fig3 shows a refrigeration circuit in accordance with an embodiment of the present invention that might be used , for example , to generate chilled water for a building air conditioning system . in a preferred form , the operating parameters of the system might be designed to provide sufficient heat output to heat a supply of low pressure hot water for the building , in addition to the chilled water for the air conditioning system . a refrigerant fluid circulates around the refrigeration system , which operates on a vapour - compression refrigeration cycle ( see fig1 ). the refrigerant enters a two - stage centrifugal compressor 10 in a superheated vapour state and is compressed , in two steps ( lower pressure stage 1 and higher pressure stage 2 ) to a higher pressure and temperature . superheated refrigerant vapour from the second stage compressor discharge outlet passes to a condenser 12 , where it is cooled , the superheat and latent heat of condensation is removed and the vapour condenses to a liquid , then cooling further to a sub - cooled liquid state . the sub - cooled liquid refrigerant output from the condenser is next passed through an expansion device ( e . g . expansion valve 14 ), where it undergoes an abrupt reduction in pressure , resulting in an adiabatic flash evaporation of a proportion of the liquid refrigerant . this dramatically lowers the saturation temperature of the remaining liquid refrigerant , resulting in a liquid and vapour mixture . the refrigerant then passes to an evaporator 16 , through which the secondary fluid to be chilled flows ( typically water or air ). the liquid part of the cold liquid / vapour refrigerant mixture evaporates and in so doing removes heat energy from the circulating secondary fluid to chill it . the saturated refrigerant vapour absorbs further heat energy to become a superheated refrigerant vapour . from the evaporator outlet the superheated refrigerant vapour then returns to the compressor inlet and the cycle starts again . in this example , an economiser is used to improve the efficiency of the compressor and the refrigeration system that it serves . a significant portion of the total liquid refrigerant mass flow leaving the condenser 12 enters the primary circuit of the economiser heat exchanger 18 prior to the primary expansion device . the remaining refrigerant liquid enters a second expansion device ( e . g . expansion valve 20 ) and flows through the secondary circuit of the same economiser heat exchanger 18 , where it evaporates at a lower saturation temperature than the temperature of the primary liquid refrigerant flow , thus sub - cooling the major portion of refrigerant liquid before it enters the primary expansion device 14 . the economiser secondary outlet superheated vapour flow is re - introduced to the main refrigerant vapour flow between the two impeller compression stages of the compressor 10 . the evaporating pressure and temperature within the economiser sits at an intermediate point between the evaporating pressure at the inlet stage of the first stage impeller 22 and the discharge pressure at the outlet of the second stage impeller 24 . conveniently , the pipework 26 connecting the outlet of the first stage 22 to the inlet of the second stage 24 provides an easily accessible location to introduce this economiser refrigerant vapour flow . the economiser flow returning into the compressor intermediate stage performs significant de - superheating of the main flow of refrigerant , improving overall efficiency of the heating / cooling cycle . in accordance with the present invention , as illustrated in fig2 , and as discussed further below , the radial bearings 32 and thrust bearings 34 that support and restrain the compressor shaft 28 are hydrodynamic vapour bearings in which the bearing fluid is the circulating refrigerant vapour itself . in the example illustrated here , the refrigerant vapour is taken from tappings 34 , 36 in the stage 2 impeller casing and passed through conduits ( schematically shown in fig2 with dashed lines ) machined in the compressor housing to the bearings 32 , 34 , to supply vapour to the bearings under pressure . the bearings 32 , 34 are therefore a hybrid between hydrostatic and hydrodynamic bearing types , and might conveniently be referred to as “ hybrid bearings ”. in other embodiments the vapour supply for the bearings may be taken from the stage 1 impeller casing , a combination of tappings in the stage 1 and 2 impeller casings , or from an external vapour source . in the example shown in fig2 separate pressure tappings 34 , 36 are used for the radial bearing 32 and the thrust bearing 34 , in order that the thrust bearing 34 can be supplied with the refrigerant vapour at a higher pressure . the compressor shown in the exemplary system of fig3 , and discussed in greater detail below , is a two - stage compressor , with the impellers 22 , 24 for the two - stages mounted at opposite ends of the drive shaft 28 , the two impellers 22 , 24 being oriented in opposite axial directions ( back to back ) to help minimise the axial thrust forces via self - cancellation . the concepts of the invention can also be applied , however , to other compressor configurations , examples of which are shown in fig4 to 7 . fig4 shows a single - stage configuration . fig5 shows a two - stage configuration with both impellers 51 , 52 mounted at the same end of the compressor shaft and axially oriented in the same direction . fig6 shows a three - stage configuration with the stage 1 and 2 impellers 61 , 62 , 63 at one end of the compressor shaft , facing in the same axial direction , and the stage 3 impeller at the other end of the shaft facing in the opposite direction , so it is back to back with the stage 2 impeller . fig7 shows a four - stage configuration with the stage 1 and 2 impellers 71 , 72 at one end of the shaft and the stage 3 and 4 impellers 73 , 74 at the other end . the stage 1 and 2 impellers 71 , 72 face in the opposite direction to the stage 3 and 4 impellers 73 , 74 . the skilled person will appreciate that numerous other compressor configurations are possible within the scope of the invention . referring again to the system of fig3 , as well as the enlarged view of the compressor of that system shown in fig8 , the components of the system and some possible variants of the system will now be described in more detail . the compressor is a two - stage centrifugal design with an overall pressure ratio of , in this example , about 19 : 1 . the design point rotational speed was selected at 21000 rpm to give the best compromise between keeping the stage 1 inlet tip relative mach number to a manageable level ( 1 . 3 ) while minimising the impeller diameters . the compressor is sized to achieve 400 kw of refrigeration , with the incorporation of an economiser circuit . as a result , the second stage compressor is designed for a refrigerant mass flow approximately 50 % higher than the first stage . the overall shaft power consumption of the compressor is 224 kw at the maximum flow condition . the resulting heat of rejection of 624 kw in this example can be rejected to atmosphere so usefully employed to heat a secondary fluid ( water or air ). other variants of the compressor , either as a single - stage or multi - stage centrifugal design will operate at lower compression ratios than 19 : 1 . these compression ratios will in some embodiments be in the order of 3 : 1 - 5 : 1 and will be applied to various refrigerants in order to provide the highest operating efficiency . the range of evaporation saturation temperatures and condensing saturation temperatures that in turn define these compression ratios might , for example , be : accordingly , the operational saturation temperatures will render the compressor suitable for use in systems that serve sub - zero refrigeration applications , process cooling , and data centre cooling , comfort and process air conditioning , and numerous other applications . moreover , the range of condensing saturation temperatures proposed will permit this compressor design with its attendant technical and efficiency features to be applied not only in northern europe , but also in higher temperature regions such as parts of the usa , japan , middle east , far east and australasia where certain refrigeration and air conditioning systems are unable to function unless the compressor is able to develop a sufficiently high condensing saturation temperature , whilst maintaining the required evaporation saturation temperature at a sufficiently low level . a ) twin centrifugal compressor impellers 22 , 24 ( sometimes referred to as “ wheels ”) operating in series . for some applications a single impeller may be adequate . b ) “ hybrid ” bearings 32 , 34 ( radial and thrust ) which allow the shaft 28 carrying the centrifugal compressor wheels 22 , 24 to rotate at very high speeds ( circa 60 , 000 rpm , or as much as 100 , 000 rpm or even 200 , 000 rpm or more ) without the need for lubricating oil . 99 % of the world &# 39 ; s air conditioning & amp ; refrigeration systems utilise oil to lubricate the compressors and this oil is inadvertently and unnecessarily transported around the refrigeration system on a continuous basis . this is due to the fact that all compressors continually pump out a proportion of their lubricating oil charge along with the recirculating system refrigerant . this oil impedes the heat transfer efficiency of heat exchangers , consumes a proportion of the compressor &# 39 ; s electrical input energy thus reducing overall system efficiency , requires careful sizing of refrigeration pipework to ensure adequate velocities for oil return to the compressor and often requires oil cooling , oil separators , oil pumps , oil heaters , oil level controls and other components to ensure satisfactory operation . the oil free design of the compressor of the present invention thus has a major impact on the efficiency , first cost and ongoing maintenance of such systems . a key advantage of hybrid bearings over an alternative low friction bearing ( magnetic bearings ) is that in the event of an electrical power failure , the shaft can simply be allowed to come to rest without the need to provide an auxiliary and temporary power supply , as is the case with magnetic bearings . more importantly , the hybrid type bearings require no electrical energy and therefore this compressor type will be more efficient in this particular regard . c ) hydrostatic bearing vapour injection — in order to reduce wear and to extend the life of the hydrostatic bearings , the wear that occurs when starting and stopping shaft rotation must be addressed . consequently , the exemplary system can employ a refrigerant vapour injection system to pass vapour into the bearings in order to “ float ” the shaft prior to start - up and this will also apply to the thrust hydrostatic bearings . the refrigerant will preferably be finely filtered prior to entering the bearings in order to protect them from system particles / contaminants . the vapour injection system might also be employed to maintain bearing temperatures within reasonable limits , should the minimal friction generated cause unacceptable temperature rise . d ) high efficiency motor design — the majority of compressor drive motors operate at an efficiency of approximately 80 %- 90 % and the waste heat energy is absorbed by the refrigerant vapour passing over the motor windings and then through the compressor mechanism in the majority of ac & amp ; r system compressors throughout the world . the motor efficiency envisaged for embodiments of the present invention will preferably be in the order of up to 98 . 5 % and consequently only 1 . 5 % heat rejection will be passed to the refrigerant vapour running over the motor . this results in a smaller condenser , lower condenser secondary fluid flow ( typically air or water ), smaller fan motors and water pumps resulting in lower electrical power input , and higher overall system efficiency , otherwise known as coefficient of performance ( cop ). f ) inverter control — a custom electronic inverter is envisaged to drive the compressor motor at the required speed and to vary the motor speed in order to regulate refrigerant mass flow through the evaporator and refrigeration system in order to match the cooling capacity or heating capacity to the instantaneous cooling or heating load . this results in precision temperature ( or evaporating pressure ) control and minimises energy consumption . moreover , when the system is operating at less than maximum speed / mass flow / cooling - heating capacity , the evaporator and condenser become ‘ oversized ’. this increases their heat exchange capability whilst dramatically improving compressor efficiency and system cop as a result of lower compression ratio demand . consequently , the increase in system efficiency is exponential at reducing capacity levels . the pressure drop through the expansion device is also reduced under these conditions , thus further improving compressor efficiency and system cop . the inverter will preferably also afford the compressor motor protection against the following : g ) typical compression ratios — typical compression ratios in conventional refrigeration and air conditioning systems are in the order of 3 : 1 - 5 : 1 and a wide family of refrigerants is used ( in conjunction with suitable mineral or synthetic lubricating oils ) to meet the required operating saturation temperature at the evaporator for cooling / refrigeration and at the condenser to suit the secondary fluid temperature ( e . g . ambient air or water ) into which the refrigeration system heat must be rejected . the range of refrigerants now available is extremely wide and includes hfc &# 39 ; s , hfo &# 39 ; s , co2 , air , hydrocarbons , ammonia and others . the primary application of the exemplary compressor described here is targeted at a compression ratio of around 20 . 1 : 1 , this being substantially higher than any prior art centrifugal compressors . h ) the impellers 22 , 24 are mounted on either end of the shaft 28 in a back - to - back arrangement . this has three main benefits : the axial thrust loads are partially balanced , as the thrust from the first - stage impeller 22 acts in the opposite direction to the second stage impeller 24 . this minimises the load on the thrust bearing 34 ; variable inlet guide vanes (‘ vigvs ’) can easily be incorporated in both stages , whereas if the wheels were mounted on the same end of the shaft , it would be virtually impossible to find the space for vigvs for the second stage ; and the mixing of the vapour from the economiser can be easily achieved — complex porting arrangements will not be needed . the system of fig3 has been initially designed to handle the most arduous temperature range envisaged , that being : a ) cooling a secondary medium such as air or water ( or another liquid ) to a temperature of approximately 5 ° c ., this requiring a refrigerant evaporating temperature and equivalent saturation pressure of about 0 ° c . b ) heating a medium such as air or water ( or another liquid ) to a temperature of approximately 80 ° c ., this requiring a condensing temperature and equivalent saturation pressure of approximately 90 ° c . many other applications can thus be served by variants of the system , these typically operating between − 5 ° c . ( or lower ) and 60 ° c . ( or higher ) evaporating / condensing temperatures . moreover , this compressor can also be applied as a single compressor or as a multiple compressor arrangement such that one or more compressors serve the requirements of the same refrigeration system / circuit , thereby providing a degree of reserve capacity , standby capacity and / or greater flexibility in variable capacity control to match the thermal load requirement . such compressors can be run singly or mutually , preferably with operating speed and inlet guide vane angle adjustment in harmony so as to ensure both compressors develop the same compression ratio in order to maintain refrigeration system stability . the control and protection system has been designed to suit both single and multiple compressor applications . the air conditioning and refrigeration system industries are served by a wide range of fluids known as refrigerants . these are specifically selected to suit the application requirements and the equipment is then specifically designed to provide optimal performance with the selected fluid at the required operating parameters . embodiments of the present invention can be adapted to suit a wide range of application operating parameters and a wide range of operating fluids . these fluids / refrigerants comprise either a single fluid ( azeotrope ) or a mixture of 2 - 3 fluids ( zeotropes ) and the following lists some , but not all , of the fluids that embodiments of the invention can be adapted to function with : hfcs are man - made chemicals containing the element fluorine used predominantly as refrigerants and aerosol propellants . they are colourless , odourless and chemically non - reactive gases . they are primarily being used as replacements to ozone damaging cfcs and hcfcs . hc &# 39 ; s are organic compounds consisting entirely of hydrogen and carbon , used predominantly as refrigerants and aerosol propellants . in some cases they are readily absorbed into the lubricating oil , thereby reducing its lubricating properties , potentially leading to compressor failure . compressors in accordance with embodiments of the invention will surmount this problem by avoiding the use of lubricant oils . a new family of refrigerants currently undergoing introduction to the industry . ( as at 2012 / 2013 ) embodiments of the present application are suited to a great variety of different applications including refrigeration , air conditioning and heat pump systems . the concepts of the present invention can be adapted to operate with alternative fluids by simply varying physical size / shape / properties of the following components : as shown in fig1 , the electronic inverter 1010 is used to speed control and protect the motor 82 ( fig8 ). in addition , however , a control and protection system 1020 is provided for the compressor 10 . this can also control and protect the refrigeration system 1030 as a whole . the control system will typically have inputs of system pressures , temperatures and power input . it may , for example , control one or any two or more of : motor speed , inlet guide vane angle , the thermostatic expansion valves ( that admit refrigerant to the evaporator and economiser of the refrigeration circuit ), and condenser fan speed ( where the condenser rejects heat to atmosphere and fans are used to assist ) to maintain the compressor within desired operating parameter ranges and preferably also to maximise system efficiency and thermal performance . the expansion valves may also be controlled to help protect the compressor from the entry of liquid refrigerant . one possible approach to enhanced control and protection is to use a variant of the commercially available climacheck ™ system . the climacheck ™ system is conventionally used to analyse system performance using existing system pressure , temperature and power input sensors ( and can be used in this manner in embodiments of the present invention whether or not it is used as part of the system control ). the climacheck ™ technology enables the performance of an air conditioning or refrigeration system to be analysed to within an accuracy of 5 - 7 % of reality . the following parameters can be determined on a dynamic basis at 1 s , 2 s , 5 s , 10 s , 30 s , 60 s and 5 minute intervals and may be made available to an operator instantaneously or recorded for future interrogation : climacheck ™ also gives advance warning of refrigerant loss , unstable system operation , and operation outside of safe system limits . emails / text messages can be automatically sent to various recipients under such circumstances . business edge limited is the distributor of climacheck ™ throughout the uk . the core plc ( programmable logic controller ) that resides within climacheck ™ can also be programmed to become a control and protection device and can also reside on the same network as climacheck ™. consequently , the readings derived from temperature and pressure sensors , current transformers and so on can be utilised both for climacheck ™ performance analysis and for the “ control and protection ” of the compressor and refrigeration system . in some embodiments it may be desirable to use a separation vessel in the system prior to the compressor inlet . an example of such a vessel , which might be termed a “ suction accumulator / filter vessel ”, is shown in fig1 . the main purpose of this vessel 1110 is to act as an emergency reservoir upstream of the compressor 10 so as to trap any liquid “ slug ” that might develop in the suction line leading to the compressor 10 . this , for example , might arise where the expansion valve 14 controlling the flow of liquid refrigerant into the evaporator 16 malfunctions or if such a valve 14 is incorrectly set with too low a superheat value . another application for this vessel 1110 is where the compressor might be utilised within a heat pump system — when such systems reverse cycle , excess liquid can often be directed towards the compressor 10 . here the combined suction accumulator / filter 1110 will provide sufficient volume to receive such slugs of liquid . when this occurs , the compressor 10 continues to run drawing vapour only from the vessel 1110 so as to protect the centrifugal wheels 22 , 24 of the compressor , whilst progressively evaporating the liquid into vapour form prior to the compression process . in such an arrangement , the compressor and the system in which it is contained might employ use of a bottom connection 1120 at the base of the suction accumulator so as to allow any liquid accumulation to be directed elsewhere , either for re - use within the system or for other purposes directly related to the function of the compressor . a secondary function of the suction accumulator / filter vessel 1110 is to provide a high degree of filtration at the inlet to the compressor 10 . where any particulate matter within the refrigeration system might exist or might develop either as a result of component malfunction or as a result of poor system maintenance or installation , the high grade filter will ensure that pure vapour only passes into the inlet of the first stage centrifugal wheel 22 of the compressor so as to protect the wheels 22 , 24 and the journal / thrust plate assemblies 32 , 34 . in a further application of the suction accumulator / filter vessel , the vapour flows from the compressor hybrid bearings and / or vapour from other parts of the refrigeration system can be connected to this component so as to act as a “ vapour collection point ” to both protect the first stage impeller from any liquid present in these vapour lines and to arrest any particulate matter that would otherwise enter the compressor , thus avoiding both impeller and bearing damage . where any particulate matter containing iron is concerned , a magnetic capture device could be included within pipework or within a suction accumulator / filter vessel upstream of the compressor of the present invention or indeed any type of compressor . the incidence of such iron - containing particulate matter is likely to be very low in the refrigeration circuit ; however , any ferrous materials that might be within the system itself at the outset of construction could be captured in this way . referring briefly to fig1 and 13 , there is shown a diagrammatic overview of a remote monitoring and control system for overseeing an hvac system 1100 which includes a compressor , as described above . sensors 1110 - 1118 are typically located about building 1150 and might measure local room temperature or external ( ambient ) temperature . each sensor 1110 - 1118 provides an input into control system 1160 , such as the one that is sold under the climacheck ™ brand , and which includes a microprocessor 1160 operating under control of proprietary software ( optionally modified by a user ) and bespoke algorithms . a transceiver 1170 is connected to the microprocessor 1160 and is adapted to communicate to mobile communication devices 1180 , via ‘ gps ’ systems , as well as to conventional fixed computer terminals 1190 , for example , via the internet ( cloud ) or conventional telecommunication networks . referring briefly to fig1 there is shown a flow diagram of typical menus that are provided to a maintenance engineer in order to verify the status or update the condition of an air - conditioning unit or water heater / chiller ( not shown ) into which a temperature signal is provided automatically from the sensors 1110 - 1118 . in response to a request or in order to perform a check , a user has to enter certain location specific data , passwords and authentication data ( such as user id codes ) in order to be able to gain access to one of several controllable systems , via a data link . these systems range from a local scale , such as the temperature of a room or floor of building 1150 , to a larger scale , such as start - up of an entire water treatment plant or hvac system for an entire building or complex of buildings , for example a shopping mall or office block . once access to the system controller has been gained , and subject to safety overrides and other failsafe systems , a user inputs desired updates or new temperature conditions and these are translated into commands to local valves or to the compressor . an acknowledgement of the command or new setting is transmitted from the transceiver 1170 at the control system and confirms receipt of the instruction and updating of particular conditions / parameters . fig1 illustrates an exemplary operating cycle for a compressor system in accordance with an embodiment of the present invention . in this example , the system includes a vapouriser arrangement for supplying the hybrid bearings of the compressor with refrigerant vapour at start - up and shut - down to float the bearings at times when the rotational speed of the impeller drive shaft is not fast enough for the hydrodynamic bearings to be self - supporting . as discussed in more detail below , in this example the vapouriser arrangement has a vapouriser chamber for generating saturated vapour and a superheat chamber for superheating the saturated vapour prior to it being fed to the bearings . the cycle will be described with reference to the operating points labelled in the figure with numbers in circles : 1 . a “ start ” command in response to control system instruction “ vapouriser energised ”. at this point the vapouriser chamber is generating saturated vapour . 2 . superheater energised — the saturated vapour from the vapouriser chamber is superheated in the superheat chamber . 4 . once the bearings are floated , the drive motor starts and accelerates centrifugal vapour compressor wheels to “ pre - set speed ” at which the bearings are self - sustaining ( self - acting spool for hybrid bearings ). pressure from one or more tapping points on one or more impeller wheels is also available at this operating point to provide additional support to the journal / thrust bearings . 5 - 8 vapouriser recharged with liquid refrigerant in readiness for next operation . 7 - 8 adjusting speed within this range to control refrigerant mass flow developed by impeller wheels of compressor in order to attain setpoint target of refrigeration system by matching refrigeration capacity to instantaneous thermal load . vigv angle is also adjusted at one or more centrifugal wheel inlets in conjunction with impeller wheel speed , to attain fine refrigerant mass flow control but at peak operating efficiency , whilst avoiding “ surge ” or “ choke ” conditions . 11 . superheated vapour injection to journal / thrust bearings to support bearings as impeller speed drops below speed at which bearings are self - sustaining . 12 . drive motor decelerating to “ pre - set speed ” ( self - supporting ) ( pressure from impellers no longer able to supply bearings ) fig1 to 17 show exemplary possible configurations for a vapouriser arrangement that can be used , as in the exemplary operation cycle discussed above , to provide a supply of refrigerant vapour to float the radial and / or thrust bearings at slow speeds , for example during start - up and shut - down , especially at times where the rotational speed of the impeller drive shaft is insufficient for the bearings to be self - supporting . in the arrangement seen in fig1 , liquid refrigerant in a vapourisation chamber (“ vessel ‘ a ’”) can be rapidly heated by a heater (‘ htr ’) in the chamber to generate saturated vapour (‘ sv ’) at an elevated pressure . this vapour is then fed , under its own pressure , to the radial and thrust bearings of the compressor through vapour supply channels (‘ sat - vs / sht - vs ’). the opening of the vapour outlet within the vapourisation chamber is elevated above the base of the chamber to ensure that no liquid refrigerant can flow into the vapour supply channels . the level of the liquid refrigerant in the chamber is monitored with a liquid level sensor (‘ lls ’), a supply of liquid refrigerant to the chamber being controlled , in response to the liquid level sensor output , using an electrically operated valve (‘ eov ’). further electrically operated valves (‘ eov ’) are used to open and close the vapour flow from the chamber to the bearings and also to open and close a flow of vapour from one or more of the impellers of the compressor to the bearings . in this example , heaters (‘ htr ’) are associated with each bearing to elevate the temperature of the bearing above the saturation temperature of the refrigerant to prevent potentially damaging condensation forming in the bearings . the heaters are driven in this example by thyristor control . temperature sensors (‘ ts ’) are used to sense bearing temperature , the signals from these sensors being used by the main control system to control the bearing heaters and to control the switching on of the vapour supply from the vapouriser at start - up . in the vapouriser arrangement illustrated in fig1 , during start - up and shut - down the bearings are fed with a supply of saturated vapour from the vapouriser chamber . it may be more preferable , however , to feed superheated vapour to the bearings as this further reduces the likelihood of the vapour condensing in the bearings . fig1 shows a vapouriser arrangement to achieve this , in which saturated vapour from the vapourisation chamber (“ vessel ‘ a ’”) is supplied to a superheating chamber (“ vessel ‘ b ’”), in which it is superheated by further heater (‘ htr ’) in this second vessel . a temperature sensor (‘ ts ’) in the superheat chamber is used to control the heater to ensure the vapour leaves for onward passage to the bearings in a superheated state . fig1 shows a further example of a possible arrangement for providing the bearings with a supply of vapour . in this example , rather than the vapouriser chamber having a liquid refrigerant feed , it is instead supplied with refrigerant vapour from a high pressure region of the refrigerant circuit ( e . g . compressor outlet ) during operation of the compressor . the vapour is then stored in the chamber until required to float the bearings , e . g . at shut - down , at which point it is further superheated and elevated in pressure by a heater (‘ htr ’) in the chamber and fed to the bearings . fig1 shows , for a compressor in accordance with an embodiment of the invention , one example for a configuration of channels formed in the compressor housing through which refrigerant vapour can be supplied from one or more impeller wheels of the compressor , or from a vapouriser arrangement , to the radial and thrust bearings of the compressor . fig1 a is a longitudinal cross section through one end of the compressor . fig1 b is a sectional view along line a - a in fig1 a and fig1 c is a sectional view along line b - b in fig1 a . looking first at the feed to the radial bearing , vapour passes from a supply ( i . e . impeller or vapouriser ) through a radial bore in the compressor housing to an annular channel that surrounds the outer bearing shell of the radial bearing . a series of radial drillings extend through the outer bearing shell to extend from this annular channel to the inner bearing surface of the outer bearing shell , so that vapour flows under pressure from the annular channel through these drillings into the gap between the bearing surfaces of the inner and outer bearing shells . the vapour is then discharged to either end of the bearing gap . the arrows in fig1 a and 18 c indicate the vapour flow through the bearings . the same radial supply bore supplies refrigerant vapour to both sides of the thrust bearing via passages that branch off the radial bore . more specifically , these passages supply vapour to drillings that are circumferentially spaced at intervals around the thrust plates on either side of the thrust bearing at the end of the impeller drive shaft . the arrows in fig1 a and 18 b indicate the flow of vapour through the bearing . while the invention has been described in conjunction with the exemplary embodiment described above , many equivalent modifications and variations will be apparent to those skilled in the art when given this disclosure . accordingly , the exemplary embodiment of the invention set forth above is considered to be illustrative and not limiting . various changes to the described embodiment may be made without departing from the spirit and scope of the invention .