Abstract:
The present invention provides a method and apparatus of inhibiting a thrust bearing capacity of a compression system from being exceeded during a surge event in which a thrust bearing is biased with a biasing force to increase the thrust bearing overload margin between the capacity of thrust bearing to absorb axial forces and the greatest force produced during the surge event. This biasing force can be produced by appropriately sizing the high pressure seal on the side of the impeller opposite to the inlet of a compressor of the compression system so that the back disk force produced in the high pressure region of the high pressure seal and the low pressure region located inwardly of the high pressure region creates the desired bias force value and direction.

Description:
FIELD OF THE INVENTION 
       [0001]    The present invention relates to a surge protection method and apparatus in which the capacity of a thrust bearing of a compressor is prevented from being exceeded during a surge event. More particularly, the present invention relates to such a method and apparatus in which a biasing force is exerted on an impeller of a compressor and through the compressor shaft to the thrust bearing, during normal operation of the compressor, by producing a force difference between an impeller eye side force and a back disk force exerted on the impeller through the back disk seal that increases the overload margin between the thrust bearing capacity and a surge force occurring during the surge event. 
       BACKGROUND OF THE INVENTION 
       [0002]    Centrifugal compressors are well known in the art and are used in many applications to compress a gas from a lower pressure to a higher pressure. The gas at the lower pressure enters an inlet of the centrifugal compressor and is compressed to a higher pressure by being accelerated by a rotating impeller and then sent into a diffuser surrounding the impeller, in which additional pressure is recovered by decelerating the gas. The gas is discharged from the diffuser to a volute and from the volute to an outlet thereof at the higher pressure. 
         [0003]    In a centrifugal compressor, a high pressure seal is provided on the back of the impeller or in other words on the side of the impeller opposite to the inlet. This seal is typically a labyrinth type of seal that has a smaller diameter than the outer diameter of the impeller as defined by the outer rim thereof. High pressure gas seeps through the outer rim during operation of the compressor and such seal prevents this high pressure gas from pressurizing the entire back of the impeller. A lower pressure gas from a location such as the compressor inlet or an even lower pressure source if available, is introduced into a cavity formed at the back of the compressor, between the motor shaft driving the impeller and the high pressure seal. Leakage from the cavity is prevented by a shaft seal on the motor shaft driving the impeller. The pressure within this cavity acts on the resulting inner annular area behind the impeller along with the high pressure acting in the outer annular area between the outer rim of the impeller and the high pressure seal. The sum of these two forces produce a force that is known as the back disk force on the impeller. Acting inwardly of the impeller, is an eye side force produced by the onrush of gas entering the impeller and acting on the surface of the impeller facing to the inlet. In the prior art, the high pressure seal is sized so that during design conditions the eye side force is in balance with the back disk force. 
         [0004]    Compressors are designed to operate within an operating envelope that can be plotted in what is referred to as a compressor map of pressure ratio between outlet pressure and inlet pressure versus flow rate through the compressor. On such a plot, a peak or best efficiency operating line is plotted in which for a given flow rate and pressure ratio, the energy consumption of the compressor is at a minimum. The plot also has another dimension of specific speeds which cross the peak efficiency operating line. If the pressure ratio falls within the compressor for a given speed, a point is ultimately reached when the compressor goes into what is referred to as surge. A surge event is therefore, produced by flow rate through the compressor falling below a minimum flow required at a given speed of the impeller of the compressor that is necessary to maintain stable operation. From a viewpoint taken from the impeller, a surge event has two phases. In a first phase there is a sudden loss of discharge pressure within the impeller flow passages arising from aerodynamic instability in the impeller. This results in a decrease in the eye side pressure. However, the back disk force, remains high due to fluidic system inertial effects. This results in a large axial force on the impeller driving the impeller towards the inlet. The second phase of the surge event is produced by the high pressure back disk pressure bleeding down to lower eye side pressure, resulting in the eye side force overcoming the weakened back disk force. This produces a large axial force driving the impeller away from the inlet. The frequency at which these forces are developed is quite high resulting in destruction of the compressor starting with an overload of the thrust bearing. 
         [0005]    The forces developed during a surge event can be computed or measured and the thrust bearings in the compressor can be designed with an overload margin to absorb these forces and prevent damage to the compressor. However, there has been the increasing use of what are referred to as oil-free bearings in centrifugal compressors. These oil free bearings are either electromagnetic bearings or aerodynamic bearings. In electromagnetic bearings, the motor shaft is suspended in both radial and axial directions by means of a magnetic force. However, such oil-free bearings when applied to the thrust bearing are less able to adsorb the large axial forces that occur during a surge event than conventional bearings that are lubricated with oil. Thus, the overload margins that can be provided by oil-free bearings are less than those that can be provided by conventional bearings. As such, during a surge event, the thrust bearing will be unable to absorb the forces and back-up-, bearings or bushings will be contacted to adsorb the force. The problem with this is that back-up bearings or bushings can only be subjected to a predefined small number of usages during these surge events. After this predetermined number, the compressor will have to be taken off-line for maintenance to replace the back-up bearings or bushings. 
         [0006]    The problems outlined above with respect to oil-free bearings have therefore, limited their use in large compressor applications such as air separation plants in which such bearings would otherwise have the advantage of not contaminating the feed to the plant with oil, having much lower frictional losses, and will require less maintenance than conventional oil lubricated bearings. In an application such as an air separation plant, if a compressor is taken off-line, the plant will not function. Unplanned outages in any large installation will in any case result in financial hardship to the plant operator. 
         [0007]    As will be discussed, unlike the prior art, the present invention provides a method and apparatus in which the impeller is preloaded with a biasing force that is larger than that required to simply balance impeller eye side and back disk forces during normal operation of the compressor. This preload force is used to increase an overload margin of the bearings that could be exceeded during a surge event and thus, is particularly applicable to a centrifugal compressor using oil-free bearings. 
       SUMMARY OF THE INVENTION 
       [0008]    The present invention provides a method of inhibiting a thrust bearing capacity of a compression system from being exceeded during a surge event. In accordance with such method, the thrust bearing is biased with a biasing force acting in one of two opposite axial directions of a compressor shaft. In such a compressor, the compressor shaft is connected to the thrust bearing and to an impeller of at least one compressor that is in turn provided with a back disk seal. The compression shaft is driven by a driver to drive the impeller through rotation of the compressor shaft and the compression system has at least one stage of compression provided by the at least one compressor. The thrust bearing has a first thrust bearing capacity and a second trust bearing capacity to absorb axial loads acting upon the thrust bearing in the two opposite axial directions of the compressor shaft, respectively. Such thrust bearing is configured such that a first overload margin exists between the first thrust bearing capacity and a first surge force and a second overload margin exists between the second thrust bearing capacity and a second surge force. The first surge force and the second surge force act in the two opposite axial directions and are created during the surge event. The biasing force is exerted on the impeller and through the compressor shaft to the thrust bearing, during normal operation of the compression system, by producing a force difference between an impeller eye side force and a back disk force exerted on the impeller opposite to the eye side force. The back disk force is exerted by a high pressure outer annular region and a low pressure inner annular region of the impeller that is separated by the back disk seal. The biasing force increases one of the first overload margin and the second overload margin such that the first overload margin and the second overload margin are no less than about twenty-five percent of the first through bearing capacity and the second thrust bearing capacity, respectively. 
         [0009]    Although the increase of one of the overload margins will be at the expense of the other of the overload margins of the thrust bearing, the largest surge force can thereby be compensated for to increase the ability of the thrust bearing to absorb surge forces. In this regard, at a minimum, the overload margins twenty-five percent of the thrust bearing capacities. As can be appreciated, even larger margins are desirable if possible, for example, fifty percent. Although the present invention is not limited to the use of oil-free bearings as the invention would in fact be advantageous even in the case of the use of oil free bearings, it has particular application to magnetic and aerodynamic bearings that have limited overload margins. This increase in the ability of the thrust bearing to adsorb such surge forces allows the use of such oil-free bearings in large-scale compressor applications and in applications where continued operation of the compressor is particularly critical. 
         [0010]    The driver can be an electric motor and the compressor shaft can be a motor shaft protruding from the electric motor. In such case, the thrust bearing is part of the electric motor. Further, the thrust bearing can be a magnetic bearing. 
         [0011]    The at least one compressor can comprise a first compressor and a second compressor connected at opposite ends of the motor shaft and in flow communication with each other such that least two successive compression stages are provided in the compression system. In such case, the force difference is produced by a difference between the eye side force and the back disk force of at least one of the first compressor and the second compressor. The at least one compressor can be an upstream compression stage of a plurality of compression stages of the compression system. In such case, the low pressure region is in flow communication with an outlet or an inlet of one other compressor of an upstream compression stage of the plurality of compression stages and is pressurized by a bleed stream from the one other compressor. 
         [0012]    The present invention also provides an apparatus for inhibiting a thrust bearing capacity of a compression system from being exceeded during a surge event. In such apparatus a thrust bearing is biased with a biasing force acting in one of two opposite axial directions of a compressor shaft connected to the thrust bearing and to an impeller of at least one compressor having a back disk seal. The compression shaft is driven by a driver to drive the impeller through rotation of the compressor shaft and the compression system has at least one stage of compression provided by the at least one compressor. The thrust bearing has a first thrust bearing capacity and a second trust bearing capacity to absorb axial loads acting upon the thrust bearing in the two opposite axial directions of the compressor shaft, respectively. The thrust bearing is configured such that a first overload margin exists between the first thrust bearing capacity and a first surge force and a second overload margin exists between the second thrust bearing capacity and a second surge force. The first surge force and the second surge force act in the two opposite axial directions and created during the surge event. The back disk seal is sized in a radial direction of the compressor shaft such the biasing force is generated by a force difference during normal operation of the compression system. The biasing force is produced by a force difference between an impeller eye side force and a back disk force exerted on the impeller, opposite to the eye side force, by a high pressure outer annular region and a low pressure inner annular region separated by the back disk seal of the impeller. The force difference is transmitted from the impeller and through the compressor shaft to the thrust bearing and increases one of the first overload margin and the second overload margin such that the first overload margin and the second overload margin are no less than about twenty-five percent of the first thrust bearing capacity and the second thrust bearing capacity, respectively. 
         [0013]    The driver can be an electric motor, the compressor shaft can be a motor shaft protruding from the electric motor and the thrust bearing can be part of the electric motor. Further, the thrust bearing can be one of a magnetic bearing. 
         [0014]    The at least one compressor can comprise a first compressor and a second compressor connected at opposite ends of the motor shaft and in flow communication with each other such that least two successive compression stages are provided in the compression system. The force difference is produced by a difference between the eye side force and the back disk force of at least one of the first compressor and the second compressor. The at least one compressor can be a downstream compression stage of a plurality of compression stages of the compression system. The low pressure region is in flow communication with an outlet or an inlet of one other compressor of an upstream compression stage of the plurality of compression stages and is pressurized by a bleed stream from the one other compressor. 
     
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
         [0015]    While the specification concludes with claims particularly pointing out the subject matter that applicants regard as their invention, it is believed that invention will be better understood when taken in connection with the accompanying drawings in which: 
           [0016]      FIG. 1  is a schematic, sectional view of a compression system in accordance with the present invention that incorporates two compression stages; 
           [0017]      FIG. 2  is a schematic, fragmentary view of  FIG. 1  that illustrates the eye side and back disk forces developed in each of the compression stages; 
           [0018]      FIG. 3  is a schematic, fragmentary view of  FIG. 1  illustrating the areas at which the eye side and back disk forces are developed on an impeller in the first stage of compression; 
           [0019]      FIG. 4  is schematic diagram of the overload margins and the forces to which an impeller is subjected to in a compressor of the prior art; 
           [0020]      FIG. 5  is a schematic diagram of the of the overload margins and the forces to which an impeller is subjected to in a compressor of the present invention; and 
           [0021]      FIG. 6  is a schematic diagram of a compression system of the present invention in which an upstream compression stage is used to help provide the biasing force on the impeller in accordance with the present invention. 
       
    
    
     DETAILED DESCRIPTION 
       [0022]    With reference to  FIG. 1 , a compression system  1  is illustrated having two successive compression stages provided with two compressors  2  and  3 , respectively, that are driven by an electric motor  4 . Compressors  2  and  3  are centrifugal compressors. Compressor  2  compresses a gas, for instance air, from a low pressure to an intermediate pressure and compressor  3  further compresses the gas from the intermediate pressure to a yet higher pressure. Consequently compressor  3  has a higher outlet pressure than compressor  2 . Although not illustrated, compressor  2  would be connected to compressor  3  by a suitable conduit and depending on the application of compression system  1  could incorporate interstage cooling. As will be discussed, however, the present invention has equal application to a compression system having a single compressor. 
         [0023]    Compressor  2  includes a shroud  10  having an inlet  12  and an impeller  14  that is driven by a motor shaft  16  of the electric motor  4 . The gas is driven by impeller  14  into a volute  18  from which the gas is expelled at a higher pressure than the gas entering compressor  2  from inlet  12 . A back face seal  20  is provided to prevent the escape of high pressure gas from the back face of the impeller  14 . As indicated above, the back face seal  20  is a labyrinth type of seal and circles the back of the impeller  14 . A seal holder  22  retains the impeller back disk seal  20  in place. A cavity  24  is formed behind the impeller  14  that lies between the back disk seal  20  and the motor shaft  16 . A shaft seal  26  that provides a seal about the motor shaft  16  thereby also sealing the cavity  24 . Shaft seal  26  is also held in place by seal holder  22 . 
         [0024]    Compressor  3  is provided with a shroud  28  having an inlet  30  that is in flow communication with an outlet (not shown) of the volute  18  of the compressor  2  from which gas is discharged at an intermediate pressure. As indicated above, an interstage cooler could be proved between the outlet of the volute  18  and the inlet  30 . An impeller  32  is driven by the motor shaft  16  of the electric motor  4  at the opposite end thereof to the end at which impeller  14  is driven. The gas is driven by impeller  32  into a volute  34  from the gas is expelled at a higher pressure than the gas entering compressor  3  from inlet  30 . Although not illustrated, a conventional outlet in the volute  34  is provided for discharging the gas at such higher pressure. A back disk seal  36  is provided to prevent the escape of high pressure gas from the back face of the impeller  32 . The back disk seal  36  is supported by a seal holder  38 . A cavity  42  is formed behind the impeller  32  that lies between the back disk seal  36  and the motor shaft  16 . A shaft seal  43  that provides a seal about the motor shaft  16  and therefore also seals cavity  42 . Shaft seal  43  is held in place by a seal holder  38 . 
         [0025]    The shroud  12  is connected to the volute  18  which is in turn connected to one end of a cylindrical motor casing  44 . The shroud  28  is connected to the volute  34  which is in turn connected to a motor cap  45 . Motor cap  45  is connected to the opposite end of cylindrical casing  44 . The electric motor  4  has a rotor  46  attached to motor shaft  16  and a stator  48  attached to the inside of cylindrical motor casing  44 . The motor shaft  16  is supported at opposite ends by journal bearings  50  and  52 . Journal bearings  50  and  52  are active magnetic bearings having conductors  54  and  56  attached to the motor shaft  16  and electromagnets  58  and  60  connected to the cylindrical motor casing  44  and the motor cap  45 , respectively. The journal bearings  50  and  52  electromagnetically suspend the motor shaft  16  for rotational movement. An electromagnetic thrust bearing  62  is also provided. Thrust bearing  62  has a disk-like thrust runner  64  that is a conductor that rotates between inboard and outboard components  66  and  68 , respectively, that are electromagnets that suspend the disk-like thrust runner  64  between the inboard and outboard components. Not illustrated, but as would be well known in the art, gap sensors are provided with associated electronics to differentially power the electromagnets to maintain the gaps between conductors  54  and  56  and the electromagnets  58  and  60  and the disk-like thrust runner  64  between its associated electromagnets of the outboard and inboard components  66  and  68 . The ability to maintain the gaps by active magnetic bearings is not without force limit. This force limit can be exceeded during a surge event. Consequently, as a back up, two sets of bushings  70  and  72  are provided that are connected to the cylindrical motor casing  44  and the motor cap  45  by plates  74  and  76 . During a power loss or upon startup or after a shutdown bushings  70  and  72  will radially support the motor shaft  16 . End elements  78  and  80  connected to motor shaft  16  that are of ring-like configuration to contact the bushings  70  and  72  should the axial force on the motor shaft  16  exceed the capability of the thrust bearing  62 . This can occur during a surge event and such axial forces imparted to motor shaft  16  through bushings  70  and  72  can be particularly severe. As such the number of surge events that the motor bushings  70  and  72  can be subjected to will be limited to a small number of events. 
         [0026]    During operation of the compression system  1 , the axial forces acting on the motor shaft  16  and therefore, the thrust bearing  62  are imparted as a result of a balance of axial forces on each of the impellers  14  and  32 . With additional reference to  FIG. 2 , the impellers are subjected to an inwardly directed eye-side forces, shown as “F 1 ” and “F 2 ”, respectively, that arise as a result of the gas passing through the inlets  12  and  30 . Opposing these forces are back disk forces “F 3 ” and “F 4 ”. These back disk forces arise from the high pressure gas produced by each of the impellers  14  and  32  escaping into the high pressure, outer annular, outer annular regions  82  and  84  that lie between the outer rims  86  and  88  of the impellers  14  and  32  and their associated back disk seals  20  and  36 , respectively. Additionally, the cavities  24  and  42  are also pressurized by low pressure gas through back disk pipes  90  and  92  (shown in  FIG. 1 ) and hence forces are exerted in the resulting low pressure, inner annular regions  94  and  96  of the impellers  12  and  32  that lie between the back disk seals  20  and  36  and the motor shaft  16 , respectively. This low pressure gas can be obtained from the inlets  12  and  30  using connections  98  and  100 , respectively. Alternatively, low pressure gas can be obtained from another source such as the inlet to any upstage compressor or in case of air compression, possibly the ambient. 
         [0027]    The foregoing can be better understood with reference to  FIG. 3  in which the high pressure, outer annular region  82  and the low pressure, inner annular region  94  are shown with respect to the back disk force “F 3 ”. As can be appreciated, the back disk force “F 3 ” is given by the equation: 
         [0000]        F 3=π*[( R 0 2   −R 1 2 )* P   d +( R 1 2   −R 2 2 )* P   c ];
 
         [0028]    Where the high pressure, outer annular region  82  is the area between R 0  and R 1 , the high pressure gas is P d  escaping from the outer rim  86  at R 0  into such region and the low pressure, inner annular region  94  is the annular area between R 1  and R 2  and the low pressure gas in cavity  24  is P c . “F 1 ” is the eye side force equal to the summation of all pressures existing on the impeller eye side and is given by the equation: 
         [0029]    F 1  is determined by an integration of the pressure field across the inlet face of the impeller  14  by well known computer modeling techniques that can be carried out with the use of commercially available computer programs such as ANSYS CFX computational fluid dynamics software that can be obtained from Ansys, Inc. of Southpointe, 275 Technology Drive, Canonsburg, Pa. 15317, United States of America. 
         [0030]    With reference to  FIG. 4 , in the prior art as shown for the normal operation, in the case of impeller  14 , the radii R 1  defining the inner radius of the high pressure, outer annular region  82  (R 0  and R 1 ) and the outer radius of the low pressure, inner annular region  94  (R 1  and R 2 ) is set so that the net impeller thrust during normal operation is set to an acceptably low level or zero. In other words, if zero, F 1 −F 3 =0. The same approach would be used in the case of impeller  32  and consequently, the net impeller thrust during normal operation would also be set to acceptably low level or even zero. In other words, if zero, F 2 −F 4 =0. Since compressor  1  has two impellers  14  and  32 , impeller back disk seals  20  and  36  with desired radii with respect to the impeller or the axis of the motor shaft are set such that the net thrust from both impellers during normal operating conditions is set to an acceptably low level or even zero. In other words, if zero, F 1 −F 3 +F 4 −F 2 =0. In a compressor with a single impeller, the net through for such impeller would be set to zero in the manner indicated above for each of the impellers  14  and  32 . 
         [0031]    With continued reference to  FIG. 4 , compressors are, however, subject to abnormal operating conditions and the thrust bearing must accommodate these safely. One such abnormal operating condition is surge which produces two alternating, asymmetric forces on the thrust bearing  62 . The first phase (Surge Phase 1) is a sudden loss of discharge pressure arising from aerodynamic instability in the impeller. This results in a decrease in the eye side pressure. However, the back disk force, for instance F 3  for impeller  14 , remains high due to fluidic system inertial effects. This results in a large axial force towards the inlet, for instance inlet  12  with respect to impeller  14  which in such first phase acts in the same direction as force F 3 . The second phase (Surge Phase 2) occurs as the high pressure back disk pressure bleeds off, dramatically reducing the back disc force F 3 , while simultaneously the flow stabilizes and pressure rebuilds in the eye side and increases the eye side force F 1 . This results in a large axial force shown as Surge Phase 2 acting in the direction of F 1 . It is to be noted that during the surge event, both impellers  14  and  32  will be subjected to these oscillating forces because they are interconnected and in fluid communication. Hence, in  FIG. 4  the solid lines shown during the surge events represent a summation of the forces acting on the thrust bearing  62  and eventually the bushings  70  and  72  when the forces exceed the thrust bearing capacity of the thrust bearing  62  to resist such forces. In  FIG. 4 , the dashed lines represent such thrust bearing capacity (F capacity(+)  and F capacity(−) ), beyond which the bushings  70  and  72  will be contacted. The “Positive Overload Margin” and the “Negative Overload Margin” represent differences between such load capacities in positive and negative force directions and the maximum forces generated during the surge event in both positive and negative force directions, namely F Phase 1  and F Phase 2 . The same considerations would apply to a compressor employing a single impeller, but without the summation of forces due to both impellers. 
         [0032]    The purpose of the thrust bearing  62  is to prevent these alternating forces from damaging the compressor. When these thrust forces exceed the thrust bearing capacity, F capacity(+)  and F capacity(−)  of the thrust bearing  62 , the end elements  78  and  80  connected to motor shaft  16  contact the back-up bushings  70  and  72  which are designed to protect the compressor during such event. However, such bushings have a limited life when contacted. As illustrated, the thrust bearings of the prior art can be designed with overload margins, positive and negative, that are designed to adsorb such alternating forces during surge events. However, in the case of oil-free bearings such as magnetic bearings, the load capacities are lower and these overload margins are not high and may not exist. Consequently, as discussed above, such bearings have found limited use. As could be appreciated by those skilled in the art, the same discussion would be equally applicable to air foil bearings which also have limited overload capacities. 
         [0033]    It is to be noted that the axial forces during Surge Phase 1 and Surge Phase 2 can be obtained on an operating compression system by direct measurement or from indirect measurements and subsequent calculation, and or can be obtained during the design phase of a compression system from geometry and expected thermodynamic operating conditions. Direct measurement of axial force F phase 1  and F phase 2  on an operating compressor system using magnetic bearings is available by monitoring the magnetic thrust bearing control currents during surge and knowing the force-current coefficient. In-direct measurement and subsequent calculation of F phase 1  and F phase 2  on an operating compressor is performed by measuring thermodynamic conditions within the compression system and using known compressor geometry to compute F phase 1  and F phase 2  axial loading. During the design phase, detailed compressor aerodynamic designs provide expected thermodynamic conditions and actual geometry present that allow calculation of F phase 1  and F phase 2 . The computation of F phase 1  and F phase 2  using geometry and either expected or measured thermodynamic conditions follow the same computational process used by those skilled in the art to calculate the impeller eye side forces F 1  and F 4  and back disk side forces F 2  and F 3 . 
         [0034]    In accordance with the present invention, contrary to the prior art, the thrust bearing is preloaded by a force that will increase one of the overload margins at the expense of the other overload margin to in turn increase the ability of the thrust bearing to resist a surge event. With reference to  FIG. 5 , as an example, it is assumed that compressors  2  and  3  will experience a system surge such that the F phase 1  will be greater than F phase 2 . In accordance with the present invention, to achieve a greater positive overload margin in Surge Phase 1 of the prior art shown in  FIG. 4 , the back disk seal radii for impellers  14  and/or  32  are adjusted to create a non-zero preload thrust value for the normal operation of compressor  1 . As shown in  FIG. 5 , a preload force F preload  at normal operating conditions will increase the positive overload margin in Surge Phase 1 of the surge event where such increase is required at the expense of decreasing the negative overload margin in Surge Phase 2 of the surge event. As can be appreciated without the preload, the system shown in  FIG. 5  would have a dangerously small Positive Overload Margin without the preload force because the force generated in phase 1, F phase 1  would otherwise be closer to the positive thrust bearing capacity F capacity(+) . It is to be noted that had the Negative Overload Margin been less than the Positive Overload Margin, before providing a preload, then the preload force F Preload , would have been exerted in the opposite direction or in other words, toward the positive or F Capacity(+) . In any event, the preload force F preload  should be sufficient that both the positive overload margin and the negative overload margin is no less than twenty-five percent of the positive and negative thrust bearing capacities F capacity(+)  and F capacity(−) . As can be appreciated, even higher overload margins are desirable and if possible 50 percent and greater. It is to be noted, however, as the overload margins are increased, it is also possible that the thrust bearing capacities must also be increased to accommodate the increased margins. In such case, the entire thrust bearing  62  must be made larger and there exists a limit on the diameter of the thrust runner  64  that will maintain structural integrity at high speeds. 
         [0035]    One possible way to produce a preload force in the direction of F 3  during normal operation is by decreasing radii R 1  on impeller  14 , thus, increasing the high pressure outer annular region  82  of the back disk, while decreasing the inner annular low pressure region  94 . The result is to create a larger value of F 3  which acts in the negative direction such that a preload force F preload  will now be present in the negative direction at normal operation of the compressor  1  as per shown in  FIG. 5 . It is particularly preferred to adjust back seal diameter R 1  to achieve equal positive and negative overload margins if consequence of an overload are the same in either direction. 
         [0036]    Although the foregoing preloading has been discussed with respect to manipulating the seal design with the appropriate sizing of R 1  for impellers of either or both compressors  1  and  2 , the preload force can also be adjusted by the pressure acting on the low pressure, inner annular region by the use of a bleed a stream of low pressure gas from an upstream stage and apply it to the low pressure, inner annular region. With reference to  FIG. 6 , a multistage compression system l′ is illustrated having a compression system  1  as has been described above and having compressors  2  and  3 . A feed stream  100  is compressed by compressor  2  to produce a compressed stream  102  at an intermediate pressure. Compressed stream  2  is cooled in an intercooler  104  and then further compressed in compressor  3  to produce a compressed stream  106 . Compressed stream  106  is cooled in an intercooler  108  and then introduced into further compressor  5  for further compression. Compressor  5  is driven by an electric motor  6  of similar design to the electric motor  4  described above and such compressor  5  is also provided with seals of the type discussed with respect to compression system  1  and compressors  2  and  3 . The resultant output compressed stream  110  can be cooled by an after-cooler  112 . A bleed stream  114  of low pressure gas for the compression step  5  back disk cavity could be obtained from stream  102  as illustrated or stream  106 . This low pressure bleed stream is fed into a back disk pipe associated with compressor  5  to pressurize its low pressure region behind the impeller to at least assist in the preloading of the compressor  5  in accordance with the present invention. Another benefit of selecting the lowest low pressure stream for the cavity  104  is that the back disk seal radii R 1  can be reduced and consequently less flow of gas will occur across this labyrinth seal. A reduction in seal flow will produce an increase in compression system efficiency. In other words, the smaller the radius of the back disk seal, the less seal area there is for high pressure gas to leak across. 
         [0037]    It is to be noted that although the present invention has been discussed with respect to oil-free bearings such as electromagnetic and air foil bearings, the present invention could advantageously be applied to compression systems having conventional bearings lubricated by oil to provide an added margin of safety in case of a surge event. Furthermore, it is not only electrical driven compressors that are subject to surge events, but also, other types of compressors that are driven by other drive mechanisms, for instance, steam or expansion turbines and therefore the present invention in its most broad aspects is not limited to compressors driven by permanent magnet electric motors as illustrated in the Figures and that can be controlled by variable frequency drives. It is also to be noted, that although the present invention has been described with respect to compressors driven at opposite ends of a motor shaft, the present invention would have equal application to a compression system having a single stage of compression. 
         [0038]    Additionally, a compression system with a single compressor at one end and an expansion turbine drive or stage at the other end would be a valid application of this invention. This arrangement, known in the industry as a turbocompressor or turbocharger, typically would not include an integral high speed motor. However, it is clear to one skilled in the art that this invention and this turbocompressor or turbocharger embodiment could also include an integral high speed motor, generator, or bidirectional motor/generator. 
         [0039]    As will occur to those skilled in the art, although the present invention has been described with reference to preferred embodiments, numerous changes and omissions thereof can be made without departing from the spirit and scope of the present invention as set forth in the appended claims.