The present invention relates to a turbo-machine provided with an automatic thrust force balancing system.

The high axial forces generated by high-speed turbo-machines make the thrust bearings thereof very critical components in order to ensure safe and reliable service.

Conventional turbo-machines achieve a balance of internally generated, fluid-dynamic thrust forces, only for a single process duty for a new machine. Variation in process

conditions and deterioration in the machine's condition (wear, fouling, etc.) generate increasing thrust forces, which can lead to bearing-overload and -failure, often with significant consequential damage.

Examples of prior art turbo-machines are described in the following documents:

US 4.884.942;

US 4.385.768;

US 3.895.689;

EP 0.550.801;

WO 95/35447; and

WO 01/16466.

The present turbo-machine resolves the problem of bearing- overload and -failure by being provided with an automatic thrust force balancing system for the internally generated thrust forces in the turbo-machine . Therefore, the thrust bearing only needs to handle a negligible residual thrust force, which remains constant, regardless of variations in process or machine condition.

The range of turbo-machines for which the automatic thrust force balancing system may be incorporated, comprises multistage (i.e. three or more stages) compressors and pumps, but also large, high-power one-stage or two- stage compressors and pumps, which are typically used in fluid pipeline systems. The turbo-machines may be driven by any suitable gas medium, such as gaseous hydrocarbons and steam, for example steam for a steam turbine .

Turbo-machine safety, reliability and performance will be improved significantly by incorporating the automatic thrust force balancing system into the turbo-machine .

According to the present invention, a turbo-machine

structured for automatic balancing of fluid-dynamic thrust forces generated within the turbo-machine during operation thereof is hereby provided. The turbo-machine is provided with an automatic thrust force balancing system comprising:

- a balance piston attached around a discharge end of a rotating shaft of the turbo-machine for providing a thrust force capable of counter balancing thrust forces generated by impellers in the turbo-machine during operation thereof;

- a balance chamber structured for receiving a leakage flow emanating from the turbo-machine during operation thereof, said balance chamber communicating with a downstream side of the balance piston;

- a balance flow return line connecting the balance chamber with a suction end of the turbo-machine for conveying said leakage flow; and

- a control valve connected to the balance flow return line for regulating the flow rate of the leakage flow. The

distinguishing characteristic of the turbo-machine is that the control valve is connected to a temperature controller for regulating the degree of opening of the control valve;

- wherein the temperature controller is connected to, and receives control signals from, at least one temperature sensor structured to sense and forward the operating

temperature of a thrust bearing supporting the rotating shaft in the turbo-machine .

The automatic thrust force balancing system allows the control valve and the leakage flow rate to be regulated based on the bearing temperature so as to regulate a back pressure on the balance piston and hence regulate the counter

balancing thrust force exerted by the balance piston on the rotating shaft and the thrust bearing.

In one embodiment, said balance piston may be cylindrical and disposed within a cylindrical bush so as to provide an annular clearance between the balance piston and the bush, the clearance of which is capable of passing the leakage flow through the annular clearance and into the balance chamber. This annular clearance is sufficiently small to allow the leakage flow to experience a pressure drop when passing through the clearance, the leakage flow having a high

pressure at the upstream side of the balance piston and a reduced pressure at the downstream side thereof, which communicates with said balance chamber. This pressure drop is adjusted by regulating the back pressure on the balance piston, hence regulating said counter balancing thrust force exerted by the balance piston. Moreover, said temperature sensor may be embedded in the thrust bearing. As such, said temperature sensor may be embedded in a set of bearing pads of the thrust bearing, for example in a so-called active set of bearing pads of the thrust bearing.

Furthermore, the present turbo-machine may be comprised of a centrifugal compressor or a centrifugal pump.

As such, the turbo-machine may be comprised of a one-stage or two-stage turbo-machine, for example a one-stage or two-stage centrifugal compressor adapted for use on a pipeline system.

Alternatively, the turbo-machine may be comprised of a multistage turbo-machine, multi-stage being defined as the three or more stages provided within the turbo-machine .

Yet further, the multi-stage turbo-machine may be provided with stages arranged in a so-called "in-line" configuration or, alternatively, arranged in a so-called "back-to-back" configuration.

Embodiments of the present turbo-machine will now be

described in context of a four-stage centrifugal compressor, as an example, and with the aid of the following drawings, where :

Figure 1 shows a conventional four-stage compressor having its stages arranged in an "in-line" configuration, the figure also showing details of the compressor in an enlarged

circular section thereof;

Figure 2 shows different pressure profiles and pressure- affected areas over the internal components of the

compressor; Figure 3 shows the manner in which the combination of

unbalanced areas/pressures on impellers in the compressor produce significant thrust forces, which are mostly balanced by a balance piston, but which leaves a residual thrust force that has to be counteracted by a thrust bearing in the compressor;

Figure 4 shows a diagram of the loading on the thrust bearing due to variation of internally generated thrust forces according to process and machine conditions;

Figure 5 shows various views and sections of a typical hydrodynamic thrust bearing used for this class of turbo- machines ;

Figure 6 shows various views and sections of four-stage compressor structured in accordance with the present

invention, the compressor of which automatically balances the internally generated thrust forces regardless of process and machine conditions;

Figure 7 shows a diagram of the correlation between thrust bearing load and pad temperature; and

Figure 8 illustrates an alternative multi-stage compressor structured in accordance with the present invention and having its stages arranged in a "back-to-back" configuration.

A centrifugal compressor with a conventional thrust balancing system is shown in figure 1. Gas at pressure Ps enters the compressor at a suction flange 1, after which the gas is compressed by four impellers 2 which are driven by a rotating shaft 3 within the compressor. The impellers 2 increase the gas pressure from Ps at entry to Pd at a discharge flange 4. Each impeller 2 discharges its gas flow into a stationary- component called a diaphragm 28. The purpose of each

diaphragm 28 is to convert the kinetic energy of the received gas flow into pressure recovery before returning the gas flow to the next impeller 2. Each impeller/diaphragm combination is called a stage 29, and the respective stages 29 are stacked together to form a complete inner assembly 16, which is commonly called a bundle. The inner assembly 16 is

inserted in a casing 11 and is closed and statically sealed by an end cover 13. The dynamic annular flow paths between the rotating shaft 3 and the casing 11 are sealed using dry gas seals 14.

Moreover, a balance piston 5 is attached at a discharge end of the rotating shaft 3. One important function of this balance piston 5 is to control a gas leakage flow from a suction end to the discharge end of the compressor. The balance piston 5 is structured so as to provide a close- fitting annular clearance between the balance piston 5 and a bush 6 in the casing 11. The gas leakage flow, which

discharges from this annular clearance, then enters a balance chamber 7 provided at the downstream side of the annular clearance, after which the gas leakage flow is returned to the suction end of the compressor via a balance flow return line 8. Another important function of the balance piston 5 is to provide a thrust force which counter balances the thrust forces generated by the impellers 2. The perfect thrust balance, however, is only achieved for one operating _ condition. Variations in process conditions, and

deterioration in the machine's condition, produce imbalances in the internally generated thrust forces. These imbalances produce residual thrust forces that must be counteracted by a thrust bearing 9 located at the suction end of the compressor and supporting the rotating shaft 3. This thrust bearing 9 comprises two sets of bearing pads 15 and 17. The residual thrust force is transferred from the rotating shaft 3 to the thrust bearing 9 by means of a thrust collar 10 which is disposed between the two sets of bearing pads 15, 17, and which is shrunk onto the rotating shaft 3 at the suction end of the compressor.

Figure 2 shows, on one side, a first pressure profile

(pressure difference) acting across the pressure-affected areas of the impellers 2 and, on the other side, a second pressure profile (pressure difference) acting across the pressure-affected areas of the balance piston 5. These pressure profiles and corresponding areas differ for the impellers 2 and the balance piston 5, respectively.

Multiplying these unbalanced areas by the corresponding pressure differences will yield the thrust forces shown in figure 3, which are indicated by way of example.

Furthermore, figure 3 shows that the net (or residual) thrust force acting on the impellers 2 is oppositely directed to the net thrust force acting on the balance piston 5. In this particular embodiment, it can be seen that the total thrust force of 50 Tons on the impellers 2 is not quite balanced by the counter thrust force of 45 Tons on the balance piston 5, leaving a residual thrust force Ttb of 5 Tons acting on the thrust bearing 9. The reason for having such a residual thrust force Ttb acting on the thrust bearing 9 instead of having perfectly balanced internal thrust forces acting thereon, is to ensure that a residual thrust force is always acting on the stationary, active set of bearing pads 17 located on the active side of the thrust bearing 9 (see figure 5) . This ensures that the reactive thrust forces from the stationary, active set of bearing pads 17 is always pushing the rotating thrust collar 10 onto a shoulder 12 of the rotating shaft 3. This is very sound practice, particularly for compressors, since the inactive set of bearing pads 15, which is located on the inactive side of the thrust bearing 9 (see figure 5) , should be exposed to minimum loads. This is also sound practice for correct location of the rotating shaft 3 during transients (i.e.

stopping/starting of the compressor). When high loads have been placed on the inactive set of bearing pads 15, there have been serious accidents in the past whereby the thrust collar 10 has dislocated from the rotating shaft 3 so as to cause extensive damage.

Figure 4 shows a diagram of the manner in which loads on the thrust bearing 9 change in response to variations in the internally generated thrust forces, which result from changes in process and machine condition (wear and fouling) .

More particularly, curve A-A shows the specific bearing load plotted against gas flow for a new compressor at rated speed and suction and discharge pressures. Moreover, curve B-B shows the specific bearing load plotted against flow for a new compressor, and at rated speed and discharge pressure, but with a 10 % increase in suction pressure. Furthermore, curve C-C shows the specific bearing load plotted against flow for a worn compressor at rated process conditions.

Lastly, curve D-D shows the performance characteristics of a compressor structured in accordance with the present

invention. This compressor incorporates an automatic thrust force balancing system for allowing the compressor to

maintain its specific bearing load on a low and constant level irrespective of the gas flow rate through the

compressor .

Figure 4 also shows a thrust bearing failure zone defined between specific bearing load failure thresholds (Fw-Fw) , which represents a worn bearing, and (Fn-Fn) , which

represents a new bearing. From this figure, it is clearly apparent that combinations of variations in process and machine conditions eventually result in specific bearing loads exceeding the failure threshold. For a new, steel- backed centre pivot bearing, the failure threshold (Fn-Fn) is at approximately 500 psi (3.45 N/mm ² ) . When the bearing itself degrades on the sliding surface and the

pivots/levelling plates, the failure threshold (Fw-Fw) can fall to as low as 375 psi (2.6 N/mm ² )

The above analysis depicted in figure 4 clearly illustrates the problems with conventionally thrust-balanced turbo- machines. Several major failures have been documented in technical papers from 1970 to the present time.

The very essence of the present turbo-machine is to provide a technical solution allowing the thrust bearing of the turbo- machine to better withstand these excursions in thrust forces at high speeds .

Figure 5 details the typical hydrodynamic thrust bearing 9 that is used for large, high-speed turbo-machines. This hydrodynamic thrust bearing 9 is provided with tilting sets of bearing pads 17 and 15, each of which is provided with a pivot 27 for allowing some tilting of the set of bearing pads. The pivots 27 are made of hardened steel and are free to pivot on a bearing carrier 20. Similar to the thrust bearing shown in figure 1, the set of bearing pads 17 is active and the set of bearing pads 15 is inactive. The base material for the tilting sets of bearing pads 17 and 15 is usually steel with a thin layer of white metal 18 bonded to the steel so as to form a bearing surface in contact with the rotating thrust collar 10. Due to its high criticality, the thrust bearing 9 is commonly equipped with an on-line condition monitoring system which provides monitoring, alarm and automatic machine shutdown on high values of the following critical parameters:

• High bearing temperature monitored by embedded

temperature sensors 21 and 22, which are located at the highest temperature part of the respective set of bearing pads 15, 17;

• High axial displacement of the rotating shaft 3

monitored by a first proximeter probe 23; and

• High shaft vibrations monitored by a second proximeter probe 24.

A compressor (turbo-machine) structured in accordance with the present invention and incorporating an automatic thrust force balancing system is illustrated in figure 6. The balancing system is designed in a manner allowing the counter balance thrust force from the balance piston to be

automatically regulated in order to balance, for all process and machine conditions, the thrust forces from the impellers 2. Referring again to figure 4 and curve D-D thereof, this balancing system provides the advantageous result of

providing a minimal and constant specific bearing load on the thrust bearing 9, as compared to the conventional thrust balanced compressor, which can overload its thrust bearing as process and machine conditions change (see curves A-A, B-B and C-C in figure 4) .

Returning to figure 6, the thrust force of the balance piston 5 on the rotating shaft 3 is regulated by a control valve 25, which is installed in the balance return line 8, and which varies a back pressure Pbc (see figs. 1 and 2) on the balance piston 5. The pressure drop over the balance piston 5 can thereby be changed, which in turn varies the thrust force (thrust force = pressure drop multiplied by pressure-affected area) exerted on the rotating shaft 3 by the balance piston 5. The selected control parameter is the temperature of the active set of bearing pads 17, the temperature of which is monitored by the temperature sensor 22. In this embodiment, the sensor 22 relays an output signal to a temperature controller 26, which in turn outputs a 4-20 mA (milliampere) signal to the control valve 25 so as to regulate the degree of opening of the control valve 25.

Figure 7 shows a temperature-load curve (Tb-Tb) of the correlation between thrust bearing load and bearing pad temperature . Either of these two parameters could have been used as the selected control parameter for regulating the control valve 25. In this embodiment, however, the embedded temperature sensor 22 is selected due to its excellent reliability and minimum requirement for re-calibration, as compared to measuring the thrust bearing load by means of load cells.

The present compressor (turbo-machine) will now be

illustrated by way of a practical example.

Referring again to figure 7, assuming that the desired bearing load is 20 % of the rating for the thrust bearing 9. According to curve (Tb-Tb) , the desired load indicates a bearing pad temperature of 70 degrees Celsius. A 70 degrees Celsius temperature setting is then set at the temperature controller 26, after which the control valve 25 regulates accordingly until the bearing pad temperature stabilizes at 70 degrees Celsius. The temperature-load curve (Tb-Tb) is obtained from the bearing manufacturer, and the temperature at zero load and rated speed can be confirmed during commissioning by- searching for the lowest temperature set point that moves the rotor of the turbo-machine from the active set of bearing pads 17 to the inactive set of bearing pads 15, as indicated by the first, axial proximeter probe 23.

As mentioned earlier, it is safe design practice to have the impeller thrust forces exceed the counter balance thrust force from the balance piston 5. This under-compensation ensures that the active set of bearing pads 17 is always pushing the thrust collar 10 against the shoulder 12 of the rotating shaft 3. The thrust collar 10 is thereby prevented from loosening from the rotating shaft 3, which has been the cause of previous major failures.

The diameter of the balance piston 5 is designed to balance the highest calculated thrust load from the impellers 2 when the control valve 25 is in its fully open position. This maximum thrust load is based on the process and machine conditions yielding the highest impeller thrust force. For all other operating parameters, such as process flow rates, pressures, gas densities and machine conditions, the control valve 25 will modulate the back pressure Pbc on the balance chamber 7, and also the counter thrust force on the balance piston 5, so as to maintain a minimum and constant load on the active set of thrust bearing pads 17.

As compared to the performance characteristics of

conventional thrust balanced compressors, the benefits of this novel compressor is shown clearly in figure 4. According to curve D-D, the automatic thrust force balancing system of the present compressor ensures that the thrust bearing 9 is always automatically maintained on a low and constant loading regardless of changes in process and machine conditions. An additional benefit of using the present compressor with such a thrust force balancing system is a reduction in the bearing and balance flow power losses, which increases the turbo- machine's efficiency in the order of 1-3 %, depending on the duty of the compressor (turbo-machine) .

Figure 8 shows an alternate design of a multi-stage

compressor structured in accordance with the present

invention and having its stages arranged in a "back-to-back" configuration. This compressor is effectively split into two sections, including a low pressure section 31 and a high pressure section 32. The two sections 31, 32 are statically connected via a crossover line 33. The two sections 31, 32 are also dynamically separated by means of a close-fitted centre bush 34 and a centre sleeve 35. The centre sleeve 35 also acts as a balance piston with a high pressure discharge pressure on one side, and a low pressure discharge pressure on the other side. To balance the pressures at each end of the compressor, the compressor is provided with an additional combination of a centre bush 36 and a centre sleeve 37. This arrangement reduces the pressure at the discharge end 38 of the compressor from low pressure discharge pressure to low pressure suction pressure at the suction end 39 of the compressor. This centre sleeve 37 acts as a separate balance piston. The back pressure Pbc on the centre sleeve 37 can be regulated by means of the control valve 25 so as to

automatically balance and maintain a minimum thrust force on the thrust bearing 9. This is carried out in exactly the same way as for the multi-stage "in-line" compressor shown in figure 6.