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500 Steam Turbines

AbstractThis section of the manual explains how steam turbines work and describes steturbine components. It also describes various control devices, auxiliary systemsand startup and operation of a steam turbine. Rerating of steam turbines is alsodiscussed.

Contents Page

510 Historical Overview 500-3

520 Engineering Principles 500-4

521 Energy Conversion

522 Turbine Types and Performance Characteristics

523 Staging

524 Admission

525 Power Cycles

530 Machine Components and Materials 500-17

531 Nozzles

532 Blading

533 Rotors

534 Casings

535 Insulation

536 Steam Chest

537 Diaphragms

538 Seals

539 Bearings

540 Control 500-33

541 Governors

542 Control Strategy

Chevron Corporation 500-1 March 1996

500 Steam Turbines Driver Manual

543 Control Stage

544 Emergency Controls

550 Operation 500-43

551 Slow Roll and Warm Up

552 Critical Speeds

553 Steam Conditions

554 Shaft Seals

555 Seal Support Steam Systems

560 Instrumentation 500-49

570 Auxiliaries 500-50

571 Surface Condenser

572 Lube and Control Oil Systems

573 Piping Systems

580 Rerates 500-50

581 Single Stage Turbines

582 Multi Stage Turbines

583 Rotordynamics

March 1996 500-2 Chevron Corporation

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510 Historical OverviewThe basic operating principles of the steam turbine have been known to man fohundreds of years. The earliest steam turbine on record appears to be Hero’s rotating sphere, Alexandria, 120 BC. Many others followed. Most early versionswere interesting demonstrations of the scientific basis; but not until the last quaof the 19th century did a practical machine emerge.

Working independently, and along slightly different lines, de Laval (Sweden, 18and Parsons (England, 1884) both produced turbines with usable power outputThe de Laval turbine was a single stage, flexible shaft machine, producing abouHP at 26000 RPM as shown in Figure 500-1. The Parsons turbine was a doubleflow, multi stage unit, producing 10 HP at 17000 RPM as shown in Figure 500-2

Development proceeded slowly at first, hampered by conservative attitudes, marials, and manufacturing limitations. Nevertheless, by 1900, Parsons was produgenerator sets of 1000 kW output at 1500 RPM and had licensees worldwide.

From this point, development was so fast that by 1910, every major military nathad largely re-equipped their navies with steam turbine power. Electric power gation for domestic and industrial consumption became commonplace in the devoped world.

By about 1930, the development of the steam path and expansion efficiency wasubstantially complete, and has improved only marginally since then. Consideraroom for improvement remained, however, in the areas of construction, reliabilitand materials for higher pressures and temperatures. Unit power output continuto grow along with population and general affluence until about 1970. At that po

Fig. 500-1 de Laval’s Turbine, 1884. From Steam and Gas Turbines, by Stodola and Lowenstein, Vol. I, 1927.

Fig. 500-2 Parson’s Turbine, 1884. From Steam and Gas Turbines by Stodola and Lowenstein, Vol. I, 1927.



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a decline set in because of conservation concerns, the growth of cogeneration,the development of the gas turbine and alternative energy sources.

Steam turbines will be needed for many years, particularly in such areas as cogation bottoming cycles and process waste heat recovery.

520 Engineering Principles

521 Energy ConversionThe steam turbine converts thermodynamic energy in steam (pressure, temperenthalpy) into mechanical energy (horsepower) in two phases:

• Conversion of thermal energy (enthalpy) into kinetic energy (velocity). • Conversion of kinetic energy into work (horsepower).

Conversion of Enthalpy to VelocityMost single stage steam turbines designed in the USA develop at least sonic velocity in the inlet nozzle row. The first stage in most multi stage turbines develops supersonic conditions (approximately Mach number 1.4). The relationbetween enthalpy and velocity is shown in this equation:

(Eq. 500-1)

where:J = mechanical heat equivalent

H1 = steam enthalpy upstream of the nozzle

H2 = steam enthalpy downstream of the nozzle

v = steam jet velocity downstream of the nozzle

Acceleration of steam occurs in nozzles. Sonic velocity, C, occurs at the throat,narrowest part of the nozzle, if the pressure drop is great enough. C is calculatefrom this equation:

where:Tt = temperature at the throat of the nozzle

k = specific heat ratio (1.3 for steam)

Z = gas compressibility

R = gas constant (85.83 ft lb/lb °F for steam)

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In practical terms, temperature is the only variable in this relationship. For the conditions found in most steam turbines, sonic velocity lies between 1400 and fps.

Critical pressure ratio, Pc, is the required pressure ratio across a convergent nozzfor sonic velocity:

Pc = {2/(k+1)}k/(k-1)

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where:

k = dry saturated steam is 1.3

T1 = absolute temperature at nozzle inlet

Tt = absolute temperature at the throat

Sonic velocity is reached in the throat. If the nozzle has a divergent section dowstream of the throat, the flow continues to accelerate. This acceleration occurs if there is critical pressure at the throat and if the down-steam pressure is belowthroat pressure. Without the required pressure drop, the flow actually deceleratthe divergent section, and turbine power is very poor.

Typical single stage turbines have sonic rather than supersonic velocity steam jfor these reasons:

• It is easier and more economical to design and manufacture a convergent nozzle than a fully-profiled, convergent-divergent nozzle.

• Single stage turbines are often controlled by inlet throttling, again for cost reasons. Inlet throttling requires an excess pressure drop across the turbinallow for the pressure drop across the throttle valve at low flow rates, while maintaining sufficient pressure drop across the nozzles to achieve sonic velocity. With supersonic nozzle design (convergent-divergent), any pressudrop below design causes velocity deceleration in the divergent section of tnozzles and reduces conversion of energy to velocity. Thus, a convergent dgent nozzle is not able to operate efficiently enough at off-design conditionsmake it a worthwhile feature on a throttle-controlled turbine.

Most multi stage turbines have supersonic steam jet velocity (typically 1500-30fps) in the first stage. Control does not usually depend on variable pressure droacross a throttle valve. Therefore, the nozzles face only one set of conditions aare not exposed to excess pressure drop or the shock buffeting that goes with iremaining stages are usually subsonic.

Conversion of Velocity to PowerExactly how this function is performed depends on blade and nozzle design. (Fmore information, see the section on impulse/reaction blading in Section 522.) basic premise is the same for both impulse and reaction design: change of velocreates a force which, if acting on a moving blade, produces power.



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In Figure 500-3, using the impulse design for illustration, the nozzle outlet speeC1, the stage outlet speed is C2, and the blade is moving at U ft/sec. Let us assumthat the steam mass flow is m lb/sec. The thrust on the blade, F, in the directionblade movement, is calculated from the product of the mass flow, m, and the change of velocity.

F = m (C1 Cos ∝1 - C2 Cos ∝2)

Power is simply force times blade speed:

HP = U × F

Maximum power and efficiency occur when C2 is minimum. Impulse stages are most efficient when:

Blade speed ratio (U/C1) = (Cos ∝1)/2

Most single-row impulse turbines have optimum blade speed ratio of approxima0.485 because the inlet angle ∝1 must be positive. The nozzle centerline is usually10-14 degrees to the blades.

Efficiency cannot be greater than Cos2 ∝1 (about 94%). In reality, many losses musbe considered. The stage will probably not exceed 80%, because nozzle efficieis unlikely to be more than 90% to 92%.

Figure 500-4 shows the typical relationship between blade-speed ratio and effi-ciency.

Reaction-design turbines have similar underlying principles, but the analysis is somewhat more complex. This design requires that a substantial pressure drop

Fig. 500-3 Conversion of Velocity to Power Fig. 500-4 Relationship Between Blade-Speed Ratio and Efficiency



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introduced in the moving blades. Optimum performance occurs at a different blspeed ratio. Efficiencies are usually up to 3% higher, but manufacture is more dcult and expensive, because more stages as well as closer clearances are requaround the blade tips.

The Condition LineIf we plot the steam condition as it proceeds through the turbine, the resulting liis called the Condition Line. You can see it on the Mollier Chart in Figure 500-5(point A to point B).

In some turbines, the condition line crosses the saturation line, below which thesteam becomes wet or “condenses.” The separating water can cause much dathrough erosion and corrosion. Usually, in a well-designed turbine, the water is partially removed through strategically-placed casing drains. In practice, the condensation does not begin until the steam enthalpy is 97% of the theoretical “dry” enthalpy. This supercooling is due to the finite time required for the physicprocess to take place. A 3% wet line drawn on the chart below the saturation lincalled the “Wilson Line,” and it is an important consideration in the design of casing drains.

522 Turbine Types and Performance Characteristics

Turbine TypesThere are three bases by which to classify steam turbines:

• Condensing and Backpressure• Single and Multi Stage• Reaction and Impulse

Condensing and Backpressure. In the condensing turbine, the steam is expandedown to very low pressures, typically 2.5 to 3.5 inches of mercury absolute (abo1.5 psia). At these pressures, the steam temperature is low, about 140°F, and the steam is wet, having begun to condense. During the expansion process, the remof useful heat is maximized. All remaining heat is then rejected to a cooling watsystem through a surface condenser. The condensing function, which starts in tturbine, proceeds only to about 90% quality in the last blade row. The condensecompletes the condensation process to 100% water, i.e., 0% steam quality.

The backpressure turbine rejects its unused energy at a pressure high enough of use to other processes. Normally, the exhaust steam, in this case, retains sosuperheat. Contrary to the popular misconception that condensing turbines areefficient than backpressure turbines, the expansion efficiency is virtually identicafor both. It is true, however, that condensing cycles may be less efficient in practice because the rejected heat is not hot enough to be of value. If there were no usethe steam at the exhaust pressure of a backpressure turbine, that cycle too wouinefficient. Clearly, we must look at both the expansion (machine) efficiency andcycle efficiency when doing energy audit or analysis.



Fig. 500-5 Mollier Chart Showing a Typical Turbine Condition Line. Courtesy of the Babcock and Wilcox Company.



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Single and Multi Stage. Staging refers to the number of discrete pressure reductions in the turbine.

A single stage turbine typically has one set of nozzles that drops the pressure ffor example, 300 to 100 psi. It may have more than one row of moving blades bonly one location in the turbine where pressure energy is converted to velocity (kinetic) energy.

A multi stage turbine, on the other hand, has several nozzle rows that reduce thpressure. A typical pressure ratio across a stage is approximately 0.58. Thus, fstages the overall pressure ratio would be 0.586 or 0.038. If the inlet pressure were 1000 psia, the backpressure would be 1000 x .038 or 38 psia. Such exhaust stewould be suitable for process heating.

In reality, the pressure ratio of all six stages would not be equal. Because of theneed to control the steam flow and distribution in the turbine, the first stage, usucalled the control stage, would have a higher pressure drop, typically 0.25 to 0.See Section 543.

Alternatively, the same staging arrangement could take 50 psia inlet steam, expto 1.9 psia (3.8 inches Hg) and exhaust to a surface condenser (cooler). Note talthough the same blading aerodynamic design could work as efficiently on botsets of steam conditions, the power output would be much less at the lower presure, and the loads on the blading would be substantially different.

Impulse and Reaction. The way in which the pressure drop is distributed betweethe rotating and stationary blades distinguishes impulse from reaction turbines.two have very different requirements in blade geometry and assembly tolerance

Impulse is the force created on an object, such as a flat plate or blade, when it a jet of fluid.

The impulse turbine stage drops the available pressure in a nozzle to create maximum velocity. The jet is then played onto the moving blades to convert thevelocity (kinetic energy) into work. There is no pressure drop in the moving bladbut in the ideal situation, the outlet velocity from the moving blades is minimal. Section 521 above.

Reaction is the thrust force resulting from the creation of a high velocity jet. A good example is a fire hose pushing against the direction of the water flow. Thisthrust is the reaction force.

A reaction turbine, like the impulse turbine, has fixed nozzles. The steam jet plaon the moving blades and creates a force by virtue of impulse through momentexchange. In addition to the impulse force on the blade, pressure drops within tmoving blade row, and the fluid is re-accelerated. The force arising on the moviblades as a result of the velocity change within the blade passage is the reactioforce.

The total work on the moving blades is the sum of the reaction and the impulseforces multiplied by the blade speed.



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The ratio of the reaction work to the total work is called “degree of reaction” andtypically limited to 50%.

Steam velocities tend to be lower in the reaction turbine, although total work dois theoretically identical. In reality, the reaction turbine is up to 3% more efficienHowever, it requires closer clearances around the blade tips, is more sensitive variations in blade section geometry, and typically needs more stages for a givepressure ratio.

In practice, reaction turbines are more efficient when the physical dimensions alarge. Impulse turbines, on the other hand, are more efficient when small and atraditionally considered to be more reliable.

Because they face high energy costs, European makers have concentrated on oping the reaction type. U.S. makers, facing lower energy costs, have concentron the more reliable impulse types.

Although much of the original development was in impulse types, many currentmachines combine impulse and reaction blading in the same casing for the besboth worlds. The first stage of all multi stage units is invariably the impulse type

Performance Characteristics

Torque/Speed Relationship. At maximum steam flow, all steam turbines show a similar torque characteristic. The relationship is linear; as the speed reduces, toincreases. See Figure 500-6.

Fig. 500-6 Typical Torque/Speed Relationship for Steam Turbines



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Speed/Efficiency Relationship. At any steam flow, efficiency varies with blade/speed ratio as follows:

Efficiency = A(u/c)(Bcos∝-u/c)where:

A, B, and ∝ = constants depending on the blade and nozzle geometry

u/c = ratio of blade speed to steam jet speed

Figure 500-4 shows a typical plot of efficiency versus speed for a single stage, single blade row turbine.

523 Staging

Velocity StagingRows of moving blades that do not exhibit pressure drop are known as velocitystages. The combination of many such rows is known as velocity compoundingFigure 500-7 (left hand side) shows a section of nozzles for a pressure stage.

Curtis Stage. A subset of velocity compound staging is called the Curtis stage, named after the man who took out early patents. It combines one pressure-dronozzle row followed by two velocity stages (see Figure 500-7). This design wastypical of many early American turbines, and remains so today, particularly for such manufacturers as GE and Worthington (Dresser). In fact, almost all multi sturbines made today worldwide have Curtis design first stages. Typically the presure drop across the stage is large (ratio 0.25 to 0.4).

Pressure StagingA pressure stage, in either impulse or reaction design, comprises a single row onozzles exchanging pressure for velocity in one operation, followed by a row ofmoving blades. In the industry, this arrangement is conventionally known as a dLaval turbine.

The “Rateau turbine” has two or more pressure stages in series (an arrangemealso known as pressure compounding or Rateau staging). Figure 500-7 shows velocity and pressure distribution as the steam flows through successive stagesPressure staging is on the right (downstream) side of the figure.

524 AdmissionA steam chest admits steam into the turbine. The steam is then directed to the stage by means of an arc of nozzles. See Figure 500-8.

The nozzles are divided into sections or blocks. The complete arc of nozzles racovers more than 90 to 180 degrees of total circumference, as illustrated in Figure 500-9. This arrangement, known as partial admission, is an important pathe power and speed control system.



Fig. 500-7 Steam Flow Through Turbine Stages. Courtesy of The Elliott Company.



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Full admission, in which the inlet steam is admitted to the whole 360 degrees, irare. More commonly seen in combustion gas turbines, it would be desirable in

Fig. 500-8 Nozzle Block Installed

Fig. 500-9 Partial Admission With Multiple Nozzle Valves



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steam turbines only at very high temperatures (above 1100°F). The purpose of full admission in this context is to achieve thermal symmetry and avoid casing distotion and warping. This type of steam turbine is not commercially available, although it has been considered in energy-efficient experimental designs.

Note that downstream of several pressure-compounded stages, the flow that stout as partial admission distributes evenly around the circumference, becomingeffectively, full admission.

525 Power Cycles

Rankine CycleAll steam turbines work on some variation of the Rankine cycle. See Figure 500

This cycle is practical and depends on the following factors:

• Compression of the working fluid in liquid phase, points 1 to 2.

• Heat transfer from low temperature source, boiler feedwater preheat, pointsto 2′.

• Boiling to change phase to vapor, boiler points 2′ to 3.

• Expansion of vapor work/heat interchange, points 3 to 4.

• Condensation of residual vapor to liquid phase, condenser, points 4 to 1.

• Re-compression of liquid working fluid.

There are many variations of the Rankine cycle. These are mainly of interest intric power generation. In a process plant, the need for flexibility usually drives thuser toward the simple cycle, although overall thermal efficiency may be lower. Figures 500-10, 500-11, and 500-12.

Fig. 500-10 Rankine Cycle. From Introduction to Thermodynamics: Classical and Statistical by Sonntag and Van Wylen. 1971. Courtesy of John Wiley & Sons, Inc., New York.



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Figure 500-10 shows a basic Rankine cycle without superheat. It has saturatedsteam at the turbine inlet. The turbine runs almost entirely in the “wet” region. Although attractive from a boiler-design viewpoint, this design is not optimal. Noonly is the turbine susceptible to severe erosion, but removing the condensate the steam path to the required degree is extremely difficult.

SuperheatTo surmount erosion problems and improve cycle efficiency, we commonly supeheat the steam to several hundred degrees above the saturation temperature, ashown in Figure 500-11. The maximum practical superheat, also called the metgical limit, is approximately 1100°F. At higher values, exotic, temperature-resistanmaterials (rarely used in the industry) are required.

Also, above 1100°F, thermal distortion of the turbine structure becomes a potentproblem, and keeping the pressure containment tight is difficult. These issues cbe addressed by making the design axisymmetric (full-admission designs, forgecylindrical or barrel type construction, etc.). Although such designs have been successfully deployed, there are no such machines available commercially.

ReheatReheat is a cycle variation in which the partially-expanded steam is extracted frthe turbine steam path, passed back to the superheater, where it is raised to itsinal temperature, and then returned to the turbine, where the expansion procescontinues. See Figure 500-12. It is common to have two or more reheats in utilisteam turbine generator plants. Industrial process plants, however, do not usuause reheat because the gain in efficiency is not worth the added complexity.

Regenerative Cycle. The basic premise of the regenerative cycle is to remove sothermal energy from the part expanded working fluid and add it to the incomingfluid prior to the addition of energy from the outside source (for example, raise tboiler feedwater to the saturation temperature before it enters the boiler). The a

Fig. 500-11 Rankine Cycle With Superheat. Courtesy of the Babcock and Wilcox Company.



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to raise the mean temperature at which the main heating process occurs, thus improving the Carnot efficiency of the cycle.

Recuperative Cycles. Recuperation is another cycle variation in which thermal energy from the spent fluid stream is added to the incoming fluid prior to the maaddition of energy from the outside source. Such cycles are common in gas turplants, where the heat rejection temperature is of necessity very high (1000°F). In condensing steam turbine cycles, recuperation is not of great value because oflow heat-rejection temperature. For back pressure cycles, recuperation may ocsionally be of value.

Topping and Bottoming CyclesA Rankine cycle is often integrated with process-waste heat recovery or added other systems as a topping or a bottoming “cycle.”

These are not true cycles, as the term really applies to the position in the “hier-archy” of energy usage. For example, process-waste heat is used to generate 1psi steam, which in turn drives a critical compressor or generates electric powepass out or exhaust steam from which is used for process heating. The steam turbine cycle, in this case, is the topping cycle.

More commonly, however, the topping cycle uses a gas turbine, and the exhausheat is used to generate steam that is then supplied to a steam turbine. The twturbines can be coupled to the same generator. In this arrangement, the gas tuis the topping cycle and the steam turbine is the bottoming cycle.

Condensing steam turbines are quite suited to bottoming duty because of their tive use of low-temperature energy. Back-pressure turbines or gas turbines, on other hand, typically reject heat at several hundred degrees F.

Fig. 500-12 Rankine Cycle With Reheat. From Introduction to Thermodynamics: Classical and Statistical by Sonntag and Van Wylen. 1971. Courtesy of John Wiley & Sons, Inc., New York.



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530 Machine Components and Materials

531 NozzlesNozzles convert pressure energy into velocity (kinetic energy). The two basic nozzle types are convergent, which achieves sonic velocity, and convergent divgent, which achieves supersonic velocity.

Sonic NozzlesThese nozzles are usually found in small, single stage, throttle-controlled turbin

The sonic nozzle is of converging cross section only. It is usually drilled and mabe reamed to achieve a smooth and continuous reduction of diameter. The pasis usually straight, with the outlet set at an angle of approximately 10 degrees todegrees to the rotor blade row. Efficiency is no more than 90% to 92%.

Major losses are due to the oblique inlet and discharge. Pressure ratio (P2/P1) is usually no more than approximately 0.58, and enthalpy conversion in the steamlimited to about 75 BTU/lb.

Supersonic NozzlesSupersonic nozzles are usually found in the control (first) stage of a multi stagesteam turbine. They can be manufactured several ways, but their form is usuallblade row or cascade, as shown in Figure 500-13.

The nozzle cross-sectional area is rectangular, which better matches the movinblade row entrance, and the flow profile is convergent divergent. See Figure 50

Although a few rare supersonic nozzles are circular or straight/oblique, most arcurved with axial entry and oblique exit. Typical efficiencies are 85% to 90%. These nozzles have a much higher enthalpy drop than subsonic types, typicallyto 200 BTU/lb, which gives velocities of Mach 1.4 to 1.6. They are usually mademartensitic or ferritic chrome steel and cast or fabricated into multi-blade blocksThese blocks are welded or bolted to the outlet side of the steam chest. See Figure 500-8.

Fig. 500-13 First-Stage Nozzle Ring. Courtesy of The Elliott Company.



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532 BladingThe two categories of blade are impulse and reaction. Impulse blades are sometimes called buckets because they function much like the buckets in early wateturbines.

Impulse BladingThe passages between the blades are designed for constant velocity without exsion. The passage cross-sectional area is therefore constant. See Figure 500-1Often, the blade has identical inlet and outlet geometry, although on occasion thsection is rotated to make better use of the available energy and velocities.

For short blades (low ratio of root to tip diameter) the sections are not usually tapered or twisted. See Figure 500-15 for a definition of terms.

The blades commonly have tip shrouds, which may be integral with the blade oriveted strip steel. Sometimes the shroud is laminated with overlapping joins, githe effect of a continuous band. More commonly, however, the shroud band is ishort sections (five or six blades), giving rise to the term “blade packets.”

The primary purpose of the shroud is to dampen blade vibrations and, secondato prevent leakage and flow losses around the blade tip.

Fig. 500-14 Impulse Blading



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Figure 500-16 shows a rotor with two different arrangements of tip shroud. Theblade rows in the foreground have integral shrouds with a continuous steel rolleinto the tip— an excellent design feature. The three rows in the background arelower pressure stages and have a standard packeted arrangement with a steel riveted into position.

All the blades on this rotor are impulse type, clearly identifiable by the sharp leading edges and symmetrical inlet and outlet angles.

A variety of impulse blades are shown in Figure 500-17, each one designed forspecific application. Note the integral shrouds and rivet tangs on the shorter bla

Often, a strip seal (labyrinth) around the tip reduces flow losses past the blade.Theoretically, these seals are unnecessary because there should be no pressuacross the blades. However, in practice there is always some resistance, and thseals do contribute to efficiency.

Reaction BladingReaction blading is considerably more complex in its geometry. The passage abetween blades is variable; the blade cross section is a true aerofoil. Inlet and ogeometry is not symmetrical, and blades are usually tapered and twisted, espeif they are long.

Historically, reaction blades were often free standing, with feathered tips. The feered tips were designed to reduce the impact of damage by inadvertent contac

Fig. 500-15 Blade Nomenclature



Fig. 500-16 Integral and Packeted Blade Tip Shrouds. Courtesy of Dresser Rand Company, Steam Turbine Division.



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the casing. Such contact was common, because tip clearances had to be kept to avoid tip bypassing under the influence of pressure drop across the moving blade. Later designs tend to favor tip shrouds and seals in the interest of reliabi

Dampers for Blade VibrationLong blading often features lacing wire at an intermediate blade radius. See Figure 500-18. The purpose of the wires is to provide friction damping. The wireplaced at the vibration antinode (the point of maximum displacement under vibrating conditions), and the centrifugal force presses the wire radially onto theblade and restrains movement, thus providing frictional damping. Note that the and blade are not firmly fixed together. Relative movement is allowed but is restrained by friction.

Wire damping is less popular now than in times past because a hole at the benantinode causes inherent weakness. (The antinode is the point of maximum bemoment.)

Other forms of damping are occasionally used, but modern designs rely more ocareful prediction of natural frequencies and conservative stressing to avoid prolems. Modern analytical methods make this the preferred approach.

Blade RootsEvery blade needs to be attached to the rotor or disc. This fixture is a critical asof most designs. There are, however, as many fixture variations as there are suppliers, as was seen in Figure 500-17.

Fig. 500-17 A Variety of Impulse Blades. Courtesy of The Elliot Company.



Fig. 500-18 Blades With Damping Wires. Courtesy of Dresser Rand Company, Steam Turbine Division.



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The one common feature is a hook or dovetail. Sometimes the hook is single oshaped, sometimes double. Accuracy of manufacture and fit up is paramount ifload is to be shared between two hooks.

To save on costs, blades are usually assembled onto the disc in a tangential dirtion, with a special locking root to close the radial entry gap.

Axial-entry, fir tree roots are preferred for high-speed or highly-loaded blades, ethough they are much more expensive. Tight tolerances are still required, but thcan be attended to individually and thus more easily.

Materials of ConstructionBlade material is usually 400 series chrome steel, such as AISI 416 or 422, evethough it is somewhat susceptible to pitting corrosion or chloride stress corrosio

Duplex alloys, used occasionally, are more expensive alternatives, but probablymuch better.

Titanium blading is rare. Most manufacturers in the steam turbine industry are umiliar with the required procedures for design and manufacture of titanium blad

Blade Stress and VibrationBlade stress is conventionally categorized as either steady state or vibratory.

Steady state stress is the stress resulting from rotation (i.e., centrifugally induceIt is a tensile stress and is usually calculated using finite element methods.

Alternating or vibratory stress is usually seen as the stress arising from steam (gas) bending. Obviously, there is some force on the blade as a result of the steplaying onto it. As a conservative but practical approach, we assume that the loreaches a peak and then decays to zero as the blade passes a nozzle passage

The critical areas are the fillet between the blade shank and the platform blade attachment detail. See Figure 500-15.

Typically, there are 50 to 100 blades or nozzles in a row, and turbines run at 1815000 RPM. Clearly, fatigue must be carefully addressed in the design stages because of the many millions of stress cycles. As it takes only a few minutes toreach 107 cycles, stress must be kept below the endurance limit.

The Goodman diagram is an essential tool in this work. Figure 500-19 shows atypical diagram with sample values plotted.

Note that as the steady state stress reduces, the permissible alternating stress increases. Note also that the calculated stresses must include the effects of strconcentrations caused by root fillet and firtree radii. Typical blade root fillet stresconcentration factors in use today, for example, are 1.3 to 1.5.

A generous radius and a good surface finish are essential. Scratches and dingscaused by foreign-object damage or poor handling can cause damage quickly. Large low-pressure blades such as those used in the last stage in a condensingturbine are particularly vulnerable.



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The critical stress location is usually at the blade root fillet, or in the first hook othe root itself. The alternating component of stress at this location, when plottedthe Goodman diagram, should be considerably less than the permitted maximuThe ratio of the actual to the permitted stress is called the “Goodman Factor”. Afactor of 2 is marginal, but a factor of 5 is most desirable.

Blade VibrationResonance resulting from the alternating stresses must also be considered. A screening process is used to check for interference between the frequency of thalternating stress (the exciter) and the mechanical natural frequency of the individual blades. The Campbell diagram represents the results graphically. See Figure 500-20.

The purpose of the diagram is to highlight any situation where the blade naturafrequency coincides with an exciter within the normal steady operating speed range. In this situation, resonance can set up and cause blade failure. Under renance conditions, stress is multiplied many times, by a factor dependent mainlydamping. Without damping, the amplification factor (Q) may be up to 100.

Clearly, resonance must be avoided if possible, preferably by careful blade seletion or introduction of damping features, such as lacing wires, tip shroud packeetc.

The example in Figure 500-20 is interesting from a design point of view. The intference between the first lateral blade mode and the nozzle passing frequency occurs just below the operating-speed range. This situation would be barely acable. If the turbine were to stop just short of the normal speed range during acction, it could be disastrous for the blade row in question, unless the blade is verrobust, with Goodman factor at 20 or more.

In the same example, the interference between the blade torsional and twice nopassing frequency is in the operating-speed band, so a closer look is justified. Iblade in question is a short, shrouded or packeted blade, it’s unlikely that a probwould arise. If the blade is a long and free standing blade, additional damping i

Fig. 500-19 Goodman Diagram



raise
justified, and a change to the blade thickness or taper ratio would be needed tothe response frequency above the speed range.
533 RotorsSome rotors are built up (discs shrink-fitted to a forged shaft, as shown in Figure 500-21) and sometimes are solid forgings (integral discs and shaft).

Fig. 500-20 Campbell Diagram

Fig. 500-21 Typical Built-Up, Single-Flow Rotor. Courtesy of the Elliot Company.



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Built-up rotors usually have hot-rolled AISI 4130 shafts (4330 if over about 6 in.maximum diameter) and forged 4140 discs.

Solid forgings are usually found on units running above about 10,000 RPM, espcially if the units are multi stage.

One-piece rotor forgings are not easy to categorize; typically they are similar toA293 class 5. Although this is a tough, moderately strong steel, hot repairs (webuild up, for example) are difficult, especially if the yield point is over 80% of UTS. Welding on such rotors is usually a complex procedure requiring considerexpertise.

534 CasingsCasings perform two functions:

• Hold parts such as diaphragms and bearings in correct relative position. • Contain the steam pressure.

When steam temperatures reach approximately 700°F or above, a double casing design is often used, as shown in Figure 500-22. This design approach allows touter part of the casing to contain pressure without the attendant high temperathus avoiding complications such as creep and thermal transients.

The inner casing, although exposed to the full temperature of the steam path, hrelatively modest pressures to handle. Being lowly stressed, it is better able to handle thermal transient stresses.

During construction, casings are pressure tested to stress levels according to AVIII pressure vessel code. Neither testing nor periodic re-testing, however, is legally mandated.

The initial pressure test, especially of condensing turbines, may be staged. Blapartitions are inserted in the casing at several locations so that the high pressuis tested at high pressure, mid sections are tested at intermediate pressure, etc

If a turbine is re-rated to higher pressure, a partitioned test should be revalidatespite of the major inconvenience, this requirement must never be ignored.

If the original design has been subject to a detailed finite element analysis (FEAand the results verified by comparing them to the strain gauge measurements iical areas, it may be possible to qualify the casing for higher-pressure duty by rning the FEA and comparing it to the allowable stress levels. Otherwise, a hydrtest is needed for safety reasons, if not for legal compliance.

535 InsulationInsulation on steam turbine casings performs three functions:

• Maintains uniform temperature to avoid thermal distortion and internal misalignment.



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ar in

• Protects personnel.

• Increases efficiency.

Uniform temperature is important, especially in multi stage turbines, where rapiuneven changes can cause contact in close-clearance areas, leaking casing joand other problems.

The effect of insulation on efficiency is slight, perhaps less than 0.5%. Avoidingthermal distortion and internal misalignment caused by uneven temperature is fmore important than achieving a small increase in efficiency. In certain cases, itmay even be better to remove insulation altogether than have parts missing or poor repair.

Fig. 500-22 Turbine Components. Courtesy of Borsig Turbines.



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aft

Personnel protection is self explanatory. The dangers of uninsulated casings arslight compared to those of a superheated steam leak from a casing split line.

536 Steam ChestThe steam chest is a small volume between the inlet isolation valve (or trip andthrottle) and the nozzle valves where steam at header pressure accumulates fodistribution to the different nozzle sectors. See Figure 500-23.

537 DiaphragmsDiaphragms are the partitions in the casing between successive stages. See Fi500-23 and 500-24. They carry the stationary blades or nozzles for the downstrstage and the shaft seal between stages. Because they hold the pressure differbetween stages, they need to be strong and rigid.

Diaphragms are constructed of steel or cast iron, depending on pressure differeand user preference.

538 SealsThere are three types of seals:

• Labyrinth seals• Segmental ring seals• Dry gas seals

Labyrinth SealsLabyrinth seals are a close-clearance leakage seal, typically used for casing/shseals and internal stage-to-stage seals. See Figure 500-25.

Fig. 500-23 Turbine Casing Components. Courtesy of The Elliott Company.



Fig. 500-24 Fabricated Diaphragm Designed for 100% Admission

‘

Fig. 500-25 Labyrinth Seal. Courtesy of Elliott Company.

LABYRINTH seal showing how high and low sealing strips combine with stationary baffle to hinder flow of steam. The spring allows the stationary baffle to move away from the shaft if a rub occurs. The heat generated is absorbed by the stationary baffle. This protects the shaft and minimizes rotor damage.

LEAKAGE flow and sealing steam arrangement for steam end of condensing turbine. Conditions at right are typical values.

1. First-stage pressure equals 300 PSIA at full load.

2. Leakoff to turbine stage at approximately 125 PSIA.

3. Leakoff to turbine stage at approximately 40 PSIA.

4. Sealing steam at startup approximately 18 PSIA.

5. Steam and air drawn on gland condenser at approximately 13.5 PSIA.

6. Small amount of atmospheric air at 14.7 PSIA drawn in.



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Pressure drop comes from leakage through the small annular gap (typically 0.00.010 inch). These seals are usually made of phosphor bronze or AISI 316 runagainst carbon steel sleeves. This combination is moderately forgiving in a consituation but can cause a serious “friction whirl” if set too tight.

A friction whirl is a rotor-dynamic response to rotor-stator contact in which therea backwards precession relative to the forward rotation. Because the frequencythe whirl precession is lower than the rotor speed, it shows on vibration-monitordisplay as a subsynchronous vibration, usually at a precise fractional frequencysuch as 1/2 or 1/3 forward speed. The danger in such a whirl is that, being backwards relative to the rotation, the damping inherent in the fluid film bearings is reduced, perhaps critically, allowing the amplitude of the whirl orbit to grow beyond permissible limits. The large amount of available energy produced can cause mechanical damage to the machine structure.

Other material combinations such as nickel aluminide honeycomb and 316 introduce the advantage of abradability. Clearances can be relatively tight, and rotoorbit creates the required clearance by the cutting action of the labyrinth tips. Sfrom friction whirls comes from the low friction coefficient of the seal material combination.

On condensing turbines, a slightly positive steam pressure must be applied to tseal to prevent air from entering the casing or condenser. Thus a supply maniforequired (See Figures 500-26 and 500-27). During normal operation, leakage sfrom the inlet control valve stems is ideal for the purpose, but during startup, letdown steam must be piped in from the supply header.

Carbon Ring SealsLike labyrinths, carbon ring seals work by means of a pressure drop through smannular gaps in a series of stages along the shaft. See Figure 500-28.

Their clearances, however, are closer than those of labyrinths for several reaso

• Carbon has good rubbing characteristics with the shaft material.

• The ring is segmental and can adjust its diameter to fit the shaft closely. Thsegments are held in place by a circumferential spring (Garter spring).

• The ring is arranged to “float” on the shaft, thus following rotor displacemen

These seals have been used for many years and can be found in most single sunits.

Segmental carbon or graphite rings are prone to wire drawing or steam cutting they are left in the stationary condition with a pressure to seal. For this reason, many standby steam turbines are slow rolled. Lately, however, there has been atrend toward retrofitting such machines with dry gas seals. Refer to the Utilities Manual, Appendix E, for the Company’s “Best Practices to Eliminate Turbine SloRolling”



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Dry Gas SealsDry gas seals are a variation of the pump mechanical seal. See Figure 500-29.only essential difference is that the face of the rotating seat is treated with a pro(a rayleigh step) that develops a dynamic pressure under running conditions. Tpressure is enough to separate the faces so that there is little friction. Leakage minor and controlled. Under static conditions, the seal faces close up and thereno leakage.

These seals are unsuitable for condensing turbines because they can seal onlymodest reverse pressure (about 5 psi). They can be used, but setup is very comand ideally requires a double seal and closely-controlled buffer gas pressure.

Such seals are available from Crane, Durametallic and BWIP.

Fig. 500-26 Turbine Shaft Seal Arrangement



Fig. 500-27 Gland Sealing and Leak-Off Equipment

Fig. 500-28 Segmental Carbon Ring Seal. Courtesy of Dresser Rand Company, Steam Turbine Division.



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539 BearingsSee the section on bearings in the General Machinery Manual.

540 Control

541 GovernorsIn any steam turbine, speed and power are the two basic quantities that must bcontrolled. The same governor can control both. However, the required perfor-mance of the driven machine usually does not permit them to be controlled by tsame governor at the same time.

Fig. 500-29 Dry Gas Seal. Courtesy of Durametallic Corporation, Kalamazoo, Michigan.



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For many years, steam turbines had flyweight-type governors or a developed version thereof (see Figure 500-30). All early governors operated in droop modwhich the control speed varies (droops) as the power increases (see Figure 500The simple governor droops automatically. However, many applications must hisochronous control (constant speed regardless of control output), which requircompensating the governor, usually hydraulically. The compensation on some governors is adjustable, allowing the amount of droop to be controlled. Sometimdroop is controlled to improve stability and sometimes to keep driver output at aconstant speed, such as is often required for electric power generation.

Today, most mechanical governors are hydraulically compensated to eliminate control droop, and are coupled to a hydraulic servosystem to operate the steamcontrol valves.

Some direct-acting (no servo) systems are still available commercially for smallnon-critical drives (See Figure 500-32). These generally do not give adequate performance for refinery process applications.

Electronic governors and electric servosystems are becoming more common. Tcan be very expensive and difficult to justify for normal refinery use but can provide many benefits:

• Eliminate the need for high-pressure control oil systems.• Provide better resolution of control.• Result in a narrower dead band.• Operate automatically (depending on setup).• Provide compatibility with computerized plant-control strategy.

Many electronic governors are analog. Digital types are becoming more commoand are suitable for plant digital control systems (DCS) interface. Although they

Fig. 500-30 Flyweight-Type Governor. Courtesy of Woodward Governor Company.

Fig. 500-31 Governor Droop Phenomenon. Courtesy of Woodward Governor Company.



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relatively expensive because it is early in their design life-cycle, competition amsuppliers is beginning to bring down the cost.

Some Typical Governors

Woodward PGPL. This governor has been used for many years on critical, vari-able-speed drives such as centrifugal compressors that require a NEMA class Dgovernor. NEMA class D calls for speed control to within 0.25% of set point andin the event of sudden loss of load, maximum speed rise when operating at rateconditions is 7%. Also, control speed must not vary by more than 0.5% across tpower band 0% to 100% full load.

It is an isochronous governor, controlling at a constant speed, irrespective of threquired load. See Figure 500-33.

The PGPL governor is driven through a reduction gear from the turbine non-coupling end. It typically ranges from 70% to 110% of rated speed. The output by rotation of a torque arm. The torque is quite low and normally requires a sersystem.

Woodward PGG. This is a more complex governor, which can operate in isochrnous mode, droop speed control mode (see Figure 500-34), or load control mosubset of droop speed control. It is typically used to load control a fixed-speed machine, such as a synchronous generator connected to a utility supply grid, fowhich NEMA class C is required.

Fig. 500-32 Section of a Single-Stage Turbine and Governor System. Courtesy of The Elliott Company.



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NEMA class C calls for speed control to within 0.25% of set point and no more than 7% speed rise on sudden load loss. A 4% increase in steady-state controlis permitted for load reduction from 100% to 0% load, which can be interpreted4% maximum droop.

The PGG governor can operate in true isochronous mode. More often, howeveoperates in droop mode because the speed is dictated by outside influences. Fexample, the frequency of the local utility supply to which the drive is connected

Fig. 500-33 Isochronous Control

Fig. 500-34 Droop Control Single Machine



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can affect the speed. The droop is adjustable from 0% to about 10%. The effecchanging droop would show on Figure 500-34 as a change in the slope of the control line. Normally, droop would be set at about 3%.

Speed is selected by adjusting the speed input control. This adjustment translathe control line up or down the page.

Load is also selected by adjusting the load-control input. This adjustment translthe control line sideways across the page.

A typical operating scenario would be to bring the generator up to 95% of synchnous speed manually, set the load at zero, set the droop at about 3%, let the aumatic synchronizing gear raise the speed to synchronous, and to synchronize automatically. Finally, the load control would be manually adjusted to give the required output. Alternatively, the final 5% speed and synchronizing can be donmanually.

Woodward TG13. The TG13 is a simple unit intended for non-critical service. It a NEMA class A governor with 10% fixed droop as standard, a 13% speed increfor sudden load loss, and a 0.75% speed resolution. The 10% droop gives goodcontrol stability with single-valve control of single stage, general-purpose steamturbines, without the need for a servosystem.

Typically, however, the TG13 cannot be used with steam pressures above aboupsig, depending on valve size, as the control valve loads are too high.

The TG13 has many refinery applications because it is ideally suited to single sturbine pump drives. It is often used in upgrades from manufacturers’ standard gral non-servo governors, such as those found on Elliott and Terry general-purpturbines.

Electronic GovernorsElectronic governors are becoming more common. They typically have very gooresolution (0.1%), a narrow deadband (0.1%), and an excellent control range (1to 100%). They can be extremely fast and commonly achieve load-loss speed-rbetter than 5%.

A number of electronic governors are on the market. They can be preprogrammto run drives up to speed smoothly, either at a predetermined rate or in stages. it is possible to be absolutely certain that critical speed bands are traversed proand promptly.

Most units rely on a toothed wheel and magnetic or proximity probes to generathe speed signal. Some units offer redundancy by duplicating circuits. API 612 fourth edition calls for duplication of the pickups, but the better units already offthe vastly superior two-out-of-three voting logic.

Woodward 2301. The 2301 is an analog unit intended for mechanical drives andgenerators. It can perform isochronous or droop speed control, isochronous or droop load control, and load sharing.



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Woodward 509. The 509 is an electronic unit intended for both fixed-speed and variable-speed drives. It would typically replace the PGPL or the PGG. A digitalunit, featuring two-out-of-three voting logic, it has a high degree of internal faulttolerance as well as some self-diagnostic capability. It can be maintained on linaccepts analog input commands, and communicates with DCS’s, LAN’s, and various input/output devices.

The control output of the 509 is to an electric servomotor, e.g., the Woodward T25, which operates steam-control valves either directly or, with the appropriate interface, through a hydraulic servosystem.

A point often overlooked in the quest for speed of response is that the governoreven the old mechanical/hydraulic flyweight types, is usually the fastest part of the control loop. Trying to increase speed without speeding up the servo is pointlesHowever, speeding up servos can lead to instability.

542 Control StrategyThere are several control strategies but only two basic parameters to control in steam turbine operation, speed and load. On occasion, secondary functions, e.turbine pass-out flow or pressure, also need control.

Speed ControlThe most common control mode in refinery applications is speed. Speed has twsubsets, isochronous and droop.

Isochronous Control. Isochronous means constant speed or, more precisely, constant speed irrespective of any other consideration, such as load. Once thegovernor has been set for a speed, nothing changes the speed indication of themachine in the steady state condition.

A short-term speed “excursion” may occur, however, when load changes are fathan governor response time. NEMA SM23 classifies the allowed “excursion” focomplete instantaneous load loss from 100% speed and load.

Figure 500-33 shows the isochronous function graphically.

Droop Control. Droop control is more complex. In this mode, speed changes (droops) slightly with load variation (increase). Droop control is used for one of two reasons: to maintain speed stability on a driven machine with poor stability characteristics or to gain control over load when the permitted speed band is extremely narrow or dictated by an outside influence such as the frequency of tlocal utility. See Figure 500-34.

Load ControlLoad control is the normal way to run electric generators. The two basic types adroop control and direct load control.

Droop Control. For an isolated generator, control can be either isochronous (Figure 500-33), or droop (Figure 500-34). The isochronous condition is a spec



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subset with zero droop, and the droop is not necessary unless the goal is to patwo generators, either pairing them with each other or pairing one with the utilitycompany.

The basic principle in parallel operation is to make one generator the base-loadmachine and to make the other the load-swing machine. One machine (or the utility) dictates the speed by being on isochronous control. This machine takes load swings. The other generator is on droop control and supplies a fixed load. load can be set at any position from 0% to 100%. This mode of control is generreferred to as “droop isochronous.”

Droop isochronous control is illustrated in Figure 500-35. The base-load gener-ator is set up to take, for example, 90% of its rated load. Because the speed is other means—either by the utility or by the swing machine—it cannot deliver another load unless the speed changes. Because the governor output position is to this set of conditions, which in turn puts the steam admission valve in a uniquposition, the steam flow is fixed.

If the total load is greater than 90% of the base-load machine rating, the swing (isochronous) machine assumes the difference. However, if the total load is lesthan the set point on the droop-controlled machine, the swing machine unloadscompletely and the generator “motors” the turbine. This could be catastrophic, the turbine can overheat and be damaged unless a supply of “cooling steam” ismaintained. Fortunately, cooling steam is usually available and the generator ohas and anti-motoring protection or trip.

This system works well enough and can be performed by any NEMA C governosuch as the PPG. With such governor types, any number of machines can be controlled in parallel, but only one machine can be isochronous. All the others mbe on droop control but can be on a different droop percentage.

Fig. 500-35 Droop Control Dual Machine



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There are two major inconveniences in using dual-machine droop control:

• If the base-load machine separates from the swing (utility), a step change occurs in the load the base machine sees, and a speed and frequency chaoccurs on the droop-controlled machine. If the base machine picks up loadspeed or output supply frequency decreases (droop). If the load is droppedspeed rises. The changes may present difficulties for the electric machinerythe dependent system or for the two generators that must be re-synchroniz

• Also, the total load swing of multiple-droop units is limited to the capacity othe isochronous machine. Beyond that, one of the droop machines may tryexceed full load or may reach zero load. The protection systems describedabove are usually in place, which means that one or more generators will pably go off line.

Modern electronic governors give us some options in dealing with the limitationmechanical governors.

Direct Load Control. With the addition of load control, multiple units can run in parallel on isochronous control. A detailed description can be found in supplier manuals such as Woodward manual 01740B Power Management. A brief descrip-tion follows.

The system has two critical components, a load sensor and a summing junction

The load-sensor task is usually carried out by voltage and current transformersthe generator output line(s). The signals are reduced to a simple signal; for example, 0-6 volts DC would represent 0% to 100% generator output (kilowattsThis signal is fed via a bridge circuit to a summing junction on the governor. Thsumming junction also receives input from the steam-turbine speed sensor (toowheel and pickup) and the speed setting input. The output of the summing juncis used to reset the governor, and the governor output repositions the steam-adsion valves through the servosystem.

Introducing the load signal to the summing junction biases the set point to makegovernor pick up load. The signal is fed through a bridge circuit to permit addi-tional control inputs (biases). See Figure 500-36.

Two or more systems can be balanced through the bridge and each held at a fipercentage of its rating. If the total load changes, all units vary their actual outpstay at the same percentage of total.

Electronic governors with these capabilities are on the market from several suppliers such as Woodward and Tri-Sen.



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543 Control StageThe governor does the actual controlling by positioning the inlet steam control valves.

Small turbines normally use a throttling control method. Unfortunately, throttlingreduces cycle efficiency because the available enthalpy drop is reduced. See thMollier Chart in Figure 500-5. A typical enthalpy drop is shown by the conditionline AB. The throttling process is typically AC, reducing the work process to CB′. The lost energy appears as additional temperature (quality) in the exhaust stea

The effect of throttle control on efficiency is shown in Figure 500-37. At low powoutput, steam usage is disproportionately large and efficiency low. The Willens represents the best performance achievable, theoretically, given the blade and nozzle design, operating speed, and supply steam properties in question. Cleaany losses puts the operating point above the Willens line.

An ingenious way around this problem eliminates over 90% of the loss in most cases. The inlet flow is split into several parallel paths. See Figure 500-9. Eachflows through a valve before reaching the nozzles. The valve opening is sequenso that only one valve at a time is actually throttling. The other valves are eitherfully open or fully closed. No valve starts to open until the preceding valve is fulopen.

The sequence is preset and the valves are opened in turn by a rack or a camshTypically there are 3 to 7 valves. The most efficient arrangement is to have the smallest valve modulating at normal operating conditions.

Fig. 500-36 Balanced Load Bridge. Courtesy of Woodward Governor Company.



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Figure 500-38 shows the effect on steam flow and power. Compare it to Figure 500-37. Without throttling losses, the Willens line is straight. By increasinthe subdivision of the inlet flow, the turbine can operate more efficiently over a wider range of conditions.

544 Emergency Controls

OverviewIn the event of a sudden loss of load, e.g., a coupling failure or a generator circbreaker trip, a turbine rapidly accelerates. NEMA SM23 lays down the limiting speed excursions for loss of load but does not address loss of driven inertia. EvNEMA SM23 is fully satisfied (107% speed peak), a coupling break can cause rapid acceleration, and runaway speed—1.5 to 2.5 times design speed—can bereached in 2 to 5 seconds.

Immediate intervention is required to prevent a serious self-destruct incident. Inextreme cases, the turbine casing is usually damaged in some way to allow a psure release of steam. Blades are typically torn off, rotor discs centrifugally dilaand plastically deform, shafts bend or break, coupling hubs and shaft ends are ejected several hundred feet. Such self-destruct events start as low as 140% despeed for machines such as generators.

Overspeed ProtectionTraditional design of protective systems was based on a spring-loaded bolt set eccentrically in the rotating shaft. Above the set speed, centrifugal force dislodgthe bolt against the pre-loaded spring. The dislodged bolt, now protruding from

Fig. 500-37 Single Throttle Control Fig. 500-38 Throttle Control Alternate Method



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shaft, trips a latch, which in turn allows a quick-acting valve to close under sprinpressure.

Some systems are all mechanical, some partly hydraulic. In either case, the enchain of events takes place in about 0.1 to 0.5 seconds and is usually sufficientprevent damage.

More recently, electronic systems have been developed in which the eccentric band latch are replaced by a toothed wheel and magnetic pickups. Electronic processing of the signal sends a command to the quick-closing valve which clounder spring pressure in the traditional manner. The electronic part of the sequis very fast (elapsed time below 0.01 seconds), but the controlling factor is the tthe valve takes to close (0.25 to 0.5 seconds).

Sentinel Valves. If the condenser vacuum is lost, the exhaust pressure of a steaturbine rises rapidly, exacerbated by the speed control, which increases the steflow to compensate for loss of efficiency and power.

The pressure rating of the exhaust hood is usually low (30 to 50 psig). A pressurelief valve sized for maximum flow must be present to supply overpressure protion.

For many years, manufacturers fitted a sentinel valve on the exhaust casing, rigto blow a whistle at approximately 15 psig. However, the sentinel valve is not a legal protective device in any country in the world and must never be used as snor must it be present alone. There are, in fact, many locations where sentinel valves are illegal even if used in conjunction with a pressure-relief valve.

550 Operation

551 Slow Roll and Warm UpStart-up of a steam turbine is a critical operating phase. Inadequate procedureseasily cause reliability problems. A turbine can be started in very different waysdepending on its size, number of stages, and inlet steam temperature.

See “Best Practice to Eliminate Turbine Slow Rolling” in the Appendix E of the Utilities Manual.

Three potential problems arise from poor warm up or rapid starts of the turbine:

• Thermal Transients: Thermally induced stresses in heavy or thick metal comnents can result in fatigue cracks after surprisingly few start-ups.

• Differential Expansion: Differing warm-up rates of components such as casings and rotors can cause loss of critical clearances.

• Condensate: Accumulation of condensate in the inlet piping or in the casingcan cause severe impact damage if the turbine is allowed to accelerate beyslow roll.



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Thermal Transients

Heat takes time to soak through any material, metals included. The property thadescribes this behavior is called thermal diffusivity. This property has dimensionft2/hr and is defined in the following equation:

α = k/(δc)where:

α = thermal diffusivity

k = thermal conductivity

δ = density

c = specific heat

Some typical values are:

Note the wide range of values. An in-depth study is somewhat beyond the scopthis manual. However, note that the typical construction material for steam-turbrotors and casings has a low value. Also, note the very low value for chrome stewhich is used from time to time in special applications.

The lower the value, the longer and slower the warm-up period required to avoithe risk of crack propagation. Other factors affect the peak transient stress as wsuch as elastic modulus and thermal coefficient of expansion.

Typically, metal thickness of less than about 3 inches and steam temperatures about 700°F cause no problem. Thus a single stage turbine up to about 1000 bhcan probably go on line cold, without warm up. No harm comes from thermally induced stresses. However, the steam header must be properly warmed and cosate removed to avoid re-entrainment of condensate into the steam flow and intion into the turbine.

By comparison, a multi stage turbine with a heavy drum-type rotor 12 inches ormore in diameter, running on 800 psi steam, may have a problem and could decracks in the rotor core after a few dozen rapid starts. A rapid start is a functionseveral factors and has no precise definition. However, slow-roll, warm-up perioof several hours are often quoted by manufacturers. Take these recommendatioseriously. A slow roll ensures even heating and a straight rotor. It provides enoutime to make sure that the rotor is warmed to the core and that no thermal stresremain before centrifugal stresses are added by rotation at speed.

0.5% Carbon steel α = 0.57

Copper α = 4.34

Aluminum α = 3.48

Glass α = 0.05

12% chrome steel α = 0.28 (typical)



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If a problem develops, it starts as crack growth originating at the center of the roUsually, no symptoms appear until crack growth accelerates at a dangerous ratthis point two symptoms become observable:

• The vibration signature changes. A strong 2X or 1X signal develops, whichnot explainable by such conventional diagnoses as misalignment, balance,

• Ultrasonic examination of the rotor may reveal diametral or radial separatio(cracks) close to the rotor core.

A large cracked rotor can be repaired by hollow boring (bottle boring) to removethe damage. Because metal at the rotor center adds little strength or stiffness, inot absolutely necessary. Before the repair, however, check out the rotordynamas the reduced weight pushes the first critical speed up towards the operating r

Only very large rotors can be treated this way. Typical refinery drives are usualltoo small for a bottle-bore job and must be scrapped.

Differential Expansion

Single Stage Turbines. Single stage turbine rotors are short and not likely to suffeloss of internal clearance due to differential expansion. Typically, the rotor is located axially by a thrust bearing at inlet or the high-temperature end. Thus, exsion of the rotor, which gets hot faster than the casing, tends to increase the intclearance between the nozzle outlet and the rotor blade leading edge. Also, betypical single stage turbines do not have or need tip seals on the rotor blades, cclearance is not a concern.

We can conclude that a single stage turbine can have a fast, cold startup withousuffering internal damage from loss of internal clearance. Some thermal bowingthe rotor may occur, but the relatively generous clearances prevent rotor-stator contact. A short period (minutes) of rough running (high vibration) may be observed, but shouldn’t cause concern. Most single stage machines are robustenough to withstand vibration for a few minutes at each start.

Eliminating the slow rolling of standby machinery can save steam. Of course, thturbine must have the configuration described above, and the driven machine mbe considered as well.

Multi Stage Turbines. Multi stage turbines with their longer rotors present a verydifferent scenario. Internal clearances could easily be compromised during the warm-up transient state unless the warm up is slow. Fortunately, the warm up isusually slow for other reasons.

Also, attempting to run at speed with a rotor that is not fully stabilized thermallycould cause imbalance and result in major damage.

CondensateCondensate can accumulate in the casing of turbines during initial warm up. It calso accumulate in condensing turbines.



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Most single stage turbines have self-draining cases and do not need casing draunless the exhaust header is higher than the casing. In this situation, the lowesof the exhaust system should be steam trapped.

In a multi stage condensing machine, the last two or three stages operate belowsaturation line and are wet. If water is allowed to accumulate or pass downstreafrom stage to stage, severe erosion of the blades occurs, particularly at the tips

The wet region of the casing must be provided with drain points at each stage tallow the condensate to exit the casing at both low pressure (during warm up) aat normal operating pressures. The drains help prevent erosive damage.

In the wet region, the stage spacing is usually set further apart to allow adequawater separation. If supply-steam quality or superheat is significantly reduced, twet zone extends upstream where the casing is ill-equipped to deal with it. Severosion of blades, nozzles, and even discs and shafts can occur. On occasion, rotors have to be replaced.

Casings are usually warmed by allowing a small amount of steam to flow througthe inlet nozzles and vent into the atmosphere until the exit steam is dry.

The steam flow is usually controlled by hand throttling on the steam inlet isolatiSometimes a small bypass valve around the inlet trip and throttle valve serves tpurpose.

552 Critical SpeedsA critical speed or resonance occurs wherever a natural frequency coincides wexciting frequency or running speed. Normally, turbines are designed to avoid rnance in the normal operating speed range. However, most turbines have to pathrough the resonance when accelerating to operating conditions. Typically, theical speed band is about 200-500 RPM wide. The single most important task is to keep the speed moving—never allow it to remain constant or otherwise settle out within that band.

Many of the newer electronic governors have start-up sequencing capability ancan be programmed to avoid undesirable conditions. Older manual systems, however, require that the operator pay strict attention.

Blade resonances are, perhaps, even more troublesome than rotor criticals. Tharise when the nozzle passing frequency coincides with the blade natural frequAgain, these potentially harmful resonances do not normally occur in the operarange. But at low speeds (slow roll) there may be some interference, particularllong blades in large condensing turbines.

Blades can be designed to be robust enough to run on resonance at reduced sSuch resonance is not unusual for high-pressure blading but is very rare for lowpressure (long) blades.



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553 Steam ConditionsRefer to the utilities manual for a detailed treatment of steam and boiler-water chemistry issues. The following is an overview from a turbine standpoint.

Steam has many parameters:

• Pressure• Temperature• Enthalpy• Entropy• Purity• Superheat• Quality (dryness)

The following sections summarize how these parameters affect the performanca turbine.

Pressure A turbine is rated for a given inlet pressure. If the pressure is low, less steam floand control valves open to maximum positions, beyond which, rated power is nreached.

TemperatureSmall changes have little effect. At constant steam mass flow, the best efficiencspeed increases proportionately with the square root of the absolute temperatuand the best efficiency power increases in proportion to the absolute temperatuThe limiting mass flow at full throttle, however, reduces proportionately with thesquare root of the (increasing) absolute temperature.

EnthalpyEnthalpy is the total heat or energy potential of the steam.

EntropyEntropy describes the usefulness of the steam’s enthalpy.

PurityPurity refers to the level of contamination. Ideally, contamination is minimal, butoften the level is high enough to cause several problems:

• Chlorides come from impure feed water and are typically carried over from boiler as liquid spray, a fault condition, through the superheater. There theyout and become solid particulates in the steam stream. Finally, they get intoturbine, where they re-dissolve at the transition zone, causing chloride strecorrosion and pitting of the blades. The typical 12% chrome steel blades arvery susceptible. Stress corrosion cracking may also occur in high-stress asuch as disc keyways and shrink fits.



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• Silicates come from poor demineralization plant performance. Many modersystems use reverse osmosis, which gives a relatively poor silicates purity no scope to adjust. In critical cases, some form of polishing is needed.

As the steam saturation temperature in the boiler increases, silicates in theboiler water show an increasing tendency to vaporize. The vapor passes through the superheater and later precipitates in the turbine, leaving a hardscale. Although the scale is metallurgically harmless, it causes poor aerodynamic performance and can be extremely tenacious. The only sure cure is water jet or glass bead blast the rotor during a teardown. Most steam turbinmanufacturers call for steam silicate contamination to be under 20 ppb for psi systems. Higher pressures call for lower figures (e.g., 1500 psi systemsneed 5 ppb or lower). See the utilities manual.

• Carbonates come from boiler carryover, poor boiler solids control, and CO2 contaminated feedwater. They leave a soft deposit that can quickly choke unused nozzle block sections. These solids are harmless, apart from their temporary effect on efficiency and power output, and are easily removed bywater washing.

SuperheatSuperheat is the margin of temperature over saturation. (Otherwise the term is synonymous with temperature.) Ideally, superheat is sufficient to keep all but thlast few stages above the Wilson line (97% dry). Because many systems in thecompany do not achieve this standard, critical components erode and must be rebuilt unnecessarily and at great cost.

Quality Quality refers to the dryness of the steam. Conventionally, 90% quality means 9dry saturated steam plus 10% by weight of water at saturation temperature.

554 Shaft SealsThe purpose of shaft seals is twofold:

• To keep steam inside the casing.• To keep air out of the vacuum condensing system.

There are two basic seal types:

• Labyrinth • Carbon ring

Labyrinth SealsLabyrinth seals are probably the most common seals on large turbines. The predrop is created by allowing leakage through small clearances. See Figure 500-2for a typical arrangement.



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Figure 500-27 shows a typical seal support system. Each manufacturer has a different way of arranging the parts. For details on each system, refer to the mamanual.

Carbon Ring SealsCarbon ring seals are more common on older equipment and small back-pressunits. They function in much the same way as labyrinth seals but have segmencarbon rings instead of steel or bronze strips. Clearances are tighter because tcarbon rings are much wider than the single steel strips. However, there are fewrings, and each has a higher pressure drop.

Support systems are similar to those for labyrinth seals.

Carbon ring seals have one major problem. They have a tendency, in standby ement, to wire draw or steam cut the working faces if there is any leakage. In a uheld ready for APS, this can be a major issue because of backleakage from theexhaust steam header.

555 Seal Support Steam SystemsThe purpose of the seal steam system is to control the steam leakage and to pingress of air to condensing systems. Figure 500-27 show a typical arrangemenalthough most real systems vary somewhat in the details. Always refer to the mfacturer’s manual.

Supply steam in Figure 500-27 comes either from the turbine casing, selected tgive an appropriate pressure, or from the steam header. The header supply is uonly during start up, when casing pressure is too low. As an alternative to seal steam drawn from the casing, many manufacturers use leak steam from the steseals of the inlet control valve rack.

An eductor is sometimes used to draw off any excess flow and thus prevent escof seal steam close to the turbine bearings (see Figure 500-27). This action draair at the seal, and then the mixture of air and steam is sent to the seal steam condenser. Note that this is not the main condenser, which needs to run as freepossible from air ingress (incondensibles).

560 Instrumentation

Steam PathInstrumentation on a single stage turbine should include, at minimum, pressuretemperature of the inlet and exhaust steam. If the pressure downstream of the control (throttle) valve is included, efficiency monitoring is straightforward, as loas the exhaust steam is not in the condensing state.

Multi stage turbines should have temperature and pressure at inlet and at the fistage. This point is also known as the wheel chamber and is actually the pressudownstream of the control stage.



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The first stage pressure is a useful indicator of blade condition. When used in conjunction with a mass flow indication or an alternative, such as known poweroutput, an increase in first stage pressure indicates increase of flow (loss of blafouling, etc.).

On backpressure turbines, exhaust temperature and pressure are needed for eciency monitoring.

On condensing turbines, efficiency is difficult to monitor, as steam quality indication is also required. Because no such instrumentation is available, a condensincalorimeter is used. But it yields variable test results.

The above and other instrumentation is not in any way unusual. For details refethe Instrumentation and Control Manual.

570 Auxiliaries

571 Surface CondenserSee “Troubleshooting Surface Condenser Vacuums” in Appendix L of the Utilities Manual.

572 Lube and Control Oil SystemsRefer to the General Machinery Manual.

573 Piping SystemsRefer to the Piping Manual.

580 ReratesRerating covers any physical work done on the turbine which either changes thsteam inlet capability or changes the output characteristics (speed, power, effi-ciency, reliability). Each case must be assessed on its own merits, and assistanfrom the manufacturer or a specialist may be necessary.

Here are a few examples of the sort of work that can be done.

581 Single Stage TurbinesGenerally, single stage turbines use standard, off-the-shelf designs, with very lithat is engineered for a specific application. You may have to refer to the makercatalog or contact the maker directly to establish the maximum rating of the frasize. The maker can indicate which components keep performance below the fsize maximum. A competent engineer can then address those components.



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Nozzle BlockTypically, the nozzle block can be changed to bring power up to the frame-size maximum. One possibility—re-drilling the nozzles to enlarge them—is not recomended because it increases the bending stresses on the blades appreciably arequires that a maximum nozzle diameter be established and the blade stress rculated. A user probably wouldn’t want to risk running above the blade-stress erience of the existing design, especially without proof of reliability at those streslevels.

A better solution is to use more nozzles—as many as the steam-chest dimensioallow. They are probably available in a standard block from the maker. The addtional nozzles increase the power and torque of the turbine without changing thblade stress or the optimum speed of the turbine.

In extreme cases, efficiency could be affected. Also check the coupling and drivequipment stressing.

It will not usually be necessary to revisit rotordynamics issues.

BladesUsually modifying blades on a single stage unit accomplishes very little. It may help to change from a single-row to a two-row Curtis arrangement if the operatispeed is less than 60% of the turbine best-efficiency speed. However, in additioa rotor changeout (or at least a disk change), this arrangement requires the instion of a reversing blade row between the two moving rows.

Such a modification is fairly straightforward in most respects, unless it requires machining of the case to locate the new blade row. In this event, a casing stresysis is needed or, at minimum, a pressure test to 1.5 times the casing-design psure. Note that the actual test pressure selected is a function of the design temperature and may in fact be much higher than 1.5 times actual operating prsure.

Change of Steam PropertiesChanges of stream properties within the frame capability are permissible and straightforward.

Increasing pressure and superheat has a major effect. Power increases accordP × T0.5 where P is inlet absolute pressure and T is inlet absolute temperature. efficiency does not change noticeably, but the best efficiency speed increases iproportion to T0.5.

Check that pressure and temperature design values are not exceeded, or compstress analysis and/or pressure test. Increases in flow beyond the manufacturelimits are frequently easy to achieve. But beware; excess flow is likely to overlothe thrust bearings and couplings.



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582 Multi Stage TurbinesUnlike the single stage turbine, a multi stage turbine is usually an engineered product. Many components are already operating close to their maximum capability. Do a complete and careful analysis, including all components, before attempting any design changes.

CasingsA casing is designed as a pressure containment and, whether cast or fabricatedusually limited by its stiffness, not its strength.

Pressure tests are usually carried out during manufacture and not repeated theafter. The “new” pressure test is calculated to give a stress equivalent to 1.5 timdesign at the design temperature. It is often close to twice actual operating pres

A casting often has a potential pressure capability much higher than the applicacalls for. Take care in a pressure rerate situation. Any recalculation shows only gross excess capability. A new higher test pressure representative of the new oating pressure could reveal a casting fault not uncovered during original testinglower pressure. Alternatively, a careful NDT such as U/T or x-ray could reveal hidden faults. Re-verification by hydrotest is recommended.

The hydrotest will probably have to be staged with internal baffles to test the HPend adequately without over-pressuring the LP end.

NozzlesMulti stage turbine nozzles are treated much like those in single stage turbinesExtra nozzles can often be added by substituting nozzle blocks in the steam chThe control-stage blading probably does not need changing, but the downstrea(Rateau) staging has a bending stress increase. The Goodman diagrams mustredrawn and evaluated for those stages. See the subsection on blades. Also, thadmission valves probably need to be enlarged. Velocity through the valves shonot exceed about 120 feet per second.

BladesThe control stage is not usually a limitation because it is designed to run partialadmission. Of the remaining stages, the first and last rows frequently pose a problem for different reasons.

First Pressure Stage. The first pressure stage is designed to run at or close to fuadmission. Consequently it has very little spare open area and may be the first to choke (sonic velocity in the throat). If further pressure is allowed to build up ahead of this stage to “force” more steam through the turbine, the function of thcontrol stage is affected; power is lost and control may become marginal.

In many cases, choking on the first pressure stage is a turbine casing limitationso, an uprate can still be achieved by removing this row, making the second presure stage the first. This new first pressure stage now has about two times highmass flow capacity than it was originally designed for. The remaining stages distribute the pressure drop somewhat evenly, but tend to load up the last stage



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Most likely all the pressure staging must be redesigned to take full advantage ohigher pressures. This is a difficult task. As the enthalpy drop per stage becomegreater, one or more stages may approach choke, and the turbine may becomeliable and inflexible.

Last Pressure Stage. The last stage has the longest blades and the highest probbility of limiting blade stress. Increased steam flow increases the bending (vibratory) stress. It is essential to revisit the Goodman diagrams for all stages, especthe last row or two. Having Goodman factors (ratio of permissible alternating stto actual alternating stress) above about ten is best, although five is serviceable

The last stage, however, can present special problems. If the projected pressurdrop across the last stage increases, that increase operates largely on the rotoblades, increasing the reaction and developing additional thrust on the rotor anthrust bearing. There may also be a random distribution of energy in the form obuffeting, which can excite blade vibrations even when the turbine is running weaway from resonance conditions. These loads are very difficult to estimate.

Typically, exit axial velocity component should not exceed about 350 feet per second on condensing turbines and 250 feet per second on backpressure turbi

The last stage (or two) are probably limiting on condensing turbines. Backpressunits, on the other hand, may have some margin.

If the last stages must be re-bladed, involve the OEM and select blades that habeen adequately proven by tests or field experience. Larger blades lead to highroot loads and often necessitate replacing the rotor disk as well.

CouplingsCoupling type changes are currently fashionable for several valid reasons. Diaphragms couplings, usually the multiple-membrane type, also known (incor-rectly) as shim pack, are readily available, competitively priced, and well manuftured. The overall dimensions are comparable to the gear-tooth types they typicreplace, making for easy conversion.

Peripheral velocities are comparable between gear tooth types and multi-membtypes. However, windage can cause overheating if no oil spray is used. Pay attetion to anti-windage baffles and ventilation if the surface velocity exceeds abou300 feet per second.

Weights are comparable. Therefore lateral rotordynamics issues are not a probSome types, notably the Zurn Ameriflex, are heavy and may depress the seconical speed slightly. Make sure not to allow critical speeds to encroach on the opating separation margin.

Torsional stiffnesses are considerably different; torsional analysis may need re-fication.



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Change of Steam PropertiesIncreased steam pressure and temperature operate as they do for single stageturbines. However, the design margin over rated conditions is not usually enougmake it an attractive proposition.

583 RotordynamicsAny changes to the speed range or to rotating masses or stiffnesses calls for alook at rotordynamics. This is a job for the specialist. The OEM is a good choicConsultants are readily available, but costs are considerable. CRTC has some bility. Contact the Machinery Group.

Here are some indications of the directional result of certain changes:

• Added mass to rotors reduces first and second critical speeds.

• Added coupling mass depresses second criticals but have little effect on fircritical.

• Increased bearing stiffness, support stiffness, oil viscosity, and dynamic damping all increase critical speeds.

• Coupling-type changes impact the torsional frequencies that are critical fordrives with electric motors, generators, gears, and reciprocating equipment

• Uprates on turbines with lower-half steam chests can cause rotordynamic ibilities.

For a more complete description see the Rotordynamics section of the Compressor Manual.


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