Steam Turbine

8 Steam Turbines 8.1 8.2

Introduction. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Features of Steam Turbines . . . . . . . . . . . . . . . . . . . . . . . . . . .

8-1 8-2

Classification

8.3

Turbine Design and Construction . . . . . . . . . . . . . . . . . . . .

8-3

Bearings • Bearing Housings and Bearing Housing End Seals • Steam Control Valves, Governors, and Control Systems • Turning Gear • Couplings • Additional Tribological Components and Issues • Driven Units

8.4

Lube Oil Systems . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

8-17

Nonpressurized Oil Ring Lubrication • Pressurized Lubrication Systems

8.5

Turbine Oil . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

8-23

Physical Properties • Formulation

8.6

Performance Features of Turbine Oils. . . . . . . . . . . . . . . .

8-26

Viscosity • Oxidation Stability • Freedom from Sludge and Deposits • Corrosion Protection • Water Separability (Demulsibility) • Air Separability and Resistance to Foaming

8.7

Degradation of Turbine Oils in Service . . . . . . . . . . . . . .

8-28

Contamination • Additive Depletion • Thermal and Oxidative Degradation • Biological Deterioration • Turbine Oil Severity

8.8

Lubricant Maintenance . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

8-30

New Oil Makeup • Lube Oil Purification • Refortification

8.9

B.C. Pettinato Elliott Company

Fire-Resistant Fluids . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

8-32

Properties • Degradation • Condition Monitoring • Maintenance

References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

8-34

8.1 Introduction Steam turbines are used extensively in the power generation industry as prime movers for generators. They are also used for mechanical drive application in petrochemical and other industries where they power centrifugal pumps, compressors, blowers, and other machines. In addition, they continue to be used for shipboard propulsion. Sizes range from as low as 0.75 kW for some mechanical drive applications to as high as 1,500 MW for electric generator drives in large nuclear power plants [1]. Steam turbines are particularly well suited for continuous operation, and in many cases are operated for years without shutting down.

8-1

© 2006 by Taylor & Francis Group, LLC

8-2

Handbook of Lubrication and Tribology

8.2 Features of Steam Turbines Steam turbines operate by taking high-pressure steam and converting it into useful mechanical work through expansion. The steam is fed into an inlet casing then throttled through a set of inlet valves, which control the rate of steam admission into the turbine. The steam is then allowed to expand and accelerate through stationary blades or nozzles, which directs the flow onto the rotating blades. The rotating blades convert the steam’s kinetic energy into torque, which results in rotation of the turbine shaft along with a loss of pressure and temperature in the steam. The rotating shaft is used to drive machinery coupled to the exhaust end of the turbine shaft. Absence of lubrication from the steam path is an important feature. Since the exhaust steam is not contaminated with oil vapor, this allows the steam to be condensed and returned directly to the boilers for reheat, or extracted and used for direct heating or other purposes. The lack of internal lubrication also results in a relatively low rate of lubricating oil consumption [2].

8.2.1 Classification Steam turbines have numerous configurations and means of classification. A steam turbine is generally classified as being either high-pressure or low-pressure, condensing or noncondensing, single-stage or multi-stage, single-valve or multi-valve, extraction or nonextraction, direct drive or gear drive, and for either electric generator, mechanical drive, or propulsion service [3]. In addition, steam turbines are classified in accordance with recognized engineering standards, which govern various aspects of turbine design and construction. Some of these classifications are discussed further. High-pressure designs refer to the internal pressure to be contained by the main shell and casing parts. High pressure generally refers to pressures in excess of 13,800 kPa (2,000 psig) where double shell construction is often used. The pressure and temperature of steam are interrelated. Higher inlet steam pressure is often accompanied by higher steam temperature. Temperatures can range from 200◦ C to over 600◦ C. High temperature generally refers to applications with inlet temperatures in excess of 540◦ C (1,000◦ F). Condensing turbines exhaust steam at less than atmospheric pressure, whereas noncondensing (back pressure) turbines exhaust steam at higher than atmospheric pressure. Condensing machines tend to be larger and more complex than noncondensing designs due to the increased volume expansion of the steam at the exhaust end as well as the additional hardware required to drop the exhaust end pressure below atmospheric. In direct drive arrangements, the turbine is directly coupled to the driven machine; whereas gear drive applications have either a speed increasing or speed reducing gear between the turbine and driven equipment. The use of a speed increasing or reducing gear creates added complexity, cost, and power losses along with additional requirements of the lube oil system. However, the use of gears greatly increases the application range whether the need is for high torque as in marine propulsion or high-speed requirements such as integrally geared compressors. Gear drives also enable the efficient use of small turbines, which can operate at higher speeds when a reduction gear is used. Generator drive turbines operate at single speeds to synchronize the generators with the electric grid. Typically, the synchronization speed is either 1,800 or 3,600 rpm in regions with 60 Hz power, or 1,500 or 3,000 rpm in regions with 50 Hz power. On the other hand, mechanical drive turbines are variable speed with shaft speeds as low as 1,000 rpm or as high as 20,000 rpm depending on the turbine and the application. A number of different engineering standards have been developed for the design and procurement of steam turbines as shown in Table 8.1. American Petroleum Institute (API) standards pertain to design, manufacture, and testing of mechanical drive turbines for petrochemical application [4,5]. National Electrical Manufacturers Association (NEMA) standards pertain to design and application of mechanical drive turbines and turbine generator sets for electric utility application [6,7]. Military standards generally apply to steam turbines for shipboard use [8–10]. Other international


Steam Turbines TABLE 8.1

8-3 Steam Turbine Design and Procurement Standards

Standard designation API 611 API 612 / ISO 10437 IEC 60045-1 NEMA SM-23 NEMA SM-24 MIL-T-17286D MIL-T-17600D MIL-T-17523

Standard title General-purpose steam turbines for refinery service Special-purpose steam turbines for refinery service Steam turbines — part 1: specifications Steam turbines for mechanical drive service Land-based steam turbine generator sets 0 to 33,000 kW Turbines and gears, shipboard propulsion, and auxiliary steam; packaging of Turbines, steam, propulsion naval shipboard Turbine, steam, auxiliary (and reduction gear) mechanical drive

recognized standards such as IEC 60045-1 are also used to assist in steam turbine specification and procurement [11]. Figure 8.1 shows a general-purpose (API 611) turbine. These turbines are either horizontal or vertical units used to drive equipment that is usually spared, is relatively small in size (power), or is in noncritical service. General-purpose steam turbines for refinery service are intended for applications where the inlet gauge pressure does not exceed 4,800 kPa (700 psi), the inlet temperature does not exceed 400◦ C (750◦ F), and the speed does not exceed 6,000 rpm [4]. The turbine shown in Figure 8.1 has lubrication consisting of sumps at each journal bearing with oil ring-lubricated bearings. An isolated mechanical–hydraulic governor with oil sump is used to control speed. Figure 8.2 shows a special-purpose turbine for refinery application that meets API 612/ISO 10437 specifications. Such units are usually not spared and are used in uninterrupted continuous operation in critical service. They are not limited by steam conditions, power, or turbine speed. The equipment (including auxiliaries) covered by these standards are designed and constructed for a minimum service life of 20 yr and at least 5 yr of uninterrupted operation [5]. The turbine shown in Figure 8.2 has lubrication provided by a circulating oil system console (not shown) providing oil at high volumes to the bearings and to the servo valve actuator.

8.3 Turbine Design and Construction The parts of a steam turbine may be thought of as being in four groupings (1) the rotor, or spindle, (2) stationary parts, (3) the governing and trip systems and valves, and (4) auxiliary systems consisting of the lubrication system and other components such as the condition monitoring system. The rotor, depending on turbine type, may consist of wheels mounted on a shaft or may be machined from a solid forging or a forging made up of welded sections. In each case, the rotor carries securely fastened radial blades or buckets. Principle stationary parts consist of the steam-tight casing, nozzles, shaft seals, and bearings. Turbine governors control speed by controlling steam-admission valves through mechanical, pneumatic, or hydraulic actuators. Those parts of the turbine requiring lubrication consist of the bearings supporting the rotor, hydraulic actuators and governor components, and the trip system; and in some cases, a turning gear, geared couplings, and front standard. Lubricated parts reside external to the steam path, and when properly isolated will not contaminate the steam or become contaminated by the external environment. The lubrication system may be simple reservoirs in the pedestals of ring-oiled bearings, or elaborate circulation systems, having pumps, coolers, filters, and monitoring devices [12]. Figure 8.3 shows a typical unit of an oil-piping diagram for a turbine, gear, and generator string. Lubricating oil is supplied at two pressures by an oil console (not shown). Lube oil is supplied at low pressure of 100 to 125 kPa (15 to 18 psig) to the bearings. High-pressure oil of 1,000 kPa (150 psig) is supplied to the trip and throttle valve, to the valve actuator, and, if needed, to the governor mechanism. Bearing and coupling housings are part of the lube oil circuit and act to return oil to the reservoir.


8-4

Sentinel valve

Rotor disk assembly

Casing cover

Shaft sleeve seal

Exhaust end sealing gland Carbon ring assembly

Steam end sealing gland

Rotor locating bearing Overspeed thip assembly Coupling (governor drive)

Carbon ring assembly

Governor Governor linkage

Oil rings

Oil rings

Rotor shaft

Steam end bearing housing

Shaft sleeve seal Steam chest Governor valve

Exhaust end bearing pedestal Exhaust end casing

FIGURE 8.1

Reversing blade assembly

Steam end casing Nozzle ring

Steam end journal bearing

Steam end support

General-purpose steam turbine. (From Installation, Operation, and Maintenance Instructions for YR Turbines, Elliott Company, Jeannette, PA, 2003. With permission.)



Exhaust end journal bearing

Steam Turbines

Rocker arm bearing

Lubrican connection

Governor linkage assembly Valve stem and packing Valves, seat, and bar assembly Breather cap Bearing Journal housing bearing Bearing housing deflector Shaft end with coupling bolt pattern

Steam chest

Turbine case

Gland packing assembly case

Exhaust end packing gland assembly

Breather cap Interstage shaft seals

Journal bearing Thrust bearings

Steam end packing gland assembly

Rotor

Bearing housing end seal

Oil drain Steam exhaust

Bearing housing end seals


Gland seal leak off

8-5

Steam turbine for special-purpose refinery service.

Bearing housing

Steam end flexible support Casing drain

FIGURE 8.2

Oil drain

8-6

Lube oil supply to unit conn’s

P1

Flexible hose

PSH PSLL

S P1

Breather Servo motor Driven equipment

IP

Coupling

Breather Gear

To reservoir

FIGURE 8.3

SG

SG

1/2” per foot minimum

SG

Concentric reducer I

Inlet SY servo motor/ valves

Trip and throttle valve S

T1

SG

e

Slop

Unit oil piping diagram for turbine-gear-generator set.


Orifice PSLL

Thermowell Thermowell

P

Current/pneumatic convertor

PCV

Pressure control valve

P1

Pressure indicator

PSH

Pressure switch, high

PSLL

Pressure switch, trip

SG

Site glass

SY

Speed relay

T1

Temperature indicator

T1

SG


SG

T1

2 valve manifold instrument valve/bleed valve

High pressure control oil supply

Trip solenoid dump valve

Spare vent

Spare vent

Solenoid valve (2-way) Ball valve

Nitrogen precharge

Turbine

Coupling

SG

Governor control signal

N2

Control oil accumulator

Steam Turbines

FIGURE 8.4

8-7

Tilt pad journal bearing. (From Elliott Company. With Permission.)

8.3.1 Bearings Proper rotor position is maintained by journal and thrust bearings. Journal bearings are used in pairs for radial positioning of the turbine shaft supporting the gravitational load of the shaft. Thrust bearings are used for axial positioning and support thrust loads that arise from steam forces within the turbine case. Thrust bearings are located at the steam end of the turbine opposite the coupling, and are used in pairs to accept thrust loading in either direction along the axis of the rotor. Steam turbine bearings can be either hydrodynamic, rolling element, or magnetic. Hydrodynamic bearings are the most prevalent. 8.3.1.1 Hydrodynamic Bearings Hydrodynamic bearings are highly advantageous because they suffer little or no wear and have exceptionally long life thereby enabling long periods of continuous operation, often in excess of 5 yr. In addition, the bearings possess dynamic characteristics that allow for vibration control thereby enabling high-speed operation, and traverse of rotor critical speeds. For this reason, hydrodynamic bearings are the most common type of bearing applied to steam turbines. Journal bearings are most often of the plain cylindrical, elliptical, multilobed, pressure dam, or tilt pad design. Figure 8.4 shows a schematic of a tilt pad journal bearing. Tilt pad journal bearing designs consist of several pads arranged in a ring around the shaft with the pads free to tilt about their respective pivots. Tilt pad journal bearings may include several design variations such as self-aligning features to compensate for misalignment, and special oil feed and drain configurations for temperature and power loss control [13,14]. One particular advantage of tilt pad journal bearings is their dynamic characteristics and inherent resistance to rotordynamic instability, which allows for control of vibration even at high speeds. Thrust bearings are usually of the tapered land or tilt pad design. Figure 8.5 shows a six shoe selfequalizing tilt pad thrust bearing. Tilt pad thrust bearings may also have features to compensate for misalignment, as well as special oil feed and drain configurations for temperature and power loss control [15,16]. Hydrodynamic bearings are lubricated with turbine grade oil either by a low-pressure circulating supply system or by ring lubrication where appropriate. In low-pressure supply systems, the oil flow is metered to each bearing by an orifice or other flow-controlling device. The oil flows into the clearance spacing of the bearing where it forms a wedge separating the bearing and shaft surfaces. The oil exits axially out the sides of journal bearings; and exits radially and tangentially from thrust bearings. Observation of drain oil flow through sight boxes can be taken as an indication of at least partial flow through the bearing and is often used as a quick indication that the oil pump is running and that the oil supply is probably sufficient. Oil supplied to the bearings functions as both a lubricant and as a coolant to


8-8


FIGURE 8.5 Self-equalizing tilt pad thrust bearing. (From A General Guide to the Principles, Operation and Troubleshooting of Hydronamic Bearings, Publication HB, Kingsbury, Inc., Philadelphia, PA, 1997. With Permission.)

counteract the heat generated by shearing of the oil during operation and conduction from the hot rotor. Hydrodynamic bearings are limited with respect to minimum film thickness, maximum bearing temperature, and peak oil film pressure. These restrictions are inherently related to the load, speed of operation, and design of the bearing [17]. The bearings may be boundary lubricated during startup and turning gear operation, developing a full film shortly after startup. Operational film thickness is typically 25 to 75 µm (0.001 to 0.003 in.). Bearing metal temperature at the instrumented location may range from less than 55◦ C (130◦ F) for an unloaded inactive thrust bearing up to 130◦ C (265◦ F) for a bearing operating near its design limits. Peak oil film pressure is typically 2.5 to 3 times the specific load defined as P=

W A

(8.1)

such that P is the specific load (N/mm2 ), W is the load (N), and A is the projected area (mm2 ) [17]. For journal bearings, the area is the product of the diameter and length. For thrust bearings, the area is the area of the loaded surface. Bearing surfaces consist of a soft metal bonded to a hard metal backing. For North American operation, the soft metal surface is most often an ASTM B23 grade 2 babbitt comprised of 89% tin alloyed with antimony, lead, copper, iron, and trace amounts of other metals. Equivalent specifications can be found in ISO 4381 as SnSb8Cu4 [18], and Federal Spec Q-T-390 Grade 2. In some cases, an ASTM B23 grade 3 babbitt is used. Babbitt bearing surfaces generally cause the least damage to steel shafts when operated with inadequate lubrication or with contaminants. Babbitts are good for embedding hard contaminant particles and for resistance to seizure and galling [19]. In addition, tin-based babbitt is highly resistant to corrosion from organic acids and can provide satisfactory operation in the presence of oxidized and contaminated oils. A disadvantage of babbitt bearing materials is their relatively low compressive, tensile, and fatigue strengths especially at high temperature. To provide additional strength, the babbitt surface is cast and bonded as a thin layer to a hard metal backing, which may be steel, bronze, or chromium copper. Steel is the most prevalent and least expensive backing material. Chromium copper is used for its superior thermal conductivity enabling reduced bearing metal temperature. A good babbitt bond is critical, and can be inspected by nondestructive ultrasonic testing as described in ISO 4386 Part 1 [20]. The journal or thrust collar/disk is usually polished steel with surface finishes not exceeding 0.8 µm (32 µin.) Ra . The rotating element is either an integral part of the turbine shaft or else attached mechanically to the shaft. Bearing surface materials are normally steel containing less than 2.5% Cr, to prevent


Steam Turbines

8-9

a type of failure known as wire wool [20]. In those cases where 12-Cr shaft material is required due to erosion concerns, a sleeve may be used as the bearing journal or the rotor may be inlaid with an acceptable steel material. To supervise the satisfactory and safe operation of turbine bearings, one or more of the following quantities may be monitored: inlet and drain temperature of oil, bearing metal temperature, drain flow of oil, shaft position, vibration amplitude, oil film pressure, and lift oil pressure. Bearing metal temperatures are measured using temperature sensors embedded in the bearing backing metal near the babbitt bond line and bearing surface [21]. High bearing metal temperature can be indicative of potential bearing failure. Bearing metal temperature that rises in an upward trend without corresponding change to load or speed is also indicative of potential bearing failure. There is varying opinion with respect to metal temperature limitation. In general, the manufacturer’s recommendation should be followed especially for new equipment lacking in historical data. Drain temperatures are also useful for identifying problems. Drain temperature of oil is an indicator of bearing power loss if measured separately for each bearing. It is also an indirect indication of bearing health, but not as reliable as bearing metal temperature. For this reason, drain oil temperature is not relied upon as an indication of safe operation unless the bearings are not instrumented. Radial and axial shaft position and vibration are measured with noncontacting eddy-current probes. Drastic shaft movement is an indication of bearing distress that occurs with wiping of babbitted surfaces [22]. Excessively high radial vibration is another sign of potential bearing distress and needs to be monitored as it may cause babbitt surfaces to fatigue or internal rubs to occur. Depending on the bearings, hydrodynamic bearing maintenance can consist of inspection, repair, or replacement. After installation, a lift check should be performed on each journal bearing, and the thrust bearing endplay and rotor axial position should be checked and recorded. These same checks should also be performed prior to bearing removal especially if abnormal bearing conditions were observed prior to shutdown such as high metal temperature or vibration. During visual inspection, the bearings are examined for signs of wear or distress such as scoring, cracks, pivot fretting or brinelling, heat discoloration, electrostatic discharge machining, corrosion, flaking, signs of overheated or contaminated oil such as varnish deposits, and loss of babbitt bond. The rotor journal and thrust areas also need to be examined for signs of distress such as scoring. Causes of bearing distress and failure include overloading, insufficient oil flow, insufficient bearing clearance or endplay, excessive overspeed, excessive vibration, and too high inlet oil temperature. Corrosion failures for tin babbitt bearings are fairly uncommon, but can occur in certain cases. The formation of hard deposits of tin oxide on tin rich white metal has been a problem with bearings in steam turbines caused by electrolytic action in certain environments such as when the oil contains free water with salt in solution [20]. Oil contamination from process gases that originate from the seal oil systems of driven units such as compressors can be particularly corrosive and may attack the components found in babbitt. 8.3.1.1.1 Hydrostatic Jacking Hydrodynamic bearings may include additional features such as an externally pressurized hydrostatic jacking system. The purpose of hydrostatic jacking is first to reduce the required breakaway torque during either start-up or turning gear operation and second to reduce bearing wear during turning gear operation. Hydrostatic jacking is effective by simply reducing the loading on the bearing surface such that it is within acceptable ranges. One manufacturer recommends consideration of hydrostatic jacking when the specific load on startup exceeds 1,300 kPa (190 psi) for plain journal bearings, 1,200 kPa (175 psi) for tilt pad journal bearings, and 60% of the maximum load for thrust bearings [23–25]. The need for hydrostatic jacking depends on the frequency of start-ups, duration of any baring condition, and available starting torque. Hydrostatic jacking systems are typically designed to lift the rotor off the bearings; however, this is not always practical. Hydrostatic jacking is effective so long as the friction torque is acceptable, the loading on the babbitt surface is reduced, and associated wear is negligible. Bearings with hydrostatic lift features require a high-pressure oil system, which typically supplies oil


8-10


at 7,000 to 14,000 kPa (1,000 to 2,000 psig). The high-pressure oil is turned off shortly after start-up and turned on during coast down. Lift oil pressure may be indicated by reading the pressure supplied to the lift pocket. 8.3.1.2 Rolling Element Bearings Rolling element bearings (also known as antifriction bearings) are used where service is not critical or the steam turbine is spared. These bearings can be used as a complete set to accommodate both radial and axial loads or are used as a thrust bearing in conjunction with ring-lubricated bearings. Rolling element bearings are generally less reliable than pressure fed hydrodynamic bearings and are only applied when they meet specific criteria with respect to their speed and life, which are designated by dN and L10 parameters, respectively. The dN parameter is the product of d, the journal diameter (mm) and N , the rated speed in revolutions per minute. Operation of dN in excess of 300,000 generally requires oil lubrication. The L10 parameter describes the basic rating life expressed in number of operating hours, or millions of revolutions with 90% reliability. The latest revised and updated L10 equation considers the bearing design, dynamic load, reliability factor, and life adjustment factor that involves the complex interaction of lubrication conditions, contamination, bearing material properties and other factors [26]. Rolling element bearings are generally designed and retained in accordance with American Bearing Manufacturers Association (ABMA) standards. The Conrad type or deep groove ball bearing is a typical design. The bearings are lubricated either by grease with protection against overgreasing, or by oil supplied by bath, mist, or jet lubrication. Grease fittings are required to extend outside the machine to permit regreasing during operation. Venting is provided to prevent pressure buildup in the housing. One particular disadvantage of rolling element bearings is that they cannot be horizontally split without reducing their life and degrading their performance. As a result, most rolling element bearings cannot be replaced without removing the rotor and coupling. Presence of water in oil is particularly detrimental to the life of a rolling element bearing [27].

8.3.2 Bearing Housings and Bearing Housing End Seals Bearing housings support and position the bearings such that the rotor is centered in its respective packing bores. These housings are also used to mount vibration monitoring and other condition monitoring devices. The steam end bearing housing further encases the overspeed trip assembly; as well as the governor speed sensor, which may consist of a notched wheel and speed pickup, or it may consist of flyweights or other devices. At times, a turning gear is also present. Grounding brushes may be mounted to the outboard end of the bearing housing to prevent the buildup of high voltage between the shaft and the case, which can damage the bearings through electrostatic discharge. Bearing housings also function as a part of the lube oil circuit, keeping oil in while keeping contaminants, such as steam out. In the case of pressure-lubricated hydrodynamic bearings, the housings are arranged to minimize foaming through proper design of the drain and vent system to maintain oil and foam levels below shaft end seals. Proper sizing of drains is important to minimize foaming. Bearing housings are equipped with replaceable labyrinth end seals and deflectors where the shaft passes through the housing to minimize contamination and leakage. Bearing housings and gland seals are spaced to help prevent leaking oil from entering the glands and gland steam from entering the bearings. For ring-oiled bearings, the housings further act as oil sumps and may contain water jackets for cooling the oil. Bearings housing oil seals may suffer from oil carburization, contaminant leakage into the seal or oil leakage from the seal. Contaminant leakage into the bearing housing can be a problem when using a vapor extractor on the main oil tank, which creates a slight vacuum in the bearing housings through the oil drain lines. Pressurizing the annulus in the oil seal with a gas purge such as nitrogen or air can assist with


Steam Turbines

8-11

seal leakage. This may also cool the oil seal to prevent carburization. Overheating of the oil seal may also be prevented by improvements to the heat guards [28].

8.3.3 Steam Control Valves, Governors, and Control Systems As shown in Figure 8.6, steam is directed into the turbine steam chest through either a trip or trip and throttle (T&T) valve. Trip valves are opened by fluid pressure and mechanically closed by spring force. The trip system is controlled by an overspeed governor. A trip due to overspeed or other unsafe operating condition causes the solenoid valve to open thereby causing system depressurization and immediate closure of the trip valve, which shuts off the steam thereby bringing the turbine to an eventual stop. After passing through the trip valve, the steam is directed through the steam chest, and then through control valves (also called governor valves). The control valves throttle the steam into a nozzle ring matching the turbine power to the load thereby controlling speed. The control valves may be operated by mechanical linkage, by bar-lift arrangement (Figure 8.6), by cams, or by individual hydraulic cylinders. Mechanical and pneumatic actuators can be found on the smallest turbines whereas hydraulic actuators are required on most other units. Extraction turbines have additional valves located at an intermediate stage in the turbine. Extraction valves may be of poppet or spool type for higher pressure, or of grid type for controlling large volumes of steam at lower pressure. In each case, the valve actuators are controlled by the main governor. The main governor operates independently from the overspeed governor. The main governor can be either a relatively simple system that acts directly upon a steam-admission valve; or a complex system that may control speed, extracted steam, and devices separate from the turbine such as a compressor or the boiler. Figure 8.7 shows a mechanical–hydraulic governor with hydraulic actuator. In this case, hydraulic accumulators are used to supply the high volume of fluid that is required for rapid control action during sudden changes in load. To achieve the high force levels required in multivalve applications, the governor typically controls a servo (prepilot or slave) to a master pilot that controls the flow of high pressure oil to a large piston as shown in Figure 8.7. The assembly of servo, pilot valve, and piston is called a servomotor. In such a control system, a few ounces in governor force can be multiplied through a hydraulic mechanical advantage to generate the thousands of pounds of force that may be required to operate the turbine governor valves [29]. Required hydraulic oil pressures typically range from 350 kPa (50 psi) on small turbines to 18,000 kPa (2,600 psi) on very large turbines [30,31]. Turbine oils are typically employed at pressures below 2,000 kPa (290 psi) whereas fire-resistant fluids are often used at pressures exceeding 2,000 kPa and in installations where steam pipe temperatures exceed the auto-ignition temperature of turbine oil, particularly in power plant applications [30]. The governor and actuator control system may be supplied from the same lube system as the bearings or may be fed independently from a separate system. In small turbines, hydraulic and mechanical–hydraulic governors are often self-contained units featuring a shaft driven oil pump, and an oil sump with sight glass for determining the oil level. In medium sized turbines, typical of process industry applications, the control system is normally fed off the same lube system as the bearings. In large power plant turbines, two separate circulating systems are usually employed: one for the bearings using turbine oil and one for the control system using a fire-resistant phosphate ester fluid. Control systems have long had high visibility due to reliability and maintenance shortcomings. Large quantities of mechanical components such as pins, links, levers, rod end bearings, hydraulic relays, springs, gearing, and flyball governor assemblies are present and subject to wear. The use of electronic speed sensors and electronic governor controls has enabled the elimination of some wearing mechanical parts, and has improved control and flexibility through use of noncontacting pick-ups and nonmechanical feedback circuits. Actuators have remained primarily hydraulic due to the large forces and quick response time required. Governor maintenance depends considerably on the type of governor in use, and the manufacturer’s recommendations should be followed. The proper oil must be selected, and it must be kept clean, dry, and at


8-12

High-pressure oil supply from oil console Variable-pressure control oil Drain oil to oil console reservoir

Steam to turbine

Trip and throttle valve

Inlet steam valves

Trip pin

Spring-loaded handle Solenoid valve

Trip lever Knife-edge

Bearing housing Servo motor

Orifice

High-pressure oil from oil console

FIGURE 8.6

To oil console drain

Trip system. (From Elliott Multivalve Turbines, Bulletin H-37B, Elliott Company, Jeannette, PA, 1981. With permission.)



Electrical leads

Steam chest

Variable-pressure control oil

Steam Turbines

KEY High-pressure oil supply from oil console

Drain oil to oil console reservoir Governor internal high-pressure oil supply

Steam inlet

Governor intermediate pressure oil

Woodward PG governor

Trapped governor oil Pilot valve

Inlet steam valves

Drain oil to governor internal reservoir

Flyweights Pre-pilot valve

Accumulators

Oil pump Governor drain Governor drain

Lube system drain

High pressure oil from lube system

Mechanical–hydraulic governor system. (From Elliott Multivalve Turbines, Bulletin H-37B, Elliott Company, Jeannette, PA, 1981. With permission.)


8-13

FIGURE 8.7

Worm and wheel governor drive Inlet servo-motor

8-14


the proper level and temperature. Water contamination, even in trace amounts, contributes significantly to early failure as well as forming oxides that also contribute to failures. As noted by one governor manufacturer, dirty oil causes most governor/actuator troubles [32]. Oil contamination and degradation is particularly problematic on self-contained sump units that do not have oil conditioners. Other parts of the governing system must also be maintained. Governor valves and linkages must be free from binding or sticking. Valve sticking may be related to steam impurity, which may lead to deposits on the valve stem [33], or it may be related to oil contamination causing deposits or corrosion in tight clearance hydraulic components [34], which may have clearance as small as 5 µm. Hydraulic actuators may operate on the valves through a linkage. Loose or worn linkage components can cause unacceptable governor control. Linkage bearings are usually hand-oiled or greased. Some, however, are made from low-friction materials, which may require little or no lubrication.

8.3.4 Turning Gear High-temperature steam turbines are sometimes equipped with a turning gear to prevent bowing of the rotor when at rest, especially after shutdown. The need for a turning gear depends upon the probability of rotor bow, which is related to the steam temperature, shaft diameter, and bearing span. Turning gears are primarily found on large turbines with long bearing spans, though they are sometimes needed for small turbines as well to allow for oil circulation through the bearings during cool down. The turning gear is operated prior to turbine run-up and immediately after shutdown. Turning gears are electric motor driven with a means for disengagement such as a clutch or retractable gear. The turning gear motor is typically grease lubricated whereas the actual turning gear and bearings are lubricated with oil supplied from the main circulation system. A separate, relatively small, motor-driven oil pump is generally provided to supply oil to the bearings of the turning gear system. The auxiliary oil pump, which backs up the main oil pump, may also be used for this service. During turning gear operation, oil inlet temperature may be kept cool to increase oil viscosity thereby maintaining a thick oil film in the turbine bearings during low-speed operation.

8.3.5 Couplings Couplings are used to connect the steam turbine to the driven equipment. They are made from corrosionresistant or coated materials. Couplings can be either rigid or flexible. Rigid couplings are essentially two flanges bolted together. Such couplings require no lubrication, but do not readily accommodate changes to machine position, which can be caused by thermal expansion of the equipment, foundation settling, and strain due to loading. Flexible couplings accommodate some misalignment; however, their use does not preclude the need for proper machine alignment of both the turbine and driven equipment [6]. Flexible couplings are described by a number of standards such as API 671, ISO 10441, and MIL-C-23233A. For turbine applications, special attention may be required with respect to machinery alignment due to thermal expansion. Quill shafts, membrane couplings, and contoured disc couplings run dry and without lubrication and are often preferred for their low maintenance. Gear couplings must be lubricated. 8.3.5.1 Gear Couplings Gear couplings can be advantageous because of their light weight and minimal required overhang, and because they allow for maximum axial movement between turbine and driven equipment shaft ends as caused by expansion of various parts under hot conditions [35]. In general, however, the need for lubrication and maintenance means that geared couplings are seldom used in new turbine applications though there is still a considerable population of geared couplings that must be maintained. The life of a geared coupling is primarily dependent on alignment and lubrication. The majority of geared tooth coupling failures are due to improper or insufficient lubrication [36]. Gear coupling lubrication is complicated by the centrifugal effect that a spinning coupling has on lubricants. Packed lubrication with grease can only be applied at relatively low speeds since the thickener tends to separate out of the grease under high


Steam Turbines

8-15

centrifugal force [37]. Packed lubrication with either grease or oil may require lubricant replenishment at 6 to 12 month intervals [38]. Typical grease used for high-speed geared coupling is NLGI #1 or #2 grade with R&O inhibitors. Greases that are specifically formulated for high-speed coupling application use thickeners, which have a density closer to that of oil [39]. These formulations resist separation due to centrifugal effects. Special grease formulations can also extend replenishment intervals beyond the 6 to 12 months typically cited. A test method for evaluating grease separation is ASTM D4425. In this test, the grease is subjected to 36,000 G centrifugal acceleration at 50◦ C for a period of at least 6 h. Results from the test are presented as K 36 = V /H

(8.2)

where V is the oil separation in volume percent, and H is the accumulated time of testing in hours. High speeds and low maintenance in gear couplings require the use of continuous lubricant feed at each hub that is provided by either oil spray or jet using filtered oil piped from the system bearing oil supply. Coupling teeth for such applications are often hardened usually by nitriding [39]. Lubricants must be carefully selected with additives that resist separation from centrifugal force. Oil additives, in particular silicone antifoam compounds, can separate out of the lubricating oil and form sludge [40]. The coupling lubricant must also resist reaction with metal particles that may exist in the coupling due to wear [41]. Such measures are unnecessary when using dry couplings. The coupling housings provide safe enclosure of the coupling. Coupling housings may also act as part of the lube oil system. The housings are oil tight and include provision for coupling lube oil supply if needed and drainage back to the reservoir. The drains also handle any oil that may be carried over from the coupled equipment and are consequently featured on housings for both dry and lubricated couplings. A filter breather is attached to the coupling housing to allow proper drainage or the housing is connected to the bearing oil vent system of the equipment train. Regardless of the type of coupling used, proper design of coupling housings is important due to windage losses, heat generation, and potential for oil leakage from the joined equipment [42].

8.3.6 Additional Tribological Components and Issues Several components outside of the lubricating oil circuit require batch lubrication and special material consideration to limit wear and corrosion. Among these components is the turbine casing and steam patch components. Gland seals may also be subject to wear and have considerable effect upon water contamination of the lubrication system. 8.3.6.1 Casings Steam casings expand and contract due to changes in casing temperature caused by the use of high temperature steam. Thermal movement is typically accommodated at the steam end by either a flexible support or sliding pedestals. Sliding pedestals are most common on large turbines and rarely used on small and medium sized units. Sliding pedestals may operate dry, or they may be lubricated by either grease or oil depending on the load, temperature, and expansion. The use of lubrication reduces friction thereby allowing casing thermal expansion without binding. Binding of the casing can cause distortion, misalignment, and vibration. Grease lubricated casing supports are often used for large central station steam turbines. Grease may be supplied by either a common system or grease gun. The type and application should follow the manufacturer’s recommendation. NLGI number 1 or 2 grease of sodium, lithium, or sodium–calcium soap base have been used for lubricating sliding pedestals; in addition a mixture of graphite and cylinder or turbine oil mixed to a paste consistency has also been used [43]. Problems associated with grease separation have been noted on high temperature, heavy turbines used in central station applications due to the high temperature and heavy loads associated with these applications [44].


8-16


8.3.6.2 Steam Path Parts that are located in the steam path consist of the steam chest, rotating blades, stationary nozzles, diaphragms’, seals, valves, and valve guides. These parts have tribological issues that must be solved without lubrication, which would contaminate the steam. Stationary nozzles and rotating blades are damaged by erosion and corrosion mechanisms. Turbine blades in particular suffer the most damage in turbines resulting in loss of power, efficiency, or operation. There exist many causes for wear through the steam path; however, three mechanisms are prevalent. These are (1) moisture impingement erosion, (2) erosion–corrosion, and (3) solid particle erosion. Moisture impingement erosion is caused by the presence of water droplets in the steam. When turbine blades operate in wet steam, the moist steam may cause blade erosion. Erosion is dependent on speed of the rotating blade, wetness of the steam, and blade design. Blade velocities can exceed 250 m/sec (825 ft/sec) at the tip. Moisture impingement erosion has been noted to be particularly problematic in the final stage of long multi-stage low-pressure turbines due to condensation. Allowable wetness is related to steam conditions, blade velocities, and design. In some cases, there is no need for a moisture limit. In other cases, 8% [45], 12% [46], or other moisture limit is used depending on the application. Moisture erosion also effects seals and can lead to degradation of performance and changes to the thrust loading [47]. Erosion–corrosion problems are caused by reactive steam chemistry. Steam of insufficient purity may cause deposits on the casing, nozzles, blades, seals, and sealing surfaces. These deposits may contain corrosive agents such as chlorine, which can attack the material used on these components. This results in eventual pitting and stress corrosion cracking [48]. Geothermal steam applications are known to have particularly corrosive steam with constituents of silica, sodium, ammonia, calcium, and sulfate. The acidity of geothermal waters can be very high with pH as low as 1.8 [49]. Solid particle erosion (SPE) is caused by entrainment of erosive materials in the steam. Solid particle erosion is traceable to exfoliated material coming from the boiler tubes, and in some cases, the steam leads. This type of erosion appears to be related to both the size of the unit and the pressure being employed. The solid particle erosive mechanism is most prevalent on large central station utility turbines; and it is rarely observed on small turbines operating under 540◦ C (1000◦ F) [50]. An important factor in each of these erosive wear mechanisms is the condition of the steam. For this reason, steam conditioning may be used for ensuring reliable operation. Monitors have been developed to quantify the particle loadings from the boiler [46]. Various separators and moisture removal devices may be employed upstream and inside of the turbine. Strainers are used to remove the largest particles and trap foreign objects. Some recommendations for steam purity are specified by NEMA for lowpressure turbines relating to the amount of dissolved solids, alkalinity, conductivity, and content of silicon oxide, iron, copper, sodium, and potassium [6]. Original equipment manufacturers also provide steam purity recommendations. Design methods for combating erosive wear include the use of either hardened materials or hard coatings such as Stellite on turbine blades. Stainless steels such as 12% chromium steel are also used. For turbine nozzles and internals, chromium steel cladding may be used [50]. Corrosion-resistant coatings have also been developed for this service. Other forms of wear and degradation internal to the steam path also exist. Turbines that are not in operation can experience a form of corrosion known as stand-by corrosion [45,49]. The corrosion is due to steam leaking into the turbine past a valve, which is not tight. Once the steam has leaked into the unit, it can condense and corrode the unit. This type of corrosion may cause severe pitting on stainless steel buckets. Brown specs (known as tubercles or scabs) form on carbon steel parts, such as discs and diaphragms [46]. It is therefore important that idle turbines have the inlet valve tightly seated and that all the casing drains be open [49]. Additional measures such as an additional drain between the turbine and steam inlet valve, and blanketing turbine internals with a positive flow of dry gas along with running the lubrication system and rotating the journals have also been performed [43]. 8.3.6.3 Seals and Gland System In order to maintain efficiency and performance, seals are required to limit steam leakage from the turbine case and between each stage. Casing end seals, also referred to as packing seals, are provided where the


Steam Turbines

8-17

shaft ends pass through the casings. They are used to seal against the leakage of steam to atmosphere, and to seal against air suction into the low-pressure condensing case of condensing turbines. The seals may be carbon ring, labyrinth, or noncontacting mechanical design. Properly functioning gland seals are important for maintaining turbine performance. Improperly functioning gland seals can cause excessive steam leakage, which may ingress into the bearing housing leading to oil contamination. Likewise, a loss of vacuum on condensing turbines can also cause excessive steam leakage past the gland seals. Gland leakage does not have to be visible to cause a problem.

8.3.7 Driven Units Driven units such as gears, compressors, and generators create additional complexity to a lube oil system. Many of these units will have similar requirements to the turbine with respect to journal and thrust bearing lubrication. The oil selected for a common lube oil system must be suitable to all the pieces of equipment to be supplied. Low-viscosity rust and oxidation inhibited (R&O) oils, commonly called turbine oils, are used in many high-speed gear units where the gear tooth loads are relatively low [51] and the high entraining velocity of the gear develops thick elasto-hydrodynamic (EHD) oil films. Slower speed gears, as used for propulsion, tend to be more heavily loaded. These gears generally require higher viscosity lubricants with antiscuff additives [51]. High-pressure oil seals, as used in compressors; and hydrogen seals, as used in generators, can cause contamination of the seal oil by gas such that natural or vacuum degassing is required [31]. In some cases, a separate, isolated, lube oil system is used to provide seal oil due to potential contamination of the lubricating oil [52].

8.4 Lube Oil Systems Lube oil systems may be classified as either nonpressurized or pressurized systems. Nonpressurized lube systems consist of ring lubrication and are common on very small steam turbines. Larger turbines use pressurized lubrication.

8.4.1 Nonpressurized Oil Ring Lubrication Ring-lubricated hydrodynamic bearings are used where service is not critical or the steam turbine is spared. These bearings have the advantage of not requiring an external lube oil system thereby enabling steam turbine application where initial cost is a primary concern. Figure 8.8 shows an oil ring-lubricated journal bearing. The oil ring lubrication system employs metal rings to deliver oil to the turbine bearings. The rings are rotated by the journals carrying oil from a sump below the bearings to the top half bearing liners where it is fed into the clearance between the bearing liners and the shaft journals. Oil is drained from the ends of each bearing liner and returned to the bearing housing reservoirs to be cooled. Some of the supplied oil may be used to feed a rolling element bearing that is normally required in conjunction with ring-lubricated journal bearings for thrust positioning. The use of ring-lubricated bearings is limited with respect to load capacity, journal rotational speed, and by the need for cooling. Bearing housings may be double walled to allow water circulation to remove heat from the oil bath. Under conditions of high inlet steam temperature, the bearings can be damaged after shutdown because there is no longer oil circulation to carry heat away from the shaft, and a turning gear is sometimes used to continue the rotation of the shaft and subsequent oil ring lubrication. Ring-lubricated bearing housings are equipped with constant-level sight-feed oilers that maintain a constant reservoir oil level. A permanent indication of the proper oil level is clearly marked on the outside of the bearing housing. Low oil level in the housing will cause inadequate bearing lubrication. Excessively high oil levels can also be detrimental as it may restrict oil ring rotation also causing inadequate bearing lubrication. Housings for ring-lubricated bearings are provided with plugged ports positioned to allow visual inspection of the oil rings while the turbine is running.


8-18

Handbook of Lubrication and Tribology Inspection plug Rotor shaft

Oil ring

Journal bearing

Oiler

Oil reservoir Cooling water Cooling chamber

Lubricating oil Cooling water

FIGURE 8.8 Ring-lubricated journal bearings. (From Installation, Operation, and Maintenance Instructions for YR Turbines, Elliott Company, Jeannette, PA, 2003. With permission.) TABLE 8.2

Lube System Design and Procurement Standards

Standard designation API 614 ASTM D4248 ASTM D6439

Standard title Lubrication, shaft-sealing, and control-oil systems and auxiliaries for petroleum, chemical, and gas industry services Design of steam turbine generator oil systems Standard guide for cleaning, flushing, and purification of steam, gas, and hydroelectric turbine lubrication systems

8.4.1.1 Ring Lubrication System Maintenance Proper oil level should be maintained at all times. Since oil ring lubrication systems have no means of filtering solids from the oil or removing water, periodic sampling and frequent oil changes are necessary to ensure a clean oil supply. The range of cooling water temperature must also be controlled to ensure good heat transfer without promoting condensation in the oil sump. To avoid condensation, the minimum inlet water temperature to the bearing housings should preferably be above the ambient air temperature.

8.4.2 Pressurized Lubrication Systems The pressurized lubrication system is essentially a closed loop system designed to provide an uninterrupted supply of cooled and filtered oil at the proper pressure to the bearings, control-oil system, shutdown system, and other components such as continuously lubricated couplings, as well as gears and seals on adjoining equipment. Oil consoles vary widely depending on the make, size, type, and purpose of the turbine and its adjoining equipment. Lubrication systems are designed according to the application, which may require the use of design standards such as those shown in Table 8.2. Proper lube system design is vital to machine reliability. A typical system is shown in Figure 8.9 and Figure 8.10. The oil is taken from the reservoir and passed through a cooler then filtered. The flow is split into two legs. One leg delivers high-pressure oil to a


Steam Turbines

8-19

FIGURE 8.9 API 614 lube oil console. (From Elliott Company, Jeannette, PA. With permission.)

common header for the governing and control mechanisms. A second leg delivers reduced pressure oil to a common header for the bearings. Bearing oil pressure is typically 100 to 125 kPa (15 to 18 psi), but may range from as low as 55 kPa (8 psi) in some systems to 345 kPa (50 psi) in others. Oil from the bearings and governor mechanisms will drain back to the reservoir. Figure 8.10 shows the following major components: • • • • • •

Oil reservoir Pumps and drivers Filters Coolers Control valves Piping

Additional accessories may include relief valves, transfer valves, accumulators, and instrumentation as shown in Figure 8.10. Not shown is the oil conditioning hardware. Each of the major components is described briefly along with instrumentation, commissioning, and system maintenance. 8.4.2.1 Oil Reservoir The reservoir is usually of rectangular shape, carbon steel construction with an interior coating of rust proofing paint. Solid stainless steel or stainless steel clad construction is also used. Normally the reservoir will have a sloping bottom to drain, clean out manways, gasketed openings, fill opening with strainer, oil level sight gage, and vent with weatherproof breather. The various oil levels as defined in the reservoir are shown in Figure 8.10. Depending on design requirements, the reservoir is sized to contain an amount of oil for anywhere from 3 to 5 min working capacity as measured from the minimum operating level. Large reservoir capacity enables disengagement of entrained air or gas and the settling of water and solid contaminants. A high and low oil-level indicator and alarm are usually provided. A free oil surface in the reservoir of at least 0.37 m2 /lps (0.25 ft2 /gpm) of oil is required to enhance air disengagement from the oil [53]. In addition, the oil reservoir is designed with a sloping bottom (1 unit in 24) such that supplementary water and dirt can accumulate at the area of the low point drain, and thus be drawn off during operation.


Lube oil supply

8-20

High pressure control oil supply To oil reservoir To oil reservoir

Cooling water out

Oil return from units

A

A

PCV

Fill line

Cooling water in

PCV

TE

TI

Oil coolers VENT

Cooling water out

Vent

PCV

Main oil filters

TI PI

e

Purge

RV

RV

Minimum oil level and alarm level

PI

PI

TI

LS

LI Reservoir

Heater

Transfer valve 6-way

Relief/safety valve

Check valve

Pump suction strainer

Globe valve

Primary oil pump and driver

Change capacity (oil required for initial system fill) 800 U.S. gallons

Secondary oil pump and driver

Electric motor driver AC or DC power

Cooling fan Oil filter

Pump suction level

Inside bottom of reservoir

LIT

Tank drain TE

(3 min. retention)

Level indicator

PSH

Pressure switch, high

LIT

Level indicator transmit

PSL

Pressure switch, low

LS

Level switch

RV

Relief valve

Pressure control valve

TE

Temperature sensor

Pressure indicator

TI

Temperature indicator

LI

Orifice Gate valve Ball valve

FIGURE 8.10

Lube oil console P and I diagram.


PCV Double pass, shell and tube heat exchanger

Concentric reducer PI


1/2” per foot minimum Oil to clairifier

Top of oil reservoir Maximum oil level Working capacity 360 gallons

Purge Fill conn w/strainer Oil from clairifier

PSL

Start secondary pump o setpoint (PSIG) falling press

Dirty oil drain

PDIT

Exhauster vent

Slop

PI

Drain

Cooling water in

Retention volume 600 gallons

Coolers and filter vents

Clean oil drain

B

B

Steam Turbines

8-21

Most pressurized lubrication systems are constructed with some provision for ventilation although some systems enjoy satisfactory operation with no such provision. Effective ventilation of the lubrication system enables the reduction of moist air that affects the service life of the turbine oil. The provision of adequate ventilation is also helpful in reducing foaming where trouble from this source is encountered [54]. The following methods of ventilation are commonly used: Natural ventilation, vacuum ventilation, or dehumidifier system [54]. Figure 8.10 shows a system equipped with a vapor extractor. The extractor pulls a slight negative pressure that should result in no more than −0.5 kPa (−0.07 psi) in the bearing housing to keep oil vapors from escaping, but without pulling in atmospheric contaminants. Reservoirs normally have a connection for an oil conditioning system. Such oil conditioners can provide further purification by removing water, acids, and other contaminants not removed by the filters. These are discussed in more detail under oil maintenance. 8.4.2.2 Pumps and Drivers Two or more oil pumps are normally supplied with the lube oil system. One pump is considered the main oil pump and the other, the auxiliary. The pumps are sized with additional flow capacity to provide a positive flow of oil under all normal operating conditions and most abnormal conditions to the turbine and the driven equipment. Additional smaller pumps may be used to supply oil for special purposes such as turning gear operation; hydrostatic lift oil for highly loaded bearings; seal oil for hydrogen-cooled generators; or oil transfer through filters [28]. Positive displacement pumps have relief valves located at each pump discharge line to protect the pumps and system against excessive pressure. The main oil pump can be driven off the main turbine shaft, by an electric motor, or by a small steam turbine. Most often, the auxiliary oil pump is driven from a different source of power than the main oil pump. The auxiliary pump driver is selected to reflect availability of power or steam under emergency conditions. Should the main pump fail, the auxiliary pump will automatically start. If the pressure continues to drop, the turbine and driven equipment will shut down. Emergency situations where both pumps fail are handled by either an emergency oil pump sized to provide last-resort lubrication for coastdown or a rundown tank that provides lube oil by gravity flow. Rundown tanks are common in marine applications. On lube systems where the auxiliary oil pump is driven by a small steam turbine, an accumulator is incorporated into the system. The accumulator will maintain the required oil flow while the turbine (auxiliary pump driver) is accelerating to speed preventing a system shutdown in case of main pump failure. 8.4.2.3 Filters Twin filters with multiple cartridge filtering elements are normally used in the lubricating oil system. Filters are operated in the full-flow mode such that all oil being circulated to the turbine passes through the filter. Using two filters permits filtering element changes while the equipment is in operation. Filtration ratings should be a minimum of 25 µm, and filtration of 10 µm is typically required. Filters are sized for a maximum pressure drop of 35 kPa (5 psid) when clean and passing oil at the design temperature. Filtering elements are typically replaced when pressure drop reaches approximately 100 kPa (15 psid) above original clean value [55]. The effect of water on the filter must be considered. Water and corrosion-resistant filter cartridge materials are preferred. Such water-resistant filter cartridges should not deteriorate even if water contamination reaches 5% by volume and an operating temperature as high as 70◦ C (160◦ F). Depth type elements (e.g., cotton and nylon) can suffer from a phenomenon termed “cartridge erosion,” where oil velocity enlarges or erodes the filter passages over time, which effectively invalidates the filtration rating of the element [55]. Cartridge erosion problems are eliminated by conservatively replacing filter elements every 6 months. Filters are also replaced if the pressure drop from clean increases by 100 kPa (15 psid). Recommendations of the filter element manufacturer should be considered.


8-22


8.4.2.4 Coolers Oil coolers are usually conventional shell and tube heat exchangers with removable tube bundles. Normally, water flows through the tube bundle thus allowing for easy waterside cleaning. Oil flows through the shell side in a single pass. Coolers are usually operated so that the oil is at a higher pressure than the water, thus, reducing the severity of water contamination caused by tube failure. Typical cooling requirement is to cool oil to 120◦ F. Where conditions do not lend themselves to water-cooled heat exchangers, such as desert or subzero installations, air blast oil coolers must be considered. Coolers may be used for heating during initial oil system installation and cleansing and it is important that the system be designed for such use if desired [56]. Problems with maintaining oil temperature could be caused by improper venting, malfunctioning temperature regulators, incorrect water pressure, or badly fouled coolers [28]. Tube failure may be caused by fatigue and erosion. Excessive water flow can cause flow-induced vibration of the cooling tubes, but maintaining proper flow will reduce fatigue related problems. Water treatments and sacrificial anodes are used to retard corrosion failure of the cooler [56]. Cooler failures are responsible for the worst water contamination of the turbine oil. 8.4.2.5 Control Valves The backpressure regulator is designed to maintain a constant header pressure for all operating conditions. Normally, it is a self-operated valve, but where wide control ranges are required, pneumatic regulators complete with valve positioners are used. In addition to the backpressure regulator, pressure-reducing valves are required for all pressure levels below main header pressure. These valves are normally self-operated reducing valves but where wide control range is required, pneumatic operators complete with valve positioners may be furnished. 8.4.2.6 Piping Lubrication system components are joined together by the necessary piping to make the system functional. This includes provisions for the mounting of control instrumentation, such as pressure gauges, temperature gauges, switches, and monitoring and safety devices. Piping may be either carbon or stainless steel. Stainless steel is preferred due to superior corrosion resistance and is used extensively in refinery applications. The header piping connects the lube oil console to the various components being lubricated, such as bearings and seals. Used oil is returned to the reservoir through drain piping. Oil drains are sized to run no more than half full when flowing at a velocity of 0.3 m/sec (1 ft/sec) and are arranged to ensure good drainage. Horizontal runs slope continuously, at least 40 mm/m (1/2 in./ft), toward the reservoir [57]. 8.4.2.7 Safety and Monitoring Devices Lubrication system instrumentation is located throughout the system as shown on the schematic oil flow diagram. Monitoring devices, such as pressure gauges, safety devices, and alarm and trip switches are generally mounted on header piping close to the components being lubricated. A low-pressure start-up switch signals the auxiliary pump to start if pressure is too low. Temperature indicators are provided at bearing and seal outlets and at the inlet and outlet of coolers. Pressure indicators are generally provided at each pressure level. A sight-flow indicator is provided at the outlet of each turbine shaft bearing and each turbine thrust bearing. 8.4.2.8 Cleaning and Flushing All reasonable effort must be made to limit the introduction of contaminants into the lube oil system during construction. Proper cleaning and preservation of lube system components must be performed prior to system shipment. Different preservatives are used depending on the environment and expected storage time [58]. All units employing forced-feed oiling systems should have the entire lubrication system thoroughly flushed before operation. The importance of this step cannot be overemphasized. All dirt, rust scale, weld


Steam Turbines

8-23

slag, or other contaminants that have been introduced into the oil system during storage, transportation, and fabrication at the jobsite must be removed by a continuous flushing operation or in extreme cases that do not involve stainless steel pipe, by pickling and cleaning. In addition to flushing during the commissioning, the system should also be flushed if left idle for a long time. Most turbine manufacturers provide special instructions for the oil flush. In the absence of such instructions, industry recommendations should be consulted such as those detailed in ASTM D6439 [59] or API RP 686 [58]. Flushing the system may require the use of external pump and such preparations should be made in advance. The bearings and bearing area should be bypassed until the system is proven to be clean. The flushing should continue until the required cleanliness is achieved based on inspection of the flushing filters or strainers, patch test, particle counters, or ISO 4406 cleanliness level. Flush oils, operating oils, and preservative oils must be compatible to preclude foaming, formation of emulsions, or breakdown of oil additives. Compatibilities and limitations may generally be obtained from the oil supplier. A system that is to use phosphate ester fluids must be flushed with phosphate ester fluid since such fluid is incompatible with mineral oil. The same may apply to other synthetic oils. 8.4.2.9 Lube System Maintenance Lube systems must be periodically inspected and maintained to ensure their proper operation. As a minimum, the following regular checks should be performed: • Check filter pressure drop and replace elements as recommended. • Check the oil reservoir level and add oil as required. • Periodically check operation of auxiliary oil pump by operating pump and returning to auxiliary duty. In addition, turnaround maintenance of the lube oil system should be performed at 1 to 3 yr intervals, as normal plant maintenance permits. Care must be taken to keep contaminants out of the lube oil circuit during bearing changes, filter changes, top up, and other maintenance activities.

8.5 Turbine Oil Equipment vendors often have turbine oil standards detailing the minimum characteristics required for successful turbine operation. In the absence of such standards, an internationally recognized turbines oil specification such as shown in Table 8.3 should be used.

8.5.1 Physical Properties Turbine oil performs four functions (1) Lubricate bearings and gears; (2) cool lubricated parts, carrying heat away from hot surfaces such as bearings and shafts; (3) act as a hydraulic fluid for governor, control valves, and safety devices; and (4) act as a sealant for gas seals such as hydrogen shaft seals in generators or gas seals on compressors. Each of these functions require an oil that is suitable with respect to several physical, chemical, and performance properties. Some physical properties frequently used to characterize turbine oils with corresponding American Society for Testing and Materials (ASTM) test methods TABLE 8.3

Standards for Turbine Oils and Hydraulic Fluids

Standard designation ASTM D4293 ASTM D4304 ISO 8068 MIL-PRF-17672D MIL-PRF-17331H

Standard title Standard specification for phosphate ester-based fluids for turbine lubrication Standard specification for mineral lubricating oil used in steam or gas turbines Petroleum products and lubricants — petroleum lubricating oils for turbines (categories ISO-L-TSA and ISO-L-TGA) Performance specification: hydraulic fluid, petroleum, inhibited Performance specification: lubricating oil, steam turbine and gear, moderate service


8-24 TABLE 8.4

Handbook of Lubrication and Tribology Standardized Lubricating Oil Analytical Techniques

Physical or chemical property

ASTM designation

ISO designation

ISO viscosity grade Kinematic viscosity at 40◦ C, 100◦ C Viscosity index

D2422 D445 D2270

ISO 3448 ISO 3104 ISO 2909

Pour point Flash point Total acid number (TAN)

D97 D92 D974

ISO 3016 ISO 2592 ISO 6618

Foaming characteristics Air release

D892 D3427

ISO 6247 DIN 51 381

Water separability (Demulsibility) Rust prevention

D1401 D665

ISO 6614 ISO 7120

Corrosiveness to copper

D130

ISO 2160

Oxidation stability (TOST) Rotating pressure vessel oxidation test (RPVOT) Acid number Karl Fischer titration Color ISO cleanliness code

D943 D2272

ISO 4263

D664 D1744 D1500

ISO 6619 ISO 6296 ISO 2049 ISO 4406

Purpose Overall viscosity classification Relates to viscosity at normal operating conditions Empirical comparison of viscosity and temperature characteristics Measures low temperature flow properties Low value indicates volatile components Determination of acidity of new and used oils by titration with KOH Foaming characteristics of lubricating oils The oil’s capacity to separate entrained air over a period of time Emulsion characteristics of oil Ability of oil to prevent rusting of steel surfaces in presence of water Indicates tendency of oil to corrode copper and copper alloys Oxidation stability of mineral oils Tests remaining oxidation life of in-service oils Indicates acid level Measures the water content of oil Measures color Measures oil cleanliness

are summarized in Table 8.4. Detailed descriptions of the ASTM methods are available in the ASTM Handbook [59]. The most important physical property is viscosity. Table 8.5 gives the viscosity ranges for typical mineral lubricating oils used in steam turbines. Typical viscosity grade numbers are ISO-VG-32, VG-46, VG-68, VG-78, and VG-100 such that the viscosity grade numbers indicate the average oil viscosity in centiStoke units at 40◦ C (104◦ F). In order to reduce the power losses at the bearings and improve the responsiveness of hydraulic components, the lowest acceptable lubricant viscosity is normally selected. As a result, the usual lubricant employed in a common oil system is ISO VG-32 turbine oil cooled to a supply temperature of 120◦ F after the cooler. Other viscosity grades are also used. ISO VG-46 turbine oil cooled to a supply temperature of 140◦ F after the cooler is commonly used in desert, arid, and offshore applications where air blast coolers are utilized, or where the ambient temperature is quite high [56]. Oils used for ringoiled turbine bearings tend to be higher viscosity such as ISO VG-68 or VG-100. Oils used for shipboard propulsion may be ISO VG-68 to VG-100 and may have mild antiscuff additives. It is important to note that the lube oil system and the turbine rotordynamics are designed considering a specific oil viscosity. Turbine lube systems must be maintained with lubricants of the recommended viscosity, and the viscosity specification should not be changed without proper engineering review.

8.5.2 Formulation To achieve the desired physical, chemical, and performance properties, turbine oil is formulated with a base fluid and additive package consisting of rust and oxidation (R&O) inhibitors. Steam turbine oils are essentially special grades of R&O oils, formulated to give better oxidation resistance and longer life in a steam turbine [60]. While industry standard lube oil bench tests can provide great insight into the performance and life expectancy of turbine oils, both turbine original equipment manufacturers (OEMs) and oil suppliers generally agree that past successful performance of a particular oil under similar conditions is the best overall representation of quality and performance [61].


Steam Turbines TABLE 8.5

8-25

Physical Requirements for Turbine Oils

ISO viscosity grade Kinematic viscosity, mm2 /sec at 40◦ C, min. at 40◦ C, max. Military specification Military symbol ISO viscosity grade Kinematic viscosity, mm2 /sec at 40◦ C, min. at 40◦ C, max. At 100◦ C Pour point, ◦ C, max. Flash point, ◦ C, min. Viscosity index, min. Total acid number (TAN), mg KOH/g, max. Corrosiveness to copper, max. Rust prevention Water, percent Valve sticking characteristics Foaming characteristics Sequence 1, mL max. Sequence 2, mL max. Sequence 3, mL max. Air release Water separability Oxidation stability, min.

Light turbine oil

Medium turbine oil

Medium-heavy turbine oil

Heavy turbine oil

Test method

32

46

68

100

D2422 D445

61.2 74.8

90 110 MIL-L-17331J 2190 TEP

28.8 35.2 2075 T-H 32

41.4 50.6 MIL-L-17672D 2110 T-H 46

2135 T-H 68

D2422 D445

28.8 35.2 Report −29 157 94 0.20

41.4 50.6 Report −23 163 94 0.20

61.2 74.8 Report −18 171 94 0.20

74 97 8.0 −6 204

1 Shall pass None Shall pass



1 None None Shall pass

65/0 65/0 65/0

65/0 65/0 65/0

65/0 65/0 65/0

40/40/3 1000 h

40/40/3 1000 h

40/40/3 1000 h

65/0 65/0 65/0 20 40/—/3 1000 h

0.3

D97 D92 D2270 D974 D130 D665 D95 D892

D3427 D1401 D943

8.5.2.1 Base Oil The base oil stock of a turbine oil comprises more than 98% of the formulation. The base oil is categorized as either conventional solvent refined mineral-based (API Group I), or hydroprocessed mineral-based (API Group II) oil. Group II base oils contain fewer heteroatoms (sulfur, nitrogen, oxygen), and have less aromatic content than Group I base oils. When properly formulated, Group II turbine oils will have longer oxidation life, less deposit forming tendencies, improved water shedding ability, and overall higher performance than do Group I turbine oils [60]. One advantage of the conventional mineral-based (Group I) turbine oils is better innate solvency than the hydroprocessed (Group II) oils. The better solvency of the Group I turbine oils provides better additive package retention and increased ability to dissolve oxidation products that could otherwise potentially lead to varnish and sludge. While Group I and Group II base stocks are compatible with each other, the additive packages used to formulate the respective turbine oils may be incompatible with the overall mixture. Mixing oils can therefore cause sludge formation and additive dropout [62]. For this reason, compatibility between products is an important consideration when mixing two oils. 8.5.2.2 Additives Additives are used to improve the performance of the oil. Although additives are to some extent consumed in performing their functions, they can be replenished through normal lubricant make-up thereby enabling suitable performance for longer periods. Note that newer machine designs offer less oil loss and therefore do not benefit as much from this effect as did older machines exhibiting greater oil loss. The main types of additives include oxidation inhibitors, rust inhibitors, foam inhibitors, and demulsifiers.


8-26


8.5.2.2.1 Oxidation Inhibitors Antioxidants are the additives, which have the strongest influence on the useful life of turbine oils. They generally function either by free radical inhibition, by hydroperoxide decomposition, or by deactivation of metal catalysts. The two major types of antioxidants used in turbine oils are arylamines and hindered phenols [63], and they work as free radical inhibitors. A mixed phenol-amine has certain advantages over the use of a single antioxidant system. Other additives and combinations of additives are also used to suppress oxidation. In particular, metal deactivators are used to suppress oxidation by reacting with metal ions and surfaces to inhibit their catalytic activity [64]. 8.5.2.2.2 Corrosion Inhibitors Highly refined oils lose their metal-wetting ability and are easily displaced by water. For this reason, corrosion inhibitors are necessary to prevent corrosion. New turbine oils contain a rust-inhibitor additive and must meet ASTM Test Method D 665. These corrosion inhibitors typically work based on the physical adsorption principle. In action, the corrosion inhibitor “plates out” on surfaces, forming a film that resists displacement by water and, therefore, protects the surfaces from contact with water [65]. Corrosion inhibitors used in turbine oils are polar and thus susceptible to water washout. Alkenyl succinic acids are therefore widely used due to their resistance to water washout [66]. 8.5.2.2.3 Foam Inhibitors Foam additives must be carefully selected in order to prevent excessive foam formation, but still retain short air release times [67]. Highly refined hydrotreated base oils have lower foaming tendencies than conventionally refined base oils. Foam inhibitors work by decreasing the gas-lubricant interfacial tension. Liquid silicones are an effective antifoamant, but also act as an air-emulsion stabilizer, negatively influencing the air release properties of the turbine oil as it resides in the stilling portion of the equipment. 8.5.2.2.4 Demulsifiers Demulsifiers destabilize oil–water emulsions by changing the interfacial tension of oil and water thereby allowing their separation [64]. Conventional mineral-based (Group I) turbine oils usually contain demulsifying additives whereas hydroprocessed (Group II) turbine oils have good demulsibility without an additional additive.

8.6 Performance Features of Turbine Oils The following oil performance features must be retained to ensure safe and continuous operation of the turbine (1) viscosity; (2) oxidation stability; (3) freedom from sludge; (4) anticorrosion protection; (5) water separability [68]; and (6) air separability and resistance to foaming.

8.6.1 Viscosity Viscosity is measured by ASTM Test Method D 445. Viscosity is the most important characteristic of turbine oil, as the oil film thickness under hydrodynamic lubrication conditions is critically dependent on the oil’s viscosity characteristics. Viscosity also affects journal bearing stiffness and damping properties, which determine the vibration characteristics of the turbine. Viscosity of most new oils may vary by ±10%. A change in viscosity up to 10% is not in itself likely to cause trouble; however, a change in viscosity of 5% from its original value should be investigated for the cause. A change in viscosity is usually caused by contamination or top off with the wrong lubricant rather than by degradation of the oil. Drop in oil viscosity is a particular concern where turbine driven compressors are used in the compression of hydrocarbon gases because the viscosity change may be caused by contamination of the oil from the lighter hydrocarbons [56].


Steam Turbines

8-27

8.6.2 Oxidation Stability During steam turbine operation, the lubricating oil is subjected to relatively severe oxidizing conditions. These are due primarily to the influence of heat, the presence of water and entrained air, and the catalytic action of substances in contact with the oil, particularly copper and ferrous metals [69]. Under these influences, the antioxidants are gradually used up and the oxidation stability decreases. ASTM D943 Turbine Oil Stability Test (TOST) is used to evaluate the oxidation characteristics of new inhibited steam turbine oils. This is an accelerated oxidation test; actual service should be much longer than test report hours [61]. Since TOST testing can take longer than a year, it is impractical as an in-service oil test.

8.6.3 Freedom from Sludge and Deposits Deposits are generally formed due to oxidation of the turbine oil, soap formation, microbiological growth, contamination by water containing salts, and solid particulate contamination [66]. Process gases can also react with the oil and its additives to form deposits. One such example is a turbine driven ammonia compressor in which the oil became contaminated with ammonia. The acidic rust inhibitor used in the turbine oil reacted with ammonia to form an insoluble resinous product [70]. Filtration and centrifugation can remove sludge and other products from oil as they are formed, but if oil deterioration is allowed to proceed too far, sludge will deposit in parts of the equipment and system flushing and an oil change may be required [71].

8.6.4 Corrosion Protection Protection against rusting is very important due to the common presence of water in turbine oils and water vapor in the ambient air. Rusting may occur below the oil surface, at the oil surface, or in the vapor spaces above the oil surface. Rusting requires oxygen, water, a corrodible surface, and time. Effective corrosion protection requires the elimination of any one of these items. Oil acts to protect against corrosion by coating structural surfaces with corrosion inhibitor thereby denying access of water to corrodible surfaces. ASTM D665 is used to evaluate the rust-preventing characteristics of steam-turbine oil in the presence of water. Procedure A is used for land turbines where condensed steam or humidity from air is the water source. Procedure B is used for marine-service ocean-going vessels where salt water can be a water source. Present additive technology has been found to be highly effective at preventing rusting problems below the oil surface in full flow conditions. When rusting occurs below the oil surface, it is frequently caused by galvanic corrosion, and it is noticed in areas where there is little oil movement and where free water collects, such as the bottom of the oil reservoir. Galvanic corrosion is caused by contaminant particles settling out of the oil and the presence of water. Particulate matter can create galvanic cells and act as nuclei for air bubbles [34]. Factors that influence galvanic corrosion are impurity concentrations, the pH of the water, and temperature [66]. Galvanic corrosion shows up as black rust. Rusting at the oil surface is typically caused by liquid water standing on the surface [72]. Most rust problems occur above the oil in what is known as the vapor space. Vapor spaces are present in steam turbine bearing pedestals, oil return lines, sumps, and gear cases. The air in these vapor spaces will contain water vapor from the relative humidity of the air drawn into the system and from the evaporation of water entrained in the oil. In addition, salt particles that can act as corrosion-sponsoring nuclei also may be present [73]. Water vapor tends to condense on the cooler parts of the circulation system, such as the underside of the reservoir top, inside return-oil piping above oil level, in bearing pedestals, and around governor parts [74]. Corrosion in the vapor space results in formation of scaly red rust.

8.6.5 Water Separability (Demulsibility) A lubricant’s ability to separate readily from water is one of the most important requirements of a turbine oil. Water must readily separate from oil in the drain tank so that it is dry when pumped to the system.


8-28


Demulsibility is influenced by oxidation and contamination from dirt or metallic particles. Resistance to oxidation helps preserve the demulsibility characteristics of the oil. Normally, if the oil is in good condition, water will settle to the bottom of the storage tank, where it should be drained off as a routine operating procedure [71]. Water may also be removed by purification systems. If turbine oil develops poor demulsibility, significant amounts of water will stay in the system and create problems such as increased oxidation, additive depletion, and corrosion. ASTM D1401 is used to test the demulsibility characteristics of oil.

8.6.6 Air Separability and Resistance to Foaming All oils will foam in some degree. Foaming of the present day turbine oils should not, however, occur unless the oil is contaminated or subjected to abnormal aeration. Antifoam additives suppress foam, but in doing so may also slow down air release leading to air entrainment. Air entrainment in the oil has been known to cause pressure surge in oil systems, interruptions in oil supply, excessive formation of foam [75], and reduced hydraulic control. Care must be taken such that improving the antifoam characteristics of turbine oil does not lead to unacceptable air separability characteristics. Turbine circulation systems have been constructed to eliminate conditions that have been found to cause foaming such as leaky pump suctions, excessive splashing of oil returning to the reservoir, oil-return lines of insufficient size or capacity, and insufficient venting. Wide differences in temperature between the fresh oil (as added) and the oil in the system may contribute to foaming [76]. Serious cases of excessive foaming may be due either to mechanical faults of the type listed or to oil contamination [77]. Problems with excessive foaming may also be due to mixing of incompatible lubricants [63] or the use of excessive antifoam inhibitor. Air entrainment issues are also affected by system design. In particular, the stilling period of the lube oil system can affect the air entrainment characteristics of oil. Machines that provide short stilling periods for the oil have displayed air entrainment/release characteristics that seem to counter those displayed during standard air release testing (ASTM D3427). Such machines with very short stilling periods have displayed increased air entrainment when nonsilicone antifoamants have been used and it is suspected that the silicone antifoams discourage the initial air entrainment during the agitation period [78].

8.7 Degradation of Turbine Oils in Service Factors responsible for oil degradation in service include contamination, additive depletion, oxidation, and bacteriological deterioration.

8.7.1 Contamination Contaminants will unavoidably find their way into the lubricating oil. The following types are most common: water, oil soluble contaminants, and solid particles. 8.7.1.1 Water Water is always present in oil in solution and may also be present in free or emulsified form. The solubility of water in oil is temperature dependent. Water in solution has no adverse effect on lubricating properties and will not cause corrosion; however, when hot oil subsequently cools, some water may come out of solution as very fine droplets dispersed throughout the oil [28]. This water is very likely to cause corrosion of steel parts and may also cause other problems, (e.g., foaming, sludge formation, and change of viscosity). In addition, water can also lead to oxidation, additive removal, bacteriological contamination, as well as reducing filter element life. Water enters the oil system from the condensation of humid air by system temperature fluctuation; from steam through the turbine gland seals; or from leaking oil coolers [73]. Leaking gland seals is the


Steam Turbines

8-29

most common source of water. Leaking oil coolers is the most detrimental particularly since cooling water leaks will have moderate to high concentrations of dissolved solids. In extreme cases, a rupture of the heat exchanger can cause massive amounts of water to enter the machine compartment [79]. Free water generally exists above a saturation level of around 120 to 150 ppm, and oil becomes cloudy in the range of 200 to 500 ppm [80]. A centrifuge is effective in removing free water down to about 30 ppm above the saturation level. Different methods for the testing of water exist. The simplest is visual inspection followed by “crackle” or hot-plate test, which can indicate the presence of water in oil due to boil off. Another test is the Fourier Transform Infra-Red (FTIR) spectrometry. The Karl Fischer Titration, ASTM Test Method D 1744, is the most accurate method for water testing. Differing limitations for water are noted by different manufacturers. In general, the water content should never be allowed to exceed 2500 ppm (0.25%). ASTM D4378 cites 1000 ppm (0.1%) as a warning limit. Depending on the design and application, some manufacturers will require a limit of 500 ppm (0.05%). 8.7.1.2 Soluble Contaminants Oil soluble contaminants may include gases, solvents, other lubricants, flushing oils, preservatives, and sealants. Gases and some light solvents can be removed by vacuum dehydration methods. Other contaminants cannot be removed. The presence of such contaminants requires the consultation of the oil supplier and the turbine manufacturer. A common source of dissolved gases is the oil seals used in some generators and compressors. 8.7.1.3 Solid Particles Abrasive contaminants can damage bearings, journals, and control mechanisms. Improved practices such as better preservation of the turbine and its components when not in operation, high velocity system flushing during commissioning, and use of full flow filtration during operation have led to a significant reduction in failures due to abrasive contaminants [81]. Cleanliness of the system oil can be determined by gravimetric means by ASTM F 311 or F 312 or by particle counting. Allowable contamination level is dependent upon the individual turbine application and components in the system. ISO 4406 cleanliness levels ranging from 18/16/13 to 16/14/11 are commonly applied to steam turbine service. The three digits of the ISO 4406 code refer to the number of particles per milliliter greater than 4-, 6-, and 14-µm respectively. It should be noted that further reductions in contamination beyond manufacturer recommendations might lead to improvements in reliability that can be cost justified [82].

8.7.2 Additive Depletion Additives are used up in the performance of their function. In other cases, the additives are removed due to reaction with contaminants or drop out due to problems of compatibility. Oil suppliers are often able to replenish additives by sweetening the oil.

8.7.3 Thermal and Oxidative Degradation The oil acts as a heat transfer fluid with the overall system design determining the heat load on the oil. Factors such as smaller oil reservoirs, higher shaft surface velocity, and higher shaft and bearing temperatures all contribute to environmental conditions that degrade the oil by thermal stress leading to oxidation. Oxidation occurs by chemical reaction of the oil with oxygen. The first step in the oxidation reaction is the formation of hydroperoxides. Subsequently, a chain reaction is started and other compounds such as acid, resins, varnishes, sludge, and carbonaceous deposits are formed [71]. Oxidation products may further lead to rust and corrosion, and promotion of foaming and poor demulsibility. The oxidation rate is influenced by the presence of water, contaminants, entrapped air, and temperature. The oxidation rate of a fully inhibited mineral oil is quite low at temperatures less than 60◦ C and will double for every 10◦ C rise in temperature [83].


8-30


For in-service oil testing, the oxidation stability reserve is best determined by the rotating pressure vessel oxidation test (RPVOT), ASTM test Method D 2272, and by total acid number (TAN), ASTM Test Method D 974.

8.7.4 Biological Deterioration Lubricating oils are susceptible to biological deterioration if the proper growing conditions are present. Procedures for preventing and coping with biological contamination include cleaning and sterilizing, addition of biocides, frequent draining of moisture from the system, and avoidance of dead-legs in pipes [71]. Sustained high water content can lead to bacterial and fungal growth in the system. This can cause filter blocking and formation of deposits. The most effective antimicrobial measure is the establishment of preventative procedures such as frequent draining of free water from the oil reservoir. Biocides are used to prevent microorganism growth. Sterilization by heat is also effective.

8.7.5 Turbine Oil Severity The expected service life of a turbine lubricant depends considerably on the severity of the application. Many low severity steam turbines have a history of requiring a full lubricant changeout only every 10 to 20 yr or longer, with periodic top-up with fresh oil [67]. Certain environmental conditions, however, can result in or accelerate lubricant degradation and reduce life. As noted, factors responsible for oil degradation in service include contamination, additive depletion, and thermal, oxidative, or physical breakdown. Other important factors affecting service life are (1) type and design of lubrication system, (2) condition of the system after construction, and (3) oil makeup rate. These factors vary from unit to unit so that service life is difficult to predict solely on original oil properties [84]. One method for determining the service conditions for each operating unit is to use a property called the turbine severity level (B), which is defined as the percent of fresh oil oxidation resistance or oxidation inhibitor lost per year due to oil reactions [85]. The equation for turbine severity is B = M · (1 − X /100)/(1 − e−M ·t /100 )

(8.3)

where B is turbine severity, % of fresh oil oxidation resistance lost per year due to oil reactions in the turbine, M is fresh oil makeup, % per year, t is years of oil use, and X is used oil oxidation resistance by ASTM D2272, % of fresh oil. A lubrication system with a high severity level requires frequent makeup or completely new charges, whereas one with a low severity level may have no problems with routine makeup [86]. The method requires periodic testing of the lubricating oil. Large steam turbines should have their turbine severity determined. The severity constant is different for each turbine, and varies widely between 5 and 30 for large turbines [87]. Figure 8.11 shows the importance of makeup rate for maintaining oil quality in a high-severity turbine where B = 25%/yr [85].

8.8 Lubricant Maintenance Small turbines with ring-lubricated bearings, and governors with sumps require periodic changes in lubricant. The quantities of oil are small, and it is often more economical to change the oil rather than to maintain it. Change periods of 1 yr are not uncommon and are set by regular change intervals, by monitoring the acid number or by more sophisticated monitoring. In larger turbine systems that employ circulating oil systems using more than 200 l (roughly 50 gal) of oil that require long periods of continuous operation, oil analysis generally proves more profitable than a routine time/dump program [88]. A life of up to 30 yr is desirable because of the outage and oil change


Steam Turbines

8-31

100 90 80 70 60

40 30 M =3

0%

20 M 5%

=2

0%

10 9 8 7

M=2

M = 10%

M = 0%

Percent of original oil life remaining, X

50

6 5 0

4

8

12

16

20

Years of use

FIGURE 8.11 Effect of makeup rate on oil degradation for turbine severity, B = 25%/yr. (From DenHerder, M.J., Vienna, P.C., Lubrication Engineering, 37, 67, 1981. With permission.)

TABLE 8.6

Standards for Turbine Oil Maintenance

Standard designation ASTM D4378 ASTM D4057 ASTM D6224 ASTM D6439 IEC 60962 IEC 60978

Standard title Standard practice for in-service monitoring of mineral turbine oils for steam and gas turbines Standard practice for manual sampling of petroleum and petroleum products Standard practice for in-service monitoring of lubricating oil for auxiliary power plant equipment Standard guide for cleaning, flushing, and purification of steam, gas, and hydroelectric turbine lubrication systems Maintenance and use guide for petroleum lubricating oils for steam turbines Maintenance and use guide for petroleum lubricating oils for triaryl phosphate ester turbine control fluids

costs involved. In such systems, regular sampling and testing can indicate the need for oil conditioning. Many oil suppliers offer programs to meet specific lubrication maintenance requirements. Standards for turbine oil maintenance are listed in Table 8.6. Such standards offer a guideline for oil-monitoring and maintenance. Other methods may be applied depending on the application. One such standard, ASTM D4378, Standard Practice for In-Service Monitoring of Mineral Turbine Oils for Steam and Gas Turbines is used in the power generation industry [89]. As with any oil monitoring program, proper sampling is important. In-service oil should be tested at sufficient intervals to detect contamination, oxidation, and additive depletion. Key tests include appearance and color, water content, viscosity, total acid number, rust test, cleanliness, and RPVOT [89]. Systems that are exposed to volatile gases or liquids


8-32


may also benefit from flash-point testing. Maintaining the lubricant may require new oil makeup, lube oil conditioning, and refortification.

8.8.1 New Oil Makeup Oil is lost due to leakage and to system maintenance such as draining off impurities, and filter changes. There is considerable variation with respect to the amount of makeup oil required for a steam turbine. Makeup rates can range from less than 5%/yr to more than 30%/yr in extreme cases. The average makeup rate in the United States is 7 to 10%/yr [89]. The compatibility of the system oil and the makeup oil are of critical importance. Compatibility is described as a lubricant’s ability to be mixed with another lubricant without detriment to the properties and the characteristics of either lubricant. The introduction of Group II oils has caused some concerns with respect to compatibility with Group I oils. In particular, the different additives and the solubility of those additives is a concern when mixing different oils especially those involving different base stocks. Problems involving excessive air entrainment, varnish particle build up, development of sludge, sticking of governor proportional valves, and plugging of governor filters has been noted on hydroturbines operating on turbine lubricants [90]. In some cases, a complete system flush may be required to introduce a new oil. The use of makeup oil that is the same oil as is already in the system is preferred for the elimination of compatibility issues.

8.8.2 Lube Oil Purification All circulating lube oil systems use filters to remove particle contaminants and purify the oil. Devices for removing liquid contaminants such as water will also improve system reliability. The most common devices for removing liquid contamination are settling tank, centrifuge, coalescing filter, and vacuum dehydrator. The settling tank works best on a batch basis. The centrifuge, coalescing filter, and vacuum dehydrator are applied continuously with 10 to 20% of the volume of oil in the turbine system every hour. Systems of this type tend to remove impurities as fast as they enter the oil, thereby avoiding accumulations. Settling tank — Oil contaminants that are heavier than oil can be separated by gravity alone. Such settling is best accomplished in a settling tank that is separate from the main oil tank. Settling times can be very long and the results are often less adequate than the onstream methods. Centrifuge — In a centrifugal purifier, or centrifuge, centrifugal force is used to accomplish the separation of contaminants heavier than oil. A separating force several thousand times that of gravity is produced by rotating the liquid at 7,000 to 15,000 rpm. The centrifuge is particularly effective in removing water and larger, heavier particles of solid impurities. The extent to which extremely fine solid particles are removed depends on the rate of throughput and other factors. Centrifuges are capable of removing free water and solids. Coalescing filter — A coalescing filter system uses special cartridges to combine small, dispersed water droplets into larger ones. The larger water drops are retained within a separator screen and fall to the bottom of the filter while the dry oil passes through the screen. Coalescers are capable of removing free water and solids. Vacuum dehydrator — A vacuum dehydration system removes water from oil through the application of heat and vacuum. The contaminated oil is exposed to a vacuum and heat. The water is removed as vapor. The vacuum dehydrator removes not only the free water, but also the dissolved and suspended water well below the solubility point (down to 10 ppm). In addition, vacuum dehydrators also deaerate and degasify the oil [91].

8.8.3 Refortification Refortification refers to the act of adding a predetermined amount of additive to a clean, dry, used lubricant to replenish some of the depleted additives [92]. In most cases, refortification and purification are used together.


Steam Turbines

8-33

8.9 Fire-Resistant Fluids Fire-resistant fluids are used in the hydraulic actuators of large steam turbines operating at very high temperatures in excess of the auto-ignition temperature of turbine oil. Since the early 1970s, phosphate esters have been the only fire-resistant fluids approved by the turbine builders for use as a turbine control fluid although small amounts of more flammable carboxylate or synthetic esters have been used in refurbished systems [93].

8.9.1 Properties The main advantage of phosphate esters is their fire-resistance. Phosphate esters tend to have higher flash and fire points, higher auto-ignition temperatures and perform better in spray flammability and wick-type fire propagation tests [94]. Auto-ignition temperatures are in the region of 550–590◦ C. The triaryl version of the phosphate ester possesses inherent self-extinguishing properties because the fluid does not create enough energy to support its own combustion. Triaryl phosphates, in addition to their fireresistant properties, have good thermal stability, excellent boundary lubrication properties, low volatility, fair hydrolytic stability [94], adequate air release, and low-foaming properties. The density of phosphate esters is roughly 30% higher than mineral oils necessitating some additional consideration with respect to lube oil system design. Phosphate ester-based fluids are described in ASTM D4293. Viscosity grades are either ISO VG-32 or ISO VG-46. Phosphate ester fluids can be incompatible with some seal and insulative materials as well as certain paints thus making the pressurization system design and maintenance critical.

8.9.2 Degradation In service, phosphate esters are subject to deterioration as a result of hydrolysis, oxidation, and contamination. In the case of triaryl phosphate ester hydraulic fluids, contamination may be by water, particulates, mineral oil, and chlorine or chlorinated materials [95]. The principle degradation pathway for phosphate esters in steam turbine-generator lubrication systems is hydrolysis. While water is inevitably present in the fluids, its continued high concentration can be tolerated if fluid acidity is controlled [96]. As the solubility of water in phosphates is very much higher than in oil (reaching about 2500 ppm at 25◦ C), free water is not usually a problem and the level of fluid acidity will normally determine the suitability of the fluid for continued use. Many of the problems with the use of phosphate esters in turbine applications are associated with the development of acidity due to hydrolysis or oxidation. Since acidity development can cause corrosion, further accelerate the rate of hydrolysis, and is probably an early stage in the process of deposit formation, the maintenance of acidity levels of less than 0.5 mg KOH/g and preferably less than 0.2 mg KOH/g is strongly recommended [97]. Contamination by mineral oils can impair fire resistance, as well as being incompatible with various seals. High chlorine content can cause servo valve electrokinetic wear [95].

8.9.3 Condition Monitoring The following properties are considered necessary for the in-service testing of phosphate esters; appearance, chlorine content, color, mineral oil content, total acid number or neutralization number, fluid cleanliness, particle size, resistivity, viscosity, water content, and air release. The parameters that are of most concern are the increase in acidity, water content, and particulate contamination level. When triaryl phosphates degrade the most common result is an increase in acidity with little effect on viscosity change. Triaryl phosphate ester fluids are condemned if the acid number exceeds 0.2 over the original value (typically 0.03) [98]. Water should be kept below 2000 ppm. Alternate guidelines for maintenance and use of triaryl phosphate ester turbine control fluids can be found in IEC 60978.


8-34


8.9.4 Maintenance The key to the cost-effective use of phosphate esters is the use of conditioning media to remove acid degradation products. Fuller’s earth and activated alumina have provided years of acceptable service; however, new adsorbents based on ion-exchange resins, may allow the fluid to be kept in the system for many years. Vacuum dehydration is required to remove the displaced water [93]. Phosphate ester hydraulic fluids require additional consideration of the lube oil system. Their use in high-pressure (1000 psi) systems requires fine filtration (0.5 to 5 µm) to protect more closely fitted pumps and control valves [99]. In addition, adsorbent filtration of phosphate ester hydraulic fluids using fullers’ earth, activated alumina, or ion exchange resin is needed to control fluid acidity. Adsorbent filters remove dissolved contaminants, such as acids, that are not removed economically or at all by other processes [100]. There is a tendency for these types of filters to remove additive materials. For this reason, adsorbent clay filters are typically not used on turbine oils, but are used for purifying fire-resistant phosphate ester hydraulic fluids as used in turbine control systems. The filters are most often used in a continuous bypass mode with 1–3% treatment ratio or they are used intermittently in accordance with changes in the acid number [97]. A fine particulate filter must be placed in series and downstream of the fuller’s earth filter to control particulates.

References [1] Steam, Its Generation and Use, 40th ed., The Babcock & Wilcox Company, New York, chap. 57. [2] Church, E.F., Steam Turbines, 3rd ed., McGraw-Hill Book Company, Inc. New York, 1950, chap. 1. [3] Electric Power Plant Design, Technical Manual TM 5-811-6, Department of the Army, Washington, DC, 1984. [4] API Standard 611, General-Purpose Steam Turbines for Petroleum, Chemical, and Gas Industry Services, 4th ed., American Petroleum Institute, Washington, DC, 1997. [5] API Standard 612, Special-Purpose Steam Turbines for Petroleum, Chemical, and Gas Industry Services, 4th ed., American Petroleum Institute, Washington, DC, 1997. [6] NEMA Standard SM-23, Steam Turbines for Mechanical Drive Service, National Electrical Manufacturers Association, Washington, DC, 2002. [7] NEMA Standard SM-24, Land-Based Steam Turbine Generator Sets 0 to 33,000 kW, National Electrical Manufacturers Association, Washington, DC, 2002. [8] Military Specification MIL-T-17286D, Turbines and Gears, Shipboard Propulsion and Auxiliary Steam; Packaging of, Department of the Navy, Washington, DC, 1989. [9] Military Specification MIL-T-17523, Turbine, Steam, Auxiliary (And Reduction Gear) Mechanical Drive, Department of the Navy, Washington, DC, 1982. [10] Military Specification MIL-T-17600D, Turbines, Steam, Propulsion Naval Shipboard, Department of the Navy, Washington, DC, 1982. [11] IEC Specification 60045-1, Steam Turbines — Part 1: Specifications, IEC, 1991. [12] Wills, J.G., Lubrication Fundamentals, Marcel Decker, Inc., New York, 1980, chap. 9. [13] Zeidan, F.Y. and Herbage, B.S., Fluid film bearing fundamentals and failure analysis, Proceedings of the 23rd Turbomachinery Symposium, Turbomachinery Laboratory, Texas A&M University, College Station, TX, 1999, 161. [14] Nicholas, J.C., Tilting pad bearing design, Proceedings of the 23rd Turbomachinery Symposium, Turbomachinery Laboratory, Texas A&M University, College Station, TX, 1999, 81. [15] New, N.H., Experimental comparison of flooded, directed, and inlet orifice type of lubrication for a tilting pad thrust bearing, Journal of Lubrication Technology, 96, 22, 1974. [16] Mikula, A.M., The leading edge groove tilting pad thrust bearing: recent developments, Trans. ASME, Journal of Tribology, 107, 423, 1985. [17] Gardner, W.W., Hydrodynamic Oil Film Bearings: Fundamentals, Limits and Applications, Waukesha Bearings Corporation, Waukesha, WI, 1998.


Steam Turbines

8-35

[18] Kingsburg, G.R., Friction and wear of sliding bearing materials, in Friction, Lubrication, and Wear Technology, ASM Handbook, Vol. 18, ASM International, 1992, 741. [19] Ludema, K.C., Failures of sliding bearings, in Friction, Lubrication, and Wear Technology, ASM Handbook, Vol. 18, ASM International, 1992, 483. [20] Summers-Smith, J.D., A Tribology Casebook, Mechanical Engineering Publishing Ltd., Bury St. Edmunds, 1997, pp. 5–7, 91–92, 110–113. [21] API Standard 670, Machinery Protection Systems, 4th ed., American Petroleum Institute, Washington, DC, 2000. [22] Jackson, C., How to prevent turbomachinery thrust failures, in Compressor Handbook for the Hydrocarbon Processing Industry, Gulf Publishing Company, Houston, TX, 1975, 152. [23] Glacier Rotating Plant Bearings, Glacier Designers’ Handbook No. 5B: Tilting Pad Thrust Bearings, The Glacier Metal Co. Ltd., 1989. [24] Glacier Rotating Plant Bearings, Glacier Designers’ Handbook No. 10: Standard Tilting Pad Journal Bearings, The Glacier Metal Co. Ltd., 1989. [25] Glacier Rotating Plant Bearings, Glacier Designers’ Handbook No. 14: Medium and Thick Wall Journal Bearings, The Glacier Metal Co. Ltd., 1989. [26] Snyder, D.R., Selecting rolling element bearings for modern applications: how performance criteria, operating environments, materials, lubrication requirements and other factors shape the bearingselection process, Tribology and Lubrication Technology, 60, 28, 2004. [27] Cantley, R.E., The effect of water in lubricating oil on bearing fatigure life, ASLE Transactions, 20, 244, 1977. [28] Kure-Jensen, J., Large steam turbine-generators, in CRC Handbook of Lubrication (Theory and Practice of Tribology) Volume I: Practice, Booser, E.R., Ed., CRC Press, Inc., Boca Raton, FL, 1983, 91. [29] Bloch, H.P., A Practical Guide to Steam Turbine Technology, McGraw-Hill, New York, 1996, 136. [30] Ogle, A.W. and Render, M., Impact of market changes upon power plant control, in IMechE Seminar Publication: Steam Turbine Governing and Overspeed Protection, Professional Engineering Publishing Ltd., London, 1998, 5. [31] Hayler, M.G. and Wilson, A.C.M., Lubrication of water and steam turbines, in Lubrication in Practice, 2nd ed., Robertson, W.L., Ed., Marcel Dekker, Inc., New York, 1984, chap. 4. [32] Governing Fundamentals, Manual 25195, Woodward Governing Company, 1999, chap. 7. [33] Clark, E.E., Protecting against rotating equipment loss, Chemical Engineering, 104, 106, 1997. [34] Wilson, A.C.M., Problems encountered with turbine lubricants and associated systems, Lubrication Engineering, 32, 59, 1976. [35] Neale, M.J., Drives And Seals, A Tribology Handbook, Society of Automotive Engineers, Inc., Butterworth-Heinemann, Ltd., Oxford, 1994, 29. [36] Crease, A.B., Design principles and lubrication of gear couplings, International Conference on Flexible Couplings for High Powers and Speeds, Michael Neale and Associates Ltd., 1977, Paper B1. [37] Calistrat, M.M., Extend gear coupling life — Part I, in Compressor Handbook for the Hydrocarbon Processing Industry, Gulf Publishing Company, Houston, TX, 1975, 161. [38] Wright, J., Principle engineering features required by high performance gear couplings, International Conference on Flexible Couplings for High Powers and Speeds, Michael Neale and Associates Ltd., 1977, Paper B4. [39] Calistrat, M.M., Shaft-coupling lubrication, Lubrication Engineering, 37, 9, 1980. [40] Calistrat, M.M., Mechanical shaft couplings, in CRC Handbook of Lubrication (Theory and Practice of Tribology) Volume II: Theory and Design, Booser, E.R., Ed., CRC Press, Inc., Boca Raton, FL, 565. [41] Calistrat, M.M., Extend gear coupling life – Part II, in Compressor Handbook for the Hydrocarbon Processing Industry, Gulf Publishing Company, Houston, TX, 1975, 166. [42] Carter, D., Garvey, M., and Corcoran, J.P., The baffling and temperature prediction of coupling enclosures, Proceedings of the 23rd Turbomachinery Symposium, Turbomachinery Laboratory, Texas A&M University, College Station, TX, 1999, 115.


8-36


[43] Errichello, R., Friction, lubrication, and wear of gears, in Friction, Lubrication, and Wear Technology, ASM Handbook, Vol. 18, ASM International, 1992, 541. [44] Bloch, H.P., A Practical Guide to Compressor Technology, McGraw-Hill, New York, 1996, 372. [45] Transamerica Delaval Engineering Handbook, 4th ed., Welch, H.J., Ed., McGraw-Hill Book Company, New York, 1983, chap. 5. [46] Low, M.B.J., The sliding of high temperature components in steam turbines, Proceedings of IMechE, 194, 171, 1980. [47] Hackel, R.A. and Keyes, H.M., Steam turbines in process industries, in Sawyer’s Turbomachinery Maintenance Handbook, Volume 2: Steam Turbines/Power Recovery Turbines, Sawyer, J., Ed., Turbomachinery Intl. Pubns., Norwalk, 1980, chap. 4. [48] Steam turbines and auxiliaries, Power, 133, June, 1989. [49] Pilicy, F.X. and Dundas, R.E., Insurance for steam turbines, in Sawyer’s Turbomachinery Maintenance Handbook, Volume 2: Steam Turbines/Power Recovery Turbines, Sawyer, J., Ed., Turbomachinery Intl. Pubns., Norwalk, 1980, chap. 10. [50] Westhofen, B., Enhancing the availability of industrial turbines, Brown Boveri Review, 73, 31, 1986. [51] Cameron, J.A., Materials for use in geothermal steam turbines, in Geothermal ’77, Carrier Corporation, 1977, 39. [52] Ortolano, R.J., Steam turbine blading maintenance: part IV, Turbomachinery International, 24, 56, 1983. [53] D’Innocenzio, M., Oil systems – design for reliability, Proceedings of the First Turbomachinery Symposium, Turbomachinery Laboratory, Texas A&M University, College Station, TX, 1970, 163. [54] Ventilation of steam turbines, Gargoyle Lubricants Technical Bulletin, 1945. [55] Sassos, M.J., Babyak, M.R., and Zerbe, J.P., Lubrication and seal oil systems — common problems and practical solutions, Proceedings of the 22nd Turbomachinery Symposium, Turbomachinery Laboratory, Texas A&M University, College Station, TX, 1993, 51. [56] Salisbury, R.J., Stack, R., and Sassos, M.J., Lubrication and seal oil systems, Proceedings of the 13th Turbomachinery Symposium, Turbomachinery Laboratory, Texas A&M University, College Station, TX, 1984, 151. [57] API Standard 614, Lubrication, Shaft-Sealing, and Control-Oil Systems and Auxiliaries for Petroleum, Chemical and Gas Industry Services, 4th ed., American Petroleum Institute, Washington, DC, 1999. [58] API Recommended Practice 686, Recommended Practice for Machinery Installation and Installation Design, 1st ed., American Petroleum Institute, Washington, DC, 1996. [59] Petroleum Products, Lubricants, and Fossil Fuels, Annual Book of ASTM Standards, Section 5, ASTM International, West Conshochocken, PA, 2003. [60] Schwager, B.P., Hardy, B.J., and Aguilar, G.A., Improved response of turbine oils based on Group II hydrocracked base oils compared with those based on solvent refined base oils, in Turbine Lubrication in the 21st Century, ASTM STP 1407, Herguth, W.R. and Warne, T.M., Eds, American Society for Testing and Materials, West Conshohocken, PA, 2001, 71. [61] Hannon, J.B., How to select and service a turbine oil, Machinery Lubrication, July, 2001. [62] Potential performance problems caused by mixing different types of turbine oils, Engineering and Construction Bulletin 2003-17, US Army Corps of Engineers, 2003. [63] Dolby, G.W. and Kofke, W.A., The role of base oils and additives in modern-day turbine oils, in Symposium on Turbine Oils, ASTM STP 321, American Society for Testing and Materials, Philadelphia, PA, 1962, 1. [64] O’Brien, J.A., Lubricating oil additives, in CRC Handbook of Lubrication (Theory and Practice of Tribology), Volume II: Theory and Design, Booser, E.R., Ed., CRC Press, Inc., Boca Raton, FL, 1984, 301. [65] von Fuchs, G.H., Rust inhibitors — their evaluation and performance, in Symposium on Turbine Oils, ASTM STP 321, American Society for Testing and Materials, Philadelphia, PA, 1962, 28. [66] vd Merwe, D.G., Morgan, P.M., Roets, P.N.J., and Botha, J.J., Additives in turbine and dynamic compressor oils, The South African Mechanical Engineer, 47, 13, 1997.


Steam Turbines

8-37

[67] Swift, S.T., Butler, K.D., and Dewald, W., Turbine oil quality and field application requirements, in Turbine Lubrication in the 21st Century, Herguth, W.R. and Warne, T.M., Eds, ASTM STP1407, American Society for Testing and Materials, West Conshochocken, PA, 2001, 25. [68] Roberton, R.S., Background and development of ASTM D 4378: practice for in-service monitoring of mineral turbine oils for steam and gas turbines, in Turbine Oil Monitoring, ASTM STP 1021, Young, W.D. and Roberton, R.S., Eds, American Society for Testing and Materials, Philadelphia, PA, 1989, 3. [69] The Lubrication of Steam Turbines, Shell, Alabaster Passmore Sons, Ltd., London, 1958. [70] Summers-Smith, D., The unacceptable face of lubricating oil additives, Tribology, 11, 318, 1978. [71] Engineering and Design — Lubricants and Hydraulic Fluids, Engineer Manual EM 1110-2-1424, Department of the Army, Washington, DC, 1999, chap. 12. [72] Furby, N.W., Hanly, F.J., and Vincent, J.A., Rusting in turbine oil systems, in Symposium on Steam Turbine Oils, ASTM STP 211, American Society for Testing and Materials, Philadelphia, PA, 1956, 40. [73] Layne, R.P., Vapor space corrosion inhibition of steam turbine lubricating and cleaning oils, in Turbine Lubricating Problems, ASTM STP 437, American Society for Testing and Materials, Philadelphia, PA, 1968, 73. [74] Steam Turbines and their Lubrication, Mobil Oil Corporation, New York, NY, 1981. [75] Enz, W.E. and Hausermann, A., Particular problems of steam turbine lubrication, Proceedings of the Seventh Turbomachinery Symposium, Turbomachinery Laboratory, Texas A&M University, College Station, TX, 1981, 125. [76] Steam Turbine Lubrication, 2nd ed., The Texas Company, 1947. [77] Fowle, T.I., Problems in the lubrication systems of turbomachinery, Proceedings of the Instrumental Mechanical Engineers, 186, 705, 1972. [78] Bice, C.D., Air entrainment issues in equipment using new generation turbine oils, Presented at STLE, Pittsburgh Section, November, 2004. [79] Fitch, J.C. and Jaggernauth, S., Moisture — the second most destructive lubricant contaminate, and its effects on bearing life, P/PM Magazine, December, 1994. [80] Coleman, W.L., Water contamination of steam turbine lube oils — how to avoid it, Proceedings of the Seventeenth Turbomachinery Symposium, 51. [81] Missana, A. and Steenburgh, J.H., Ensuring clean lube oil for large steam turbines, Power Engineering, 88, 46, June, 1984. [82] Bissett, W., Cost effective condition monitoring of large steam turbine/generator oil systems, Transactions of Mechanical Engineering, IEAust., ME20, 61–68, 1995. [83] Abner, Jr., E., Lubricant deterioration in service, in CRC Handbook of Lubrication (Theory and Practice of Tribology) Volume I: Practice, Booser, E.R., Ed., CRC Press, Inc., Boca Raton, FL, 1983, 517. [84] Lamping, G.A., Cuellar, Jr., J.P., and Silvus, H.S., Summary of maintenance practices for large steam turbine-generator lubrication systems, ASME/IEEE Power Generation Conference, ASME Paper 86-JPGC-Pwr-14, 1986. [85] DenHerder, M.J. and Vienna, P.C., Control of turbine oil degradation during use, Lubrication Engineering, 37, 67, 1981. [86] McCloskey, T.H., Troubleshooting bearing and lube oil system problems, Proceedings of the 24th Turbomachinery Symposium, Turbomachinery Laboratory, Texas A&M University, College Station, TX, 1998, 147. [87] Ohgake, Ryoji, Sunami, M., Yoshida, T., and Watanabe, J., The reliable control of oil quality in Japanese turbine units, in Turbine Oil Monitoring, Young, W.C. and Roberton, R.S., Eds, ASTM STP 1021, American Society for Testing and Materials, Philadelphia, PA, 1989, 35. [88] Bloch, H.P., Criteria for water removal from mechanical drive steam turbine lube oils, Lubrication Engineering, 36, 699, 1980.


8-38


[89] Roberton, R.S., ASTM in-service monitoring program for steam and gas turbine oils, Lubrication Engineering, 42, 466, 1986. [90] Micetic, J., New Generation of turbine oils, US Army Corps of Engineers, Hydroelectric Design Center, Portland District. [91] Adams, M.A. and Bloch, H.P., Vacuum distillation methods for lube oils increase turbomachinery reliability, Proceedings of the 17th Turbomachinery Symposium, Turbomachinery Laboratory, Texas A&M University, College Station, TX, 1991, 41. [92] Stein, W.H. and Bowden, R.W., Shell Global Solutions, Inc., Turbine oil reclamation and refortification, Machinery Lubrication Magazine, July, 2004. [93] Phillips, W.D., The use of a fire-resistant turbine lubricant: Europe looks to the future, in Turbine Lubrication in the 21st Century, Herguth, W.R. and Warne, T.M., Eds, ASTM STP1407, 2001, 1. [94] Brown, K., Utility Service Associates, Condition-monitoring of phosphate ester hydraulic fluids, Machinery Lubrication Magazine, November, 2002. [95] Brown, K.J. and Staniewski, J.W., Condition monitoring and maintenance of steam turbinegenerator fire resistant triaryl phosphate control fluids, in Condition Monitoring and Preventive Proceedings, STLE SP-27, Society of Tribologists and Lubrication Engineers, 1989, 91. [96] Wolfe, G.F., Experience with phosphate ester fluids as industrial steam turbine-generator lubricants, Lubrication Engineering, 8, 413, 1978. [97] Phillips, W.D., The conditioning of phosphate-ester fluids in turbine applications, Lubrication Engineering, 12, 766, 1983. [98] Thibault, R., Converting to condition-based oil changes — part II, Practicing Oil Analysis Magazine, March, 2001. [99] Steele, F.M., Filtration and reclamation of turbine oils, Lubrication Engineering, 34, 252, 1978. [100] Steele, F.M., Contamination control in turbomachinery oil systems, Lubrication Engineering, 40, 487, 1984.


Steam Turbine

Recommend Documents