the η-th ring section.

After each determination of T, T _{a is} subtracted from T. The difference is with the coefficient of thermal expansion «6 as well as

- 49 - ORIGINAL INSPECTK

009839/1180

multiplied by a constriction factor or / W, which determines the influence

of the groove base takes into account where the thermal stresses

and stretches are concentrated. The quantity thus obtained is then divided by the value (1 - V ), in which V is Poisson's number, whereby the rotor surface tension E _g is obtained. If the surface tensile _{stress S 3 is to} be calculated instead of the thermal stress E ₃ , a modified equation is used within the puncture block 164 that uses the modulus of elasticity E as a further factor for the difference (T - T-). In addition, a voltage constriction factor is used instead of the factor / U.

The surface tension E _{g obtained} is stored in a function block: 166. The following voltage _{values E 3} from each subsequent program execution are read off in order to determine the periodic effect of the voltage sequence. Preferably, only particularly characteristic voltage periods are recorded during the reading process. The puncture block I68 calculates the fatigue damage caused by the stresses of change in the size of the rotor in connection with the respectively recorded period. The damage is calculated by determining the above-mentioned quantity N on the basis of the fatigue diagram, as soon as the type of stress period and the parameters thereof have been determined.

ORiGiWAL INSPECTiD - 50 -

009839/1180

919122

The function block 17 * 0 works as a fatigue damage measurement system, since it successively adds the damage values obtained and thereby supplies a variable for the respective instantaneous value of the fatigue damage. The display parts 81 and 83 each show the respective rotor surface tension E n and the surface fatigue damage after corresponding program commands or after a query by the operator. Normally, the damage summation calculation due to the actual fatigue damage depends on the parameters for determining the stress periods, i.e. on the type of classification used to determine the periods leading to damage as well as the load period not taken into account, for example with very rough sampling the the recorded number of fluctuations in the turbo-generators must be as low as 1 per day. In such and similar applications, the puncture blocks 168 and 168 would be ineffective.

If desired, a closed, not shown, can be used for the summation of the fatigue damage in the puncture block I70 Voltage exciter loop for basic control in long-term monitoring come into use. For example, the voltage -

- 51 -

00 9 8 3 9/1180 originally inspected

913122

ει

limit Ear for ^the function block: 184 can be modified over time depending on the calculated total damage.

The temperature meter 54 for the steam in the nozzle chamber and assembly; For the determination of the rotor surface tension E _g as well as, if necessary, for the recording and accumulation of the fatigue damage can each be arranged separately in the form of control panel instruments, which are useful for the monitoring of the turbine operation, if a dynamic tension control in a closed loop is not desired. In such a case, the output of the steam temperature meter for the nozzle chamber and the output of the tachometer are coupled to a corresponding computing stage which can process the program of function block 164 and, if necessary, also the programs of function blocks 166, 168 and 170. This switchboard computer can be a special analog computer, a digital-analog computer or a digital computer. Such an arrangement is particularly useful for turbine systems that have already been set up, where improved monitoring is to be achieved by making the operator aware of the rotor tension state and the damage caused by fatigue. Similar display instruments can be used to determine the

- 52 -

009839/1180

thermal stress and / or the deformation states in the medium-pressure part of large turbine systems and other turbines. In the present case, the control loop for the dynamic surface tension ratios is obtained by executing program 1 ^ 5 and subroutine 156. With this, a voltage _{limit E SL} for the rotor surface is determined _{within function block 172 after the actual voltage E 3 has been determined.} This can, which is assumed for the sake of simplicity, be a fixed value assumed on the basis of the design values and the determination of the service life. Normally, however, the voltage limit Egj. be determined by a function, the value of which can be taken from a table of values in which each table value corresponds to a number of state variables. In the case of a variable voltage limit, the fatigue damage can be automatically graded for different operating conditions if it can be estimated that the overall gain from a more rapid turbine setting is significant compared to the costs of increased fatigue damage.

In function blocks 174 and 176, 20 and the medium pressure part 22 determines the thermal load on the shaft bore and the centrifugal load. In In any case, the thermal stress in the bore is derived from the rotor temperature gradient during the transition section

" 53 'ORiGiNAL! NSPECTED

009839/1180

determined and from the measured steam temperature as well as other characteristic values is calculated in a manner similar to the calculation in the puncture block l64. The centrifugal force is off the turbine speed determined by the tachometer 52 is specified. Block I78 provides a combined load limit as a fixed value or, as in block I72, a number of mutually related values that are the permissible Load for the various operating states.

In a similar manner, the wall stress in the high-pressure housing is determined in function block 180 on the basis of the temperature measurements the housing temperature meter 56 is determined. A fixed or other voltage limit is set in the function block determined.

Since changes in the total temperature T i in the nozzle chamber are the main cause of rotor temperature gradients and since they in turn cause rotor stresses and deformations, the maximum permissible rate of change of T-- _{is determined in a function block 184 depending on the surface tension E 3 determined in function block 164.} determined. To limit the rate of change of T ^, the change in enthalpy of the addition steam and / or the change in the steam throughput is limited. In this case

- 54 -

OHiGlNALlNSPEGTEO

00983971180

the maximum permissible rate of change (dT-j- / dt) _{M is determined} as a puncture of the respective ratio of the surface tension E _g to the limit tension E _ßL from the function block 172.

If the turbine manufacturer has usually specified or recommended maximum values for the permissible rate of change of Tj corresponding to the maximum heating and cooling rates based on its experience, maximum cooling is normally permissible if the rotor surface is exposed to compressive stress due to the previous heating transition; Correspondingly, maximum heating can be permissible if the rotor surface is exposed to tensile stresses from the previous cooling. If a compressive stress acts within the rotor surface, the extent of this compressive stress determines whether the maximum permissible rate of change of T ₁ or a lower rate of change is required. If there is tensile stress in the rotor surface, the extent of this tensile stress determines whether the maximum permissible rate of change T ^ for cooling or a lower rate of change must be observed.

These considerations are contained in FIG. 11, in which a comparatively simple and approximately valid function for determining the quantity (dT,. / Dt) _M within the function block 184 is shown. There the drawn out curve allows

; - 55 - ORtGiNAt. INSPECTEE

009839/1180

the highest value of (dT.j./dt) _M for the heating or increasing values of Tj for all tensile stress ratios and up to 50 % of the compressive stress ratios. ~ ^L between the 50% - and 100 ^ -Druckspannungsverhältnis coat the allowable Knderungszunahrae of T ^ linearly to the value zero. Similarly, the dashed curve 188 allows the highest values of (dTj / dt) _M for cooling or decreasing values of Ty for all compressive stress ratios and up to 50 % of the compressive stress ratio. Between the 50 % value and the 100 # value, the permissible rate of decrease of Tj falls linearly to zero. In this case, the curves 186 and 188 lie one above the other within the region 187. The linear portions of curves 186 and 188 make it possible to generally predict that an existing rotor voltage level, even if not excessively large, is likely to require a restriction in the rate of change of steam temperature .

Further functions with or without prediction can of course be used to determine the permissible rate of change in the steam temperature T i. can be used in the nozzle chamber. These functions can be a fixed Snderungshöchstwert for T corresponding to the level 107 Jiaben or not, ie * for the hochstzulässige Änderiangsgeschwindigkeit bet different values under untersc & iedlicihen ¥ o: rgäben for the operating behavior

ORIGINAL INSPECTED

919122

If the control is operating in load operation, which is determined in function block 190, and if, after comparing the end values in function blocks I80 and I82, the housing voltage is recognized as too high in function block 192, the maximum value (dTj / dt ) .. reduced by a proportion specified in function block 194. The actually possible rate of change of T _T is then determined from the measured temperature and time values and compared with the calculated maximum permissible value of function block .196. If the actual value of dT ^ / dt is equal to or greater than the maximum permissible value, the function block 196 calculates the maximum position setpoint value D _M for the steam valve load. If the case voltage is within the limits, the program proceeds from function block 192 to function block 196.

If the control is operating in speed mode, a determination similar to that described for function block 192 takes place in function block 198 for exceeding the limit of the housing voltage. In addition, by combining the load values of the function blocks 174 and Vj6 and by comparing the combination values for the threshold values of the function block 178 for the high pressure part and the medium pressure part, an examination is carried out to determine whether the drilling load is too great. The calculation of the bore load takes place within the speed control mode, since this is for cold start-up and similar

ORIGINAL! MSPECTID

0Q983Ö / ⁵ ff8Ö

Operating conditions is particularly important. If either the bore load or the housing tension exceed the respective threshold value, the maximum value (dTj / dt) _M from the block 184 is reduced by a corresponding amount within the puncture block 200.

As in the case of a load restriction, the actual rate of change of T _{1 is} calculated and compared in function block 202 with the calculated maximum permissible value (dTp / dt) ". If the actual temperature change rate is equal to or greater than the permissible maximum value, the function block 202 determines the value D _oia , ie in this case the increase in w _{R is} stopped in the block 1 ^ 7. If the housing stress and the bore stress remain within the threshold values, program execution proceeds immediately from function block 198 to function block 202.

When determining D-, and in the general case of Dg ", the maximum rate of temperature change (dT ^ / dt) _{M of} the steam is converted into a maximum target valve position, whereby the steam valves 25 are adjusted from the respective starting position at such a rate of change, that the change in throughput does not take place at a rate of change greater than that corresponds to _{the maximum value (dT.j./dt) M.} So if the inlet steam temperature and pressure and thus also the

- 58 -

ORSGiNAL INSPEGTiD

009839/1180

Steam enthalpy are kept essentially constant in this case, the valve position rate of change is limited so that the rate of change of the steam flow rate cannot exceed a visual wave value, which in terms of the rate of change of the nozzle chamber steam temperature T ₁ leads to the maximum value (dTp / dt) _{M being exceeded} would. The inlet steam valves 25 are expediently set to setpoint values with a somewhat supercritical gain. For this reason and for other reasons, it is preferred to limit the speed of change in steam temperature dTj / dt in the nozzle chamber by limiting the setpoint values D ₀ or O _n within the setting controls, as has already been explained in the previous case. However, if you want to use special function curves that deviate from the function curves of blocks 192 and 202, a control loop can be used for the local analog valve position control loop or for the direct digital definition of the valve position.

Strengthening can be used, creating immediate dynamic restrictions for the steam valve division and thus for the Change in turbine speed and / or turbine load are specified, -when the temperature change rate dT ^ / dt should be restricted.

As a result of using the invention, the performance improved by steam turbines, in particular that will Operating behavior more accurate, more effective and more economical. if

ORIGINAL SHSPECTED

009839/1180

the invention is used in closed control loops, the degree of improvement is increased. The preferred one programmed digital computer for the regulation according to the invention meets the requirements of an effective in an economical manner Operational regulation.

This enables the turbine service life to be extended, because the fatigue monitoring and / or fatigue control can be carried out more precisely in relation to the change in shape stresses. This increase in economic efficiency is accompanied by an improvement in the efficiency of the operating control as a function of the Load, speed or other setpoints reached. The limitations of dynamic behavior are better on the achievable Optimal values or optimal dynamic turbine characteristics matched, especially with turbo generators and similar applications where it is known that level changes in the Work behavior often make restrictions necessary. The improvement achieved is due to certain rules for collecting the fatigue phenomena due to change in stresses bound. The restrictions on the bore loads for the housing tension will be in the context of a dynamic regulation in a permissible manner with the restriction for the rotor voltage combined.

ORiGSNAL INSPECTlD

Claims (1)

Claims; ■; lJ control system for steam turbines, characterized by the combination of the following features: presence of a measuring device for the steam temperature within a heat transfer area on a selected rotor part, presence of a measuring device for the thermal stress of at least one rotor section depending on the steam temperature in said area, presence of an adjusting device for the steam state variables in the said range and the presence of a forced entrainment device for the actuating device to limit the change in the steam temperature within the said range, each depending on the thermal state of the relevant rotor section. 2. Control system according to claim 1, characterized in that said control device for the steam state variables comprises a stage for controlling the flow of steam through said area and that said positive entrainment means the adjustment of the adjusting device is limited, so that the rate of change of steam flow within of the range mentioned is limited. - 61 - OFHGtNAL INSPECTED 009839/1180 '.919122 dd 3 · Control system according to claim 2, characterized in that the adjusting device for the steam throughput steam valves as well as actuators for these steam valves .umfnahm and that the said forced entrainment device can exercise a limitation of the adjustment of these adjusting members. 4. Control system according to claim 5, characterized in that that the steam turbine is a large turbo generator, that the given steam chamber area is the turbine nozzle chamber and that the valve actuator has a control loop, that the steam flow rate control stage also comprises a component to determine the target speed valve setting for speed operation for the control loop that there is also a stage for determining the load target valve position for load operation for the control loop is and that said forced entrainment device limits one of said setpoint actuators in terms of their effect, so that the rate of change of steam flow in the Nozzle chamber is limited. 5. Control system according to one of claims 1 to 4, characterized characterized in that the thermal stress from a comparison between the instantaneous thermal state variables and a predetermined threshold value for the thermal state determined ORIGlNAi INSPECTiD 9830 ^ Ϊ6ϋ ^:: and that the device for determining the thermal Voltage a tachometer for the turbine speed and a temperature gauge for the rotor surface temperature includes. 6. Control system according to one of claims 1 to 5, characterized by means of a detection device for the determination of the total thermal stress modulus of change in the rotor after a predetermined puncture of the thermal state of the relevant rotor section. 7. Control system according to claim 6, characterized by means for collecting the deformation fatigue records of the rotor. 8. Control system according to one of claims 1 to 7 » characterized by further detection stages for the thermal load and the centrifugal load on the rotor shaft within at least one rotor section and by a detection device for determining the rotor shaft load from the thermal load and the centrifugal load, which provides the information about these loads feeds into the forced entrainment device. - 63 Ö 0 9 S 3 d / 1 1 8 0 ORlGiNAL INSPECTED 9. Control system according to one of claims 1 to 8, characterized characterized in that said determination, rule and Force-taking devices include a programmed digital computer. . - ".-. 10. Control system according to claim 9 *, characterized in that that the thermal rotor surface condition according to the relationship • - (T - Tj 1 - Y where Τ _{σ is} the rotor surface temperature, T is the mean total rotor temperature, 06 is the thermal expansion coefficient, V is Poisson's constant, T ^ is the concentration factor and Tg is the puncture of the nozzle chamber steam temperature. 11. Control system according to claim 9 or 10, characterized by a program system controlling the operation of the digital computer, in which at least one predetermined thermal rotor state at least one rotor section as a puncture of the steam temperature as well as limiting functions for the steam valve setting changes as a puncture of this thermal ORIGINAL INSPECTED 009839/1180 - 64 -: Rotor state are stored. 12. Control system according to claim 11, characterized by that the thermal rotor condition is a function of the ratio of the thermal rotor surface condition to a limit value is this thermal rotor surface condition. 15. Control system according to claim 12, characterized by that the limiting effect is within a given range between compressive stress and tensile stress values in Form of a constant function of the thermal surface condition ratio takes place and for compressive stress and tensile stress values that are greater than the constant functional range defining values are inversely proportional to the size of the thermal Surface condition ratio is. - 65 - ORiGSMAL INSPECTID 009839/1180