Turbine plant systems

British Electricity International, in Turbines, Generators and Associated Plant (Third Edition), 1991

6.2.4 Jacking-oil pumps and priming pumps

The jacking oil pumps deliver oil at around 300 bar to the individual bearings. The pumps used are motor-driven positive displacement and either multi-plunger pumps or two-shaft gear pumps, see Figs 2.57 and 2.58.

FIG. 2.57. Multi-plunger jacking oil pump

FIG. 2.58. Gear type jacking oil/priming pump

The arrangement utilises either one gear pump per bearing or one motor/pump unit for either one or two bearings. The pumps require a positive suction pressure. The multi-plunger pumps are fed from the lubricating oil manifold and the gear pumps have a motor-driven positive displacement two-shaft gear pump.

On some designs the turning gear is provided with a separate jacking oil pump. This pump is a motor-driven, positive displacement, two-shaft gear pump; the discharge pressure is the same as for the other jacking oil pumps.

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Condenser Flood and Vacuum Tightness Tests

Dipak K. Sarkar , in Thermal Power Plant, 2017

14.6.2 Condenser Vacuum Test

i.

Start the turbine auxiliary/lube oil pump;

ii.

Start the jacking oil pump (if provided);

iii.

Put the turbine on turning gear;

iv.

Start one CEP;

v.

Start flow through the CEP minimum flow recirculation line;

vi.

Line up turbine gland sealing steam supply and dump system;

vii.

Start gland steam condenser with steam exhauster;

viii.

Start both vacuum pumps or noncondensing type single-stage starting (hogging) air ejector;

ix.

On achieving about a 20   kPa vacuum, supply steam to turbine gland seals;

x.

Maintain pressure and temperature on the turbine gland steam supply header at 102.3   kPa and >   433   K, respectively;

xi.

Verify whether the desired vacuum, as recommended by the turbine manufacturer, is maintained;

xii.

In case the desired vacuum cannot be achieved, check all air ingress points, eg, glands of valves located on interconnecting piping with the condenser, vacuum-breaking valve, all atmospheric vent valves on the shell side (steam side) of the condenser, feedwater heater vents and drains connections, steam drains connections, makeup water connections, condensate spray connections, LP bypass dump line connections, air evacuation connections, and so on;

xiii.

Identify the defective point/s and rectify the defect/s;

xiv.

Repeat steps xii and xiii until the desired vacuum is achieved;

xv.

Stop one vacuum pump or change over from the hogging air ejector to the steam-jet (holding) air ejector;

xvi.

Verify whether the desired vacuum can be sustained;

xvii.

If step xvi fails, repeat steps xii and xiii until desired vacuum is sustained.

Note

In a running unit, air leakages to the condenser may be detected by adopting the "Helium leak testing" technique, which is the fastest and most-effective method for identifying condenserleaks. This method, however, requires a highly skilled operator for effective results.

An alternative technique to the above is the "ultrasound detection" technique. Although this technique offers favorable results with less-skilled technicians, noise emanating from steam leakages may interfere with the end result.

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The steam turbine

British Electricity International, in Turbines, Generators and Associated Plant (Third Edition), 1991

8.5.2 Electrical turning gear (ETG)

The rotors are turned slowly, typically less than 30 r/min, during start-up and shutdown by the ETG. An electric drive motor turns the rotors through a wormshaft and wormwheel, thereby providing a reduction gear. A jacking oil pump supplies high pressure lubricating oil to the reduction gear.

Manual control of the motor is provided in the Control Room, and automatic stop and start facilities are also included in the motor switchgear. The motor overload trip is set at a value which prevents excessive torque being applied to a seized rotor. Electrical interlocking prevents the motor being started until jacking oil pressure is established.

A self-shifting synchronous (SSS) clutch is installed between the drive motor and the turbine shaft and provides a simple mechanical means of automatically connecting or disconnecting the turning gear drive. The SSS clutch is a positive tooth-type overrunning clutch which is self-engaging when passing through synchronism, that is, immediately the speed of the input shaft exceeds that of the output shaft. The clutch disengages automatically when the torque reverses, that is, when the speed of the output shaft exceeds that of the input shaft (Fig 1.127).

FIG. 1.127. Self-shifting synchronous clutch

At standstill, when the driving shaft begins to provide torque, the clutch will engage; if after this the torque ceases, the clutch will disengage. It will reengage if the speed of the driving shaft exceeds that of the driven shaft, whether at full barring speed or at any lesser speed.

Spring-loaded pawls, acting on a ratchet, sense the relative speeds of the input and output shafts; when the input shaft is about to overtake the other, the pawls 'bite' and reactive torque is applied to the helically-splined sleeve which moves axially and slides the clutch teeth into engagement.

The positions of the pawls and ratchet teeth ensure that the clutch teeth pass between each other exactly, without making contact until full engagement is reached; at this point the pawls leave the ratchet teeth, and the flanks of the clutch teeth meet to take up the drive.

It is important to note that the pawls merely sense zero relative speed and angular location, they do not carry the main torque. To prevent ratcheting and consequent wear when the relative speed is high, the pawls are designed to disengage centrifugally.

The only load imposed on the pawls is the force required to engage the clutch. In a very large clutch, this could overload the pawls and so a relay clutch is used. Here the primary mechanism is exactly as described above, but the helically-splined sleeve, in moving forward, engages teeth to move a much heavier helically-splined sleeve on which the clutch teeth are cut. A further refinement is an oil dashpot, which cushions the engaging action, and prevents disengagement as a result of rapid transient torque reversal.

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Start-Up and Shut-Down

Dipak K. Sarkar , in Thermal Power Plant, 2015

12.7 Emergency Shut-Down of Steam Turbine

An emergency shut-down of a steam turbine may become imminent due to variety of off-normal conditions either inside the steam turbine, such as unusual noise, etc., or in its associated equipment/systems, e.g., quality of stator cooling water deviating from its recommended limit.

1.

Manually operate emergency trip push buttons

2.

Ensure the ESVs, IVs, and control valves of both the HP and IP turbines get closed

3.

Ensure the extraction steam line block valves are closed

4.

Ensure the non-return valves at HP turbine exhaust are closed

5.

Isolate the generator and switch off the field breaker

6.

Make sure the HP-LP bypass valves have opened

7.

Start the auxiliary oil pump and ensure the oil pressure in lube oil system is normal

8.

Maintain the lube oil temperature

9.

Start the jacking oil pump

10.

When the rotor speed falls to a predetermined low value, put the rotor on turning gear manually or turning gear may cut-in automatically

11.

Break the condenser vacuum

12.

Shut down the boiler and keep it boxed and bottled up

13.

Close HP-LP bypass valves

14.

Following the manufacturer's guideline open the drain valves as follows:

i.

HP turbine casing

ii.

Before-seat and after-seat drain valves of non-return valves on turbine extraction steam lines

iii.

Warm up drain valves for emergency stop valve/s and control valves

iv.

Drain valves on main steam piping

v.

Drain valves on cold and hot reheat piping.

15.

When the vacuum falls to zero, stop the turbine gland seal steam supply

16.

Stop the condensate extraction pump

17.

The CW pump may be stopped when the temperature of the exhaust part of the LP turbine falls to 328   K or to a value recommended by the manufacturer

18.

The rotor should remain on the turning gear until the temperature of the HP casing falls below a recommended limit

19.

Stop the jacking oil pump

20.

As the unit cools down and water shrinks, add make-up water intermittently to the boiler to prevent the drum level from dropping below the visibility limit of the drum level gauge glass

21.

Do not restart the boiler until identification and rectification of defects of turbine.

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Turbo Generator Control System

Swapan Basu , Ajay Kumar Debnath , in Power Plant Instrumentation and Control Handbook, 2015

1.1 Subsystems or Functional Subgroups

As indicated above, the subgroups are functionally divided to achieve better control over the entire turbine system. The main items particular subgroups control are discussed in the following clauses.

1.1.1 Turbine HP Control Fluid Subsystem

Whether combined or separate, subsystems have different standby liquid control systems (SLCs) and drives for LPBP and main valve systems to serve their assigned purposes, which are to supply HP power/control oil. This is done by using control oil pumps, recirculation pumps, cooling fans, pumps for regeneration circuits, etc.

1.1.2 Turbine Oil Subsystem

This subsystem takes care of the oil systems required for initial running of the turbine and shut-down procedure and includes SLCs and drives of auxiliary and emergency lube oil pumps, jacking oil pumps, turning gear gate valves, etc.

1.1.3 Condenser Evacuation and Gland Seal Subsystem

This subsystem includes SLCs for condensate extraction pumps and turbine drains as well (although it is a part of the turbine control subgroup) and drives for vacuum pumps, air extraction valves, instrument/service air solenoid valves, vacuum breaker solenoid valves, etc.

1.1.4 Turbine Control Subsystem (Start-Up, Loading, Unloading, and Shut Down)

The heart of the turbo generator control system is the turbine control subsystem (start-up, loading, unloading, and shut down) comprised of SLCs for drains, warm-up controller for generating the criteria for operation of drain valves, soaking of turbine metals, and holding/speeding of the turbine rotor. Also included are the start-up device (TSL), speed and load set point devices of the turbine governing system, and the auto-synchronizer.

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Plant P&ID (Process) Discussions

Swapan Basu , Ajay Kumar Debnath , in Power Plant Instrumentation and Control Handbook (Second Edition), 2019

11.1.1.2 Turbine Jacking Oil System (Fig. 3.45)

1.

During start-up and shutdown of the turbine, the jacking oil system comes into operation. Before admitting steam, the turbine is turned at a slow speed by turning gear. During the start up of the turbine, jacking oil is necessary. Also, when the turbine is shutting down, then the jacking oil system comes into operation so that the hot shaft does not rest on the white metal of the bearing. As discussed earlier, there is a need to have a separation between the white metal in the bearing and shaft with the help of oil film. At low speed, say <   75 rpm or so, this separation is achieved by jacking oil. During this time, oil at a very high pressure, say >   200 kg/cm2 (> 350 for higher MW units), is injected from the bottom to lift the shaft. The oil is kept injected until an oil film is developed.

Fig. 3.45

Fig. 3.45. Jacking oil scheme.
2.

Separate jacking oil pumps (JOPs) taking the suction from the main oil reservoir are deployed. In many places, these pumps are kept at the bearing pedestal. On account of the very high pressure (sometimes >   350 kg/cm2), there is a high importance of safety relief valves for these pumps.

3.

Pressure and temperature at the jacking oil header are monitored. The normal range may be 0–350 kg/cm2 for a medium to high MW power plant. At the main reservoir the level is monitored at the DCS with the help of a level transmitter. A level switch is used for alarm and interlock.

4.

As discussed earlier, jacking oil is used during low running of the turbine. During start-up and shutdown, the turbine rotor is rotated at a slow speed, ~   100 rpm, by the turning gear. Basically, the turning gear is driven by a motor, and is coupled to the main turbine through a series of reduction gear. Turning gear is used to rotate the turbine rotor slowly during start-up and shutdown of the turbine to avoid uneven heating/cooling (to avoid shaft deformation). This is turning gear operation. During this time, the turning gear oil pump may be used to supply oil to the bearings.

5.

When the turbine speed crosses, say 200 rpm, the turning gear stops operating. Jacking oil only operates during the time the turning gear is in operation.

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Introduction of advanced technologies for steam turbine bearings

P. Pennacchi , in Advances in Steam Turbines for Modern Power Plants, 2017

15.5 Bearing coating materials

Obviously, the material used for coating the bearing is essential for a correct and long-term operation, but it is worthwhile to also spend a few words about the main structural parts of the bearings used for steam turbines.

Usually, the shell of the bearing is made of structural steel, possibly with high strength and good machining/welding properties like S355 J0 [30], corresponding to ASTM A572   Gr 50 [31], C22+n [32] or C45 [33], of cast iron EN-GJL-250/350 [34] or equivalents.

Pads are often made of the same material as the shell (see Fig. 15.59), with sometimes inserts in CuCr1Zr alloy [35], see Fig. 15.8.

Figure 15.59. Different types of pads for journal and thrust bearings.

 Source: Courtesy of Eurobearings Srl.

Pivots are subject to high concentrated loads and are made of low alloy martensitic chrome steel, heat-treated, like 100Cr6 [36], equivalent to AISI 52100, thanks to its high hardness, wear resistance and to the suitability to cold deformation machining.

15.5.1 Low friction alloys (white metals)

The components that may potentially come into contact with the rotor are coated with a material with a low friction coefficient and characterized by a lower hardness than the material of the shaft (usually steel), in both journal and thrust bearings.

This is due to two main reasons:

1.

During the start-up of the turbine, or its run-down, when the sliding speed between the bearing surfaces and those of the rotor is low and full film hydrodynamic lubrication might not form, the possible contact between the surfaces results in wear of the less hard material (which acts as sacrificial element). In this way, the rotor shaft is protected, while the components of the bearings coated by a material with low friction coefficient are, actually, considered parts subject to wear and eventually spare parts. Obviously, the presence of jacking oil pumps in the journal bearings greatly limits the possibility of contact.

2.

At the start-up of the turbine, in the case of boundary lubrication, low friction coefficient allows limiting the breakaway torque required to start and consequently also the wear of all the components.

In steam turbines, the coating material used is, generally, the so-called "babbitt metal," a term which, actually, refers to a wide series of alloys. Isaac Babbitt invented the first type of these alloys in 1839. Since then, he has become the eponym of these alloys. The standard ASTM B23-00 [37] prefers to indicate these alloys as "white metals," identifying the term babbitt metal as the trade name. Alloys based on tin are normally used for the turbines. These alloys differentiate for the percentages of other elements, mainly antimony and copper, and a nonnegligible percentage of lead is present. The standard classifies the different alloys with a number (the "grade"), which is associated with a specific chemical composition. The white metals included in ASTM B23-00 are characterized by excellent antiseizure characteristics, corrosion resistance, conformability, embedability and resistance to galling. These white metals are generally not very resistant to fatigue (Fig. 15.60), or to creep, and have a rather low melting temperature. Equally low is the hardness, from which it follows that the Young's modulus is relatively low, about 50   GPa, less than a fourth of the Young's modulus of steel. The maximum yield point is 45.5   MPa at 20°C, which approximately halves at 100°C. However, the yield point depends strongly on the alloy grade. Due to this, the average specific load (pressure) on the bearing is generally limited between less than 2 and 3   MPa, also in order to avoid phenomena of wiping due to overloading that could lead to coating irreparable damage (see Fig. 15.27).

Figure 15.60. Application of dye penetrant to unveil fatigue cracks in the babbitt metal coating of a tilting pad journal bearing.

Nowadays, the power-generation applications prefer bearing coatings made with different types of white metals, namely ECKA TEGO V738 for heavy duty and more recently ECKA Tegostar, whose chemical composition differs from the ASTM B23-00 and becomes an industrial standard de facto, because of their better performances in terms of resistance to creep [38]. Tegostar has also the remarkable advantage that the alloy is without lead, cadmium, nickel, and arsenic.

Babbit metal is provided in pigs or bars (Fig. 15.61). The coating of bearing components is made in special heated centrifuges.

Figure 15.61. Pigs and bars of white metal.

 Source: Courtesy of Eurobearings Srl.

15.5.2 Polymeric materials

Recently, polymeric materials, which are characterized by low friction coefficients, have been taken into consideration to replace the white metal [39–41], specifically in an "environmentally friendly" vision, in order to totally eliminate the coatings that have lead as an alloying element.

An almost "natural" choice would be the polytetrafluoroethylene (PTFE) (C2F4)n, which gathers a low friction coefficient with the steel to the other positive characteristics of the white metals, and is recognized as "environmentally friendly." Also, it adheres well to steel and would easily make the coating of the bearing components that may come into contact with the rotors.

As an example, some TEHD simulations of the same pad of a thrust bearing [42,43], but with different coating layer materials (PTFE and babbitt metal), are presented for a nominal load of 2   MPa, ISO VG46 oil and rotational speed of 1500   rpm (see Fig. 15.62).

Figure 15.62. Pad layout in the x–y plane and parameters for the pad position definition.

The pressure field acting on the pad for a babbitt metal- (left) and a PTFE- (right) coating material is shown in Fig. 15.63 for the rotational speed equal to 1500   rpm. The two pressure distributions are quite similar: the maximum pressure value corresponds with the mean pad radius and is close to the trailing edge. However, the pressure on the complete pad surface is slightly lower for PTFE than for babbitt metal coating and the area with maximum pressure (dark-red (gray in print versions) area) is smaller for the polymeric material.

Figure 15.63. Pressure field (in MPa) for babbitt metal (left) and polytetrafluoroethylene (PTFE) (right) coatings for a rotational speed of 1500   rpm, nominal load of 2   MPa, and ISO VG46 lubricant.

Concerning the oil-film thickness, thicker oil-film layer is obtained for babbitt metal (Fig. 15.64, left) especially corresponding with the leading edge. This result means that, in order to obtain the same load capacity, a thinner oil-film is required in the case of PTFE coating. Therefore, the runner is closer to the pad for the polymeric material and the nozzles supply a lower quantity of lubricant.

Figure 15.64. Oil-film thickness (in μm) for Babbitt (left) and polytetrafluoroethylene (PTFE) (right) coatings for a rotational speed of 1500   rpm, nominal load of 2   MPa and ISO VG46 lubricant.

The temperature distributions (Fig. 15.65) are similar for the two cases and the maximum temperature is reached in correspondence with the outer trailing edge of the pad. However, the lubricant temperature values are quite different: whereas the hot area for the Babbitt coating does not exceed 73°C, the analogous area for the PTFE layer is characterized by a higher temperature (79°C). The heating of the lubricant for polymeric coating is due to the lower PTFE thermal conductivity coefficient with respect to the Babbitt one (0.24   W/mK for PTFE and 55   W/mK for Babbitt). Indeed, due to its low thermal conductivity, the PTFE works like a thermal wall avoiding the absorption, by the pad, of the heat generated in the lubricant by viscous stresses.

Figure 15.65. Oil-film temperature (in °C) for Babbitt (A) and polytetrafluoroethylene (PTFE) (B) coatings for a rotational speed of 1500   rpm, nominal load of 2   MPa and ISO VG46 lubricant.

Similar results to those shown in Figs. 15.63–15.65 have been obtained for the rotational speed equal to 3000   rpm and are not reported for brevity. The lower thermal conductivity along with the lower Young's modulus of the polymeric material, involves greater deformations in the coating layer, besides the increase of the lubricant temperature. The joined interaction between strain and temperature could lead to the creep phenomenon, which involves the movement of the coating material from the higher-pressure zones to the lower-pressure zones. The modification of the coating layer geometry as consequence of creep implies a variation on the pressure, temperature, and lubricant thickness fields and therefore on pad strains. In other words, the coating layer creep modifies the behavior of the whole pad. Obviously, the creep continues also in the new achieved conditions and the coating layer material moves depending on the pressure–temperature distributions. Therefore, the polymeric coating layer is characterized by a continuous movement that leads to a pad distortion known as "dishing" or "crowning" [4,44–46]. The results obtained with the model show good agreement with the creep and the crowning problem of the PTFE coating materials, which eventually make the PTFE unsuitable for bearings with long operational life, such as those of turbines.

As it is possible to see in Table 15.4, the rotation of the pad around the x and y directions ( α and β angles, respectively) is greater for higher rotational speeds. Moreover, for the same operating conditions, the babbitt metal shows higher values of rotations. Coherently with the lubricant thickness distribution previously shown, higher z f values are obtained for higher rotational speeds and PTFE coating materials.

Table 15.4. Geometrical parameters representing the pad equilibrium configuration and power loss depending on rotational speeds and coating materials

Babbitt PTFE
1500   rpm 3000   rpm 1500   rpm 3000   rpm
α (°) −0.0129 −0.0165 −0.0125 −0.0139
β (°) 0.0293 0.0375 0.0245 0.0286
z f (m) 0.0294 0.0294 0.0294 0.0294
Power loss (kW) 2.687 6.3156 2.7621 6.387

PTFE, polytetrafluoroethylene.

The power loss (see Table 15.4) is slightly greater for PTFE coating than the babbitt metal, but the difference between the two materials is about 2.8% and 1.1% for 1500   rpm and 3000   rpm rotational speeds, respectively.

The polymeric material, which has instead been applied for the coating of bearings is polyether ether ketone (PEEK). However, contrary to PTFE which is used unmixed, the PEEK is frequently mixed with some additives, such as carbon fibers or nanofibers, or nanoparticles of other materials such as CuS, Si3N4, SiO2, SiC, and ZrO2, in order to improve mechanical and/or thermal characteristics. A scanning electron microscope image of a PEEK with Si3N4 nanoparticles is shown in Fig. 15.66.

Figure 15.66. Scanning electron microscope image of a polyether ether ketone (PEEK) with Si3N4 nanoparticles.

 Source: Courtesy of Eurobearings Srl.

In addition, the PEEK being semicrystalline, it can have different degrees of crystallinity, and glass and melting temperatures. The characterization of the type of PEEK is usually made by means of a tribometer, with the pin-on-disk method in accordance with ASTM G99 [47]. By way of example, Fig. 15.67 shows the trace of wear at the end of a dry test on a sample of PEEK. Figs. 15.68 and 15.69 show, for the same sample, the trend of friction coefficient in the dry and in the presence of ISO VG 32 and the profile of the track wear in the presence of ISO VG 32.

Figure 15.67. Trace of wear at the end of pin-on-disc dry test on a sample of polyether ether ketone (PEEK).

 Source: Courtesy of Eurobearings Srl.

Figure 15.68. Trend of the friction coefficient during the pin-on-disk test.

 Source: Courtesy of Eurobearings Srl.

Figure 15.69. Wear track profile at the end of the pin-on-disk test.

 Source: Courtesy of Eurobearings Srl.

The comparison of the performances of a thrust bearing coated with PTFE and PEEK with additives is shown hereafter, by using a TEHD model [43]. The pivoted pad dimensions are reported in Table 15.5.

Table 15.5. Dimensions of the thrust bearing pad

Characteristic Value
External diameter (mm) 920
Internal diameter (mm) 720
Number of pad 16
Angle of the pad (°) 15
Pivot offset 60%
Pad thickness (mm) 28.5
Pad material Steel
Pivot diameter (mm) 14

In particular, the effect of the thickness of two different coatings has been investigated. PTFE and PEEK coatings with thicknesses in the range from 0.6   mm to 1   mm have been analyzed for a rotational speed of 1000   rpm and assuming the same oil-film thickness in correspondence with the pivot for all the simulations. This assumption allows the variation of the load to be easily evaluated for variation of the material and the coating thickness. The properties of the two commercial materials used in the simulations of this section are reported in Table 15.6.

Table 15.6. Mechanical characteristics of polytetrafluoroethylene and polyether ether ketone materials

PTFE PEEK
Young's modulus (GPa) 0.5 3.6
Poisson's coefficient 0.46 0.40
Thermal conductivity (W/(m·K)) 0.24 0.95
Density (kg/m3) 2200 1400
Linear thermal expansion coefficient (1/K) 170×10−6 5×10−6
Specific heat at constant pressure (J/(kg·K)) 1250 1800

PTFE, polytetrafluoroethylene; PEEK, polyether ether ketone.

The results of the analyses for the PTFE and the PEEK coatings and for different values of the coating thickness are reported in Tables 15.7 and 15.8, and Figs. 15.70 and 15.71.

Table 15.7. Results of simulations for the polytetrafluoroethylene coating

PTFE coating thickness (mm) 0.6 0.8 0.9 1.0
Angle α (°) −0.0072 −0.0072 −0.0071 −0.0072
Angle β (°) 0.0300 0.0265 0.0255 0.0250
Pad load (kN) 40.7 46.2 49.1 52.2
Specific pressure (MPa) 3.78 4.30 4.57 4.86
Maximum pressure (MPa) 11.39 12.96 13.71 14.32
Maximum temperature on coating (°C) 113.6 114.2 114.8 114.6
Total bearing power loss (kW) 74.1 79.7 82.7 85.6

PTFE, polytetrafluoroethylene.

Table 15.8. Results of simulations for the polyether ether ketone coating

PEEK coating thickness (mm) 0.6 0.8 0.9 1.0
Angle α (°) −0.0091 −0.0091 −0.0091 −0.0090
Angle β (°) 0.0440 0.0435 0.0430 0.0427
Pad load (kN) 28.2 27.9 27.8 27.7
Specific pressure (MPa) 2.63 2.61 2.59 2.58
Maximum pressure (MPa) 7.97 7.82 7.75 7.68
Maximum temperature on coating (°C) 101.9 101.0 103.7 102.1
Total bearing power loss (kW) 61.5 61.7 61.7 61.8

PEEK, polyether ether ketone.

Figure 15.70. Oil-film pressure distribution for the pad.

Figure 15.71. Coating temperature distribution for the pad.

Considering the PTFE coating whose numerical simulation are reported in Table 15.7, the increase of the coating thickness leads to an increase of the specific and maximum pressure that obviously corresponds to an increase of the total pad load and to an increase of the total bearing power loss. The maximum temperature on the coating remains constant, whereas it is possible to observe a reduction of the inclination of the pad described by angle β .

Considering the PEEK coating whose simulation results are reported in Table 15.8, the increase of the coating thickness has negligible influence on the pressure distribution and other quantities.

The PEEK has a Young's modulus of 3.6   GPa, lower than that of babbit metal, while the yield point is greater and equal to 90–100   MPa (variable according to the "enrichment"). The comparisons of the simulated TEHD models, for the same thickness of coating, showed that the PEEK has less crowning effect than PTFE (having much greater Young's modulus), but it still allows an increase of the specific load compared with the babbitt metal. It also has a lower power loss. The PEEK is already used commercially for coating of bearings, but also has some drawbacks, mainly the adhesion to the hard steel support, so that early applications envisaged a layer of intermediate bonding made of copper or brass. Moreover, it has a low thermal conductivity, which limits the conduction of heat generated in the oil-film towards the bearing and increases the phenomenon of the hot oil carry over [4,48].

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