GB2462635A

GB2462635A - Turbo-machine axial thrust balancing

Info

Publication number: GB2462635A
Application number: GB0814884A
Authority: GB
Inventors: William Paul Hancock
Original assignee: Individual
Current assignee: Individual
Priority date: 2008-08-14
Filing date: 2008-08-14
Publication date: 2010-02-17
Anticipated expiration: 2028-08-14
Also published as: GB2462635B; GB0814884D0

Abstract

A turbo-machine which can automatically balance its axial thrust. Counter thrust produced by a balance piston 5 may be regulated by back pressure in a balance chamber 7 adjacent the balance piston 5. A balance flow control line 8 may extend between the balance chamber 7 and an inlet 1 of the turbo-machine, a control valve 25 being used to control flow through the line 8 and as such the back pressure in the balance chamber 7. The control valve 25 may be regulated by a temperature controller 26 which receives signals from temperature sensors 21, 22 located in thrust bearing pads. The turbo-machine may be a single or two stage centrifugal pump or compressor, or may be a multi-stage centrifugal pump or compressor with in-line or back-to-back stages.

Description

Turbomachinery thrust balancer.

This invention relates to an automatic thrust force balancing system for turbomachinery.

The high axial thrust forces generated by high speed turbomachinery, make the thrust bearing, a very critical component, to ensure safe and reliable service. Conventional turbomachines, achieve a balance of internally generated thrust forces, only for a single process condition, and when the machine is new. Variation in process conditions, and deterioration of the machine condition, generate increasing thrusts which can lead to bearing overloading and failure, often with significant consequential damage.

The invention resolves the above problem, by providing an automatic balancing system for the internally generated thrusts. The thrust bearing therefore, only needs to handle a negligible residual thrust, which remains constant, regardless of variations in process or machine condition.

The ranges of turbomachinery covered by the invention are multi-.stage(3 or more stages) compressors and pumps, also large single/two stage compressors I pumps.

Turbomachinery safety, reliability and performance will be improved significantly by the invention.

The invention will now be described using a four stage centrifugal compressor as an example, and with the aid of the following drawings: Figure 1 shows a conventional four stage compressor where the stages are arranged in an in-line" configuration..

Figure 2 shows the different pressure profiles and areas over the internal components.

Figure 3 shows how the combination of unbalanced areas/pressures on the impellers, produce significant thrust, which is mostly balanced by a balance piston, but which leaves a residual thrust, that has to be handled by the thrust bearing Figure 4 shows the loadings on the thrust bearing due to variation of internally generated thrusts, according to process and machine conditions.

Figure 5 shows the typical hydrodynamic tilting pad thrust bearing used for this class of turbomachinery.

Figure 6 shows the invention, which automatically balances the internally generated thrust forces, regardless of process and machine conditions.

Figure 7.shows the correlation between thrust bearing load and pad temperature.

Figure 8 illustrates an alternative multi-stage compressor, where the stages are arranged in a "back to back" configuration.

A centrifugal compressor with a conventional thrust balancing system, and with the stages arranged "in-line" is illustrated on figure 1. Gas enters at suction flange 1 at pressure Ps, and is compressed by four impellers 2, which are driven at high speed by shaft 3. The impellers increase the pressure from Ps to Pd at discharge flange 4.

Each impeller 2, discharges the flow into a static component called a diaphragm 28,the purpose of which, is to convert the kinetic energy of the flow leaving the impeller 2, into pressure recovery, before returning the flow to the next impeller. Each impeller! diaphragm combination is called a stage 29, and the stages are stacked together to form a complete inner assembly 16, which is commonly called a bundle.

To complete the compressor unit, the inner assembly 16 is inserted in the casing 11 and is closed and statically sealed, by a bolted on end cover 13. The dynamic seals between the rotating shaft 3 and the casing 11, are sealed using tandem dry gas seals 14..

The balance piston 5. is attached at the discharge end of the shaft and fulfils the following important functions: * Controls the leakage flow from discharge to suction, by a close fitting annular clearance between balance piston 5 and a bush 6 in the casing 11. The leakage flow enters the balance chamber 7, whence it is returned to suction by balance flow return line 8.

* Provides a thrust force which counter balances the thrust forces generated by the impellers. The perfect thrust balance however, is only achieved for one operating condition. Variations in process conditions and deterioration in machine condition, produce imbalances in internally generated thrusts, which produce residual thrust forces, which have to be handled by the thrust bearing 9.

The residual thrust is transferred from the shaft 3 to the thrust bearing 9 by means of a thrust collar 10, which is shrunk onto the shaft at the suction end of the compressor.

Figure 2 indicates the differences in areas and pressure profiles over the impellers and balance piston. Multiplying the unbalanced areas by the pressure differences, gives the thrust forces shown in Figure 3.

Using figure 3, it can be seen that the total impeller thrust of 50 Tons, is not quite balanced by the counter thrust of 45Tons. The reason for this under-compensation, is to ensure that the residual load on the thrust bearing is always acting on the active side 3.

bearing pads 17. This means that the reactive thrust forces from the stationary active bearing pads are always pushing the rotating thrust collar 10 onto the shoulder 12 of the shaft.. This is very sound practice, particularly for high speed compressors, since the in-active bearing 15, should be exposed to minimum load, and is used only for locating the rotor on starting and stopping. There have been serious failures, when high loads have been placed on the in-active bearing 15, which have dislocated the thrust disc 10 from the shaft 3, causing extensive consequential damage to the compressor.

Figure 4 shows how the loads on the thrust bearing change according to variations in the internally generated thrust forces, as process conditions (Row/pressure) and machine conditions( wear/fouling etc) change: Curve AA is the specific bearing load for a new compressor, at rated speed, suction and discharge pressures and variation in flow..

Curve BB is the specific bearing load for a new compressor, at rated speed, discharge pressure but with a 10% increase in suction pressure and variation in flow.

Curve CC is the specific bearing load for a worn compressor, at rated speed, suction and discharge pressures and variation in flow.

It can be seen that thrust bearing easily becomes loaded above the specific bearing pressure thresholds for failure, For a new steel backed, centre pivoted bearing, at 120 metres/sec. collar speed, the pressure threshold for failure (Fn-Fn) is 500psi.

When the bearing degrades on the sliding surface, and the pivots /leveling plates wear, the failure threshold (Fw-Fw) can fall to as low as 375psi.(2.6 Newtons/square mm.).

The above analysis clearly indicates the potential problems with conventionally thrust balanced turbomachinery, and several major failures have been documented in technical papers from 1970 to 2008.

The thrust bearing, and its ability to handle these excursions in thrust forces at high speeds, is the very essence for this invention.

Figure5,. shows the typical hydrodyna.mic tilting pad thrust bearing 9 that is used for this demanding service The base material for the tilting thrust pads 17 and 15 (active and in-active)is usually steel, with a thin layer of white metal 18 bonded to the bearing surface, and which contacts the rotating thrust collar 10. The reason for use of the soft white metal, is that at low bearing speeds (less than 10 metres/sec) on starting/stopping, the hydrodynamic oil wedge to the pads 17 is not established. The white metal is an excellent dry running material and prevents bearing failure during transient low speed conditions.

The pivots 27 are made of hardened steel, and are free to pivot on the bearing carrier 20 4.

Due to the high criticality, the bearing has an on-line condition monitoring system which provides monitoring, alarms and automatic turbomachine shutdown, on high values of the following critical parameters: * High bearing pad temperature by embedded temperature sensors, 21/22 which are located at the highest temperature part of the pad.

* High axial displacement of the shaft 3 by proximeter probe 23 * High shaft vibration by proximeter probe.24 The invention is illustrated in figure 6. The system is designed, so that the counter balance thrust force from the balance piston, is automatically regulated to balance the thrusts from the impellers, for all process and machine conditions. Referring again to figure 4, the invention provides a minimal and constant loading line D-D on the thrust bearing, compared to conventional thrust balanced compressors, which can quickly overload the thrust bearing, as process and machine conditions change.( Curves, AA BB,CC) Returning to figure 6, the balance piston thrust force, is regulated by control valve 25, installed in the balance return line 8 and which varies the back pressure on the balance piston 5. The pressure drop over the balance piston can thereby be changed, which in turn, varies the balance piston thrust.(thrnst pressure drop multiplied by area). The selected control parameter, is active bearing pad temperature 22, which relays an output to temperature controller 26,which sends the 4-20 milli-Amp signal to the control valve Figure 7 shows the correlation (Th-Th) between thrust bearing load and pad temperatures. Either of these parameters, could be used as the control parameter, but the embedded temperature sensor is selected, due to excellent reliability and minimum requirement for re-calibration, compared to measuring the bearing thrust load directly using load cells.

The invention will now be illustrated by use of a practical example: If the desired bearing load is 20% of rating, figure 7, shows that a pad temperature set point of 70 degrees Celsius, should made on the temperature controller 26. The control valve 25 then reacts accordingly, until the pad temperature of 70 degrees C is obtained.

The temperature load curve is obtained from the bearing manufacturer, and the temperature at zero load and rated speed, can be determined during commissioning. This is achieved by progressively reducing the temperature set point on temperature controller 26, until the shaft thrust collar 10 moves from the active bearing 17 to the in-active bearing 15, as noted by the axial proximeter probe 23..

Iii accordance with safe compressor practice, it is important that the impeller forces always exceed the counter thrust force from the balance piston. The impeller thrust forces are under-compensated to ensure that the active bearing 17 is always pushing the thrust collarlO against the shoulder 12 of the shaft 3. The collar is prevented thereby in loosing from the shaft which has been the cause of previous major failures.

The diameter of the balance piston is designed, to balance the highest calculated thrust load from the impellers with control valve 25 in the fully open position. For all other process flows, pressures, gas densities and machine conditions, the control valve 25 will modulate the back pressure and balance piston counter thrust, to maintain a minimum load on the active thrust bearing.

The benefits of this invention, compared to conventionally thrust balanced compressors, are indicated in figure 4. The bearing is always automatically maintained on a low and constant loading DD,regardless of process, or machine conditions, Turbomachinery safety, reliability and operating envelope, are enhanced significantly.

An additional benefit, is that reduced bearing power loss and optimization of balance flow, will provide an efficiency gain of 1-3 %, depending on the turbomachine duty.

Figure 8, shows an alternate multi-stage compressor design, with the stages arranged in the "back to back" configuration. The compressor is effectively split into two sections, a low pressure section 31 and a high pressure section 32, The two sections being statically connected by a crossover 33 and dynamically separated, by a close fitting centre bush 34 and sleeve 35, The centre sleeve also acts as a balance piston, since on one side is 1-IP discharge pressure and LP discharge on the other.

To balance the pressures at each end of the compressor, there is an additional bush 36/ sleeve 37 combination, and a return line 8 to suction. This facility reduces the pressure at the inboard end 38 of the compressor from LP discharge pressure (LPd) to approximately the same pressure at the outboard end 39.

The sleeve 37, also acts as a separate balance piston, and the back pressure can be modulated by control valve 25 to automatically balance and maintain minimum thrust on thrust bearing 9, in exactly the same way as the multi stage "in-line "compressor shown in figure 6.

Claims

Claims.1. A turbomachinery thrust balancer which automatically balances the turbomachine' s internally generated fluid dynamic thrust forces.
2. A turbomachinery thrust balancer according to claim 1, where the counter thrust from the balance piston is regulated by the back pressure in the balance chamber.
3. A turbomachinery thrust balancer according to claims 1 and 2, where the balance chamber pressure is regulated by a control valve in the balance flow return line.
4. A turbomachinery thrust balancer according to claims 1,2 and 3, where the control valve is regulated by a temperature controller which receives signals from temperature sensors embedded in the turbomachine's thrust hearing pads.
5. A turbomachinery thrust balancer according to claims 1, 2,3 and 4, in which the turbomachine is a multi-stage centrifugal compressor with" in-line" stages.
6. A turbomachinery thrust balancer according to claims 1,2, 3 and 4, in which the turbomachine is a multi-stage centrifugal compressor with "back to back "stages..
7. A turbomachinery thrust balancer according to claims 1, 2,3 and 4, in which the turbomachine is a single or two stage centrifugal compressor.
8. A turbomachinery thrust balancer according to claims 1, 2,3 and 4, in which the turbomachine is a multi-stage centrifugal pump with "in-line" stages.
9. A turbomachinery thrust balancer according to claims 1,2,3 and 4, in which the turbomachine is a multi-stage centrifugal pump with "back to back" stages
10. A turbomachinery thrust balancer according to claims 1,2,3 and 4, in which the turbomachine is a single or two stage centrifugal pump.
11. A turbomachinery thrust balancer according to any of the above claims, in which thrust force on the machine's thrust bearing can be controlled at the optimum value consistent with high reliability and long equipment lifetime.
12. A turbomachinery thrust balancer according to any of the above claims, in which the residual fluid dynamic thrust is handled only by the active pads of the thrust bearing.
13. A turbomachinery thrust balancer according to any of the above claims, in which the efficiency of the machine is increased by reductions in thrust bearing and balance flow power losses.AMENDMENT STO THE CLAIMS HAVE BEEN FILED AS FOLLOWSClaims.1. A turbo-machine which automatically balances its internally generated fluid dynamic thrust forces, the turbo-machine comprising a balance piston located at one end of a rotor, back pressure in a balance chamber located adjacent to the balance piston being used to regulate counter thrust of the balance piston, the balance chamber pressure being regulated by a control valve in a balance flow return line, the valve being regulated by a temperature controller which receives signals from temperature sensors embedded in thrust bearing pads of the turbo machine. * . * ** * * S *S.. S. * . * *e. S...S *S.. S. * S S* * S.