GB2566675A

GB2566675A - Turbocharger

Info

Publication number: GB2566675A
Application number: GB1714107.8A
Authority: GB
Inventors: Rajkumar Batham Tushar; Vohra Ankit
Original assignee: Cummins Ltd
Current assignee: Cummins Ltd
Priority date: 2017-09-04
Filing date: 2017-09-04
Publication date: 2019-03-27
Anticipated expiration: 2037-09-04
Also published as: GB2566675B; GB201714107D0

Abstract

A turbocharger comprises a compressor and a volume 29 behind the compressor impeller 7, which is in selective fluid communication with a location 37 downstream of the compressor via a connecting channel 47. A moveable element 53 is operable to block the channel 47 when the pressure at the downstream location 37 is sufficiently large. This arrangement ensures an adequate pressure in the volume 29 to prevent lubricant leakage, whilst maintaining efficiency by only bleeding compressed air when needed because the pressure in the volume is low, e.g. during low speed and low pressure operation.

Description

Turbocharger

The present invention relates to a turbocharger.

Turbochargers are well known devices for supplying air to the intake of an internal combustion engine at pressures above atmospheric pressure (boost pressures). A conventional turbocharger comprises an exhaust gas driven turbine wheel mounted on a rotatable shaft within a turbine housing. Rotation of the turbine wheel rotates a compressor wheel mounted on the other end of the shaft within a compressor housing. The compressor wheel delivers compressed air to the intake manifold of the engine, thereby increasing engine power.

The turbocharger shaft is conventionally supported by journal and thrust bearings, including appropriate lubricating systems, located within a central bearing housing connected between the turbine and compressor wheel housing. It is desirable to provide an effective sealing arrangement at each end of the rotating shaft to prevent oil leakage from the central bearing housing into the compressor or turbine housing.

At the compressor end of the turbocharger, during normal operation, the sealing arrangement should be able to withstand the increasingly high boost pressures that are delivered by modern turbochargers. The pressure of the bearing housing is effectively the same pressure as the engine oil sump (typically around 100 millibar) and there is thus a pressure gradient between the bearing housing and the compressor housing which prevents the leakage of lubrication oil from the bearing housing into the compressor housing. The sealing arrangement typically comprises one or more ring seals arranged between the shaft and the bearing housing and received in respective grooves, in the manner of piston rings. The seals are arranged with a radial clearance so as to allow the passage of gas in small volumes across the seals but to restrict the flow to allow the pressure difference between the bearing housing and the compressor housing and/or the turbine housing to remain.

At low speeds such as when the engine is idling or during exhaust braking, the low (or even negative) boost pressure at the compressor end can drop below the pressure in the bearing housing. As a result of the low pressure difference over the seals (PDOS), oil is able to leak along the turbocharger shaft in the bearing housing, past the seals and into the compressor housing. The leakage of oil may be significant, for example, if idling or operating in engine braking mode occurs for extended periods of time. The reduction in rotational speed of the shaft means that the pressure behind the compressor wheel decreases which exacerbates the tendency of the oil to travel along the shaft to the compressor housing. Leakage of oil into the compressor housing is undesirable as it contaminates the pressurised air entering the engine intake manifold.

It is one object of the present invention to obviate or mitigate the aforesaid disadvantage.

According to a first aspect of the invention there is provided a turbocharger comprising: a turbine; a housing;

a compressor having a compressor impeller mounted within the housing on a shaft;

a volume located behind the compressor impeller, the volume being in selective fluid communication with a location within the compressor which is downstream of the compressor impeller via a connecting channel which extends between the volume and the location downstream of the compressor impeller;

a moveable element being provided at an inlet of the connecting channel, the moveable element having an outer face; wherein in an open configuration the moveable element allows the downstream location to fluidly communicate with the volume, in a closed configuration the moveable element closes the connecting channel thereby preventing the downstream location from fluidly communicating with the volume;

wherein the moveable element is biased to the open configuration by a biasing member;

wherein the configuration of the moveable element is determined by a pressure level at the downstream location relative to a pressure level in the volume; and wherein the moveable element is configured to move from the open configuration to the closed configuration when the pressure level at the downstream location is sufficiently high to provide a force sufficiently large to overcome the force arising from the pressure in the volume and a biasing force exerted by the biasing member.

The invention has a number of advantages. The invention provides a simple and low cost way of reducing the leakage of oil into the compressor. The moveable element moves to a closed configuration once a required pressure is reached in the volume. High pressure air is not bled from the compressor when not required. The system is therefore highly efficient as the system generally seals the pressurised air within the volume, thereby increasing the pressure level in the volume.

The moveable element acts passively and therefore does not require external power. The moveable element is actuated based on differences in pressure and does not therefore require any sensing or other complex components. Minimal, if any, maintenance is required to keep the invention working.

The outer face of the moveable element may align flush with a wall of the housing when the moveable element is in the closed configuration.

Flush alignment is beneficial because when the moveable element is in the closed configuration, compressed air flow may be substantially unaffected by the moveable element. This may increase the efficiency of the turbocharger. By flush alignment it is meant that there is no step between the outer face of the moveable element and the wall of the housing.

The outer face of the moveable element may have a generally circular cross-section.

A generally circular cross-section may be a cross-sectional shape which is generally, but not entirely, circular. A generally circular cross-section is beneficial because the moveable element may be able to rotate and still close. As such, a moveable element having a generally circular cross-section may not require an alignment feature to restrict rotation. A generally circular cross-section is beneficial because alternative, non-circular geometries may require an alignment feature to restrict rotation which would otherwise prevent the moveable element from closing. The alignment feature may add increased cost, complexity and maintenance requirements. Furthermore, generally circular cross-section geometries may be simpler to manufacture.

Alternatively, the outer face of the moveable element may be one of a range of noncircular geometries. A non-circular geometry may require an alignment feature to restrict rotation of the moveable element in use. The outer face of the moveable element may be shaped so as to take advantage of the pressure increase radially outward of the compressor impeller axis of rotation. For instance, the outer face of the moveable element may be wedge-shaped. The outer face of the moveable element may have a constant thickness. Alternatively, the outer face of the moveable element may have a variable thickness.

The moveable element may be a mechanical plunger.

The use of a mechanical plunger as the moveable element is beneficial due to the simplicity of the mechanical plunger. Due to this simplicity, cost and maintenance requirements are reduced.

The turbocharger may further comprise an O-ring arranged to seal the moveable element against the housing when the moveable element is in the closed configuration.

An O-ring is beneficial due to the simplicity of the sealing solution. Furthermore, Orings are known to be suitable for use in the challenging environment found inside a turbocharger including high temperatures and the presence of contaminants in the flow. O-rings are also simple to install and replace as required. Other sealing solutions may, however, be used. For instance, one or more face seals could be used.

The ratio of the cross-sectional area of the outer face of the moveable element to the cross-sectional area of an inner face of the moveable element may be between 1.1:1 and 1.6:1.

The outer face of the moveable element refers to the same outer face of the moveable element which may align flush with a wall of the housing when the moveable element is in the closed configuration. The moveable element outer face cross-sectional area may influence the pressure at which the moveable element moves from the open configuration to the closed configuration, or vice versa. The moveable element having the ratio of the cross-sectional area of the outer face to the cross-sectional area of the inner face in the above range may be beneficial because the moveable element may move from the open configuration to the closed configuration, or vice versa, once the flow reaches a desired pressure level.

The biasing member may be a coil spring. A coil spring is advantageous because it is a low cost and simple component.

The coil spring may have a spring constant of less than around 6 N/mm.

The coil spring may have a spring constant of at least around 3 N/mm.

A coil spring having a spring constant in the above range is beneficial because the spring provides a sufficiently high force to bias the moveable element to the open configuration. Furthermore, the spring constant is such that the force exerted by the spring may be overcome by the force exerted on the outer face of the moveable element by the pressure of the flow. Multiple coil springs may be used in place of a single coil spring. Multiple coil springs may be used in order to obtain the desired spring constant value. Multiple coil springs may be connected in parallel or in series.

The connecting channel may include a change of direction.

A change of direction in the connecting channel is beneficial because it may influence the pressurised air which flows through it. The connecting channel may be used to recover static pressure by reducing a dynamic head of the pressurised air flow. This may be achieved by, for example, incorporating bends or other direction changing features into the connecting channel. The recovered static pressure may increase the static pressure behind the compressor impeller. A change of direction of the connecting channel may also make the connecting channel easier to manufacture because, for example, machining two shorter bores which meet at one end may be easier than machining a single, longer bore.

A valve seat of the connecting channel may include a stepped profile.

A stepped profile may be beneficial because it is simple to manufacture. Furthermore, a stepped profile is advantageous for locating and/or constraining the moveable element during use.

A smallest cross-sectional area of the connecting channel may be at least around 2 mm².

A smallest cross-sectional area of the connecting channel may be less than around 20 mm².

Smallest cross-sectional area of the connecting channel is intended to mean a throat of the connecting channel. That is to say, the smallest cross-sectional area of the connecting channel will be at the point of the smallest diameter and so the narrowest part of the connecting channel.

The ratio of a circumferential flow path area of the compressor to a smallest crosssectional area of the connecting channel may be between 150:1 and 410:1.

The ratio of the circumferential flow path area of the compressor to the smallest crosssectional area of the connecting channel may be between 200:1 and 360:1.

Specific embodiments of the present invention will now be described, by way of example only, with reference to the accompanying drawings in which:

Figure 1 is a cross-sectional side view of a turbocharger assembly according to an embodiment of the invention;

Figure 2 is an enlarged view of part of a compressor assembly of the turbocharger assembly of Figure 1;

Figure 3 shows a moveable element of the turbocharger assembly of Figure 1 and part of the compressor assembly of Figure 2 in more detail;

Figure 4 is an enlarged view of part of the compressor assembly of Figure 1 in a first operating condition, with the moveable element in an open configuration;

Figure 5 is an enlarged view of part of the compressor assembly of Figure 1 in a second operating condition, with the moveable element in a closed configuration;

Figure 6 is a graph which shows simulation data for a prior art turbocharger;

Figure 7 is a graph which also shows simulation data for a prior art turbocharger;

Figure 8 is a graph which shows conditions within a compressor stage of a prior art turbocharger plotted using simulation data; and

Figure 9 is a compressor map of a prior art turbocharger.

Figure 1 is a cross-sectional side view of a turbocharger according to a first aspect of the invention. The turbocharger comprises a turbine 1 joined to a compressor 2 via a bearing housing 3. The turbine 1 comprises a turbine housing 4 and a turbine impeller 5. Similarly, the compressor 2 comprises a compressor housing 6 and a compressor impeller 7. The turbine impeller 5 and compressor impeller 7 are mounted on opposite ends of a shaft 8 which is supported on roller bearing assemblies 9 and a thrust bearing assembly 16 within the bearing housing 3. The roller bearing assemblies 9 support a predominantly rotational load whilst the thrust bearing assembly 16 supports a predominantly axial load. Although a fixed geometry turbocharger is shown in Figure 1, the invention is equally applicable to a variable geometry turbocharger.

The turbine housing 4 is provided with an exhaust gas inlet 10 and an exhaust gas outlet 11. The exhaust gas inlet 10 directs incoming exhaust gas to an annular inlet volute 12 surrounding the turbine impeller 5. The exhaust gas flows through the turbine 1 and out of the exhaust gas outlet 11 via a circular outlet opening which is coaxial with the turbine impeller 5. Rotation of the turbine impeller 5 rotates the compressor impeller 7 which draws in air through axial inlet 13 and delivers compressed air to the engine intake via an annular outlet volute 14. The turbine impeller 5, shaft 8 and compressor impeller 7 are co-axial and rotate about a turbocharger axis 33.

The bearing housing 3 provides a lubricating system for the turbocharger assembly. The bearing housing 3 includes a series of channels 15, 15a-d, through which oil is supplied to the roller bearing assemblies 9 and thrust bearing assembly 16. A generally horizontal manifold 15a opens into a thrust bearing assembly oil line 15b and first and second roller bearing assembly oil lines 15c, 15d respectively. The channels 15, 15a-d receive oil from an engine oil circuit (not shown).

The bearing housing 3 also includes a turbine sealing assembly 25 and a compressor sealing assembly 24. The sealing assemblies 24, 25 seek to prevent oil leaking out of the bearing housing 3 and into either the turbine 1 or the compressor 2. Compressor sealing assembly 24 comprises an oil seal plate 17, an oil slinger 18 and an oil seal ring 18a. The oil seal ring 18a is located in a groove of the oil slinger 18 to reduce the leakage of oil from the bearing housing 3 into the compressor 2.

Oil leakage into the compressor 2 is particularly undesirable because any leaked oil may be drawn into the engine intake via the annular outlet volute 14 along with the compressor inlet flow. Oil leakage into an air intake can be very damaging to an engine. The sealing assemblies 24, 25 therefore seek to isolate the bearing housing oil from the turbine 1 and compressor 2. In general, the sealing assemblies 24, 25 are not complete sealing arrangements and so do not entirely prevent the leakage of oil from the bearing housing 3 into the compressor 2 and/or turbine 1. This is due, at least in part, to the relative movement of components as a result of the rotation of the shaft 8, compressor impeller 7 and turbine impeller 5.

As discussed in the paragraph above, it is undesirable for oil to leak out into either the turbine 1 or compressor 2 from the bearing housing 3. Also as discussed above, it is particularly dangerous for oil to leak into the compressor 2, as the compressor 2 supplies the engine with compressed air. Supplying the engine with compressed air mixed with leaked oil can damage an engine.

One way to reduce the leakage of oil from the bearing housing 3 into the compressor 2 is to ensure that a pressure on a compressor-side of the compressor sealing assembly 24 is higher than a pressure on a bearing housing-side of the compressor sealing assembly 24. Such a difference in pressure should reduce the leakage of oil from the bearing housing 3 into the compressor 2.

In order to obtain a pressure differential, it is known in the art to provide pressure to a volume 29 behind the compressor impeller 7. The term “behind the compressor impeller” is defined as a side of the compressor impeller 7 opposite blades 7a of the compressor impeller 7. The volume 29 is located on the compressor-side of the compressor sealing assembly 24. In the illustrated embodiment, the volume 29 is defined by a back face 27 of the compressor impeller 7, the bearing housing 3, the oil seal plate 17 and the oil slinger 18. Alternatively, the volume 29 may be defined by a combination of one or more of the back face 27 of the compressor impeller 7, the bearing housing 3, the compressor housing 6, the shaft 8, the thrust bearing assembly 16 or the roller bearing assembly 9. The volume 29 may be any of a range of sizes such that pressure can practically be maintained within the volume 29. The volume 29 may have a continuous cross-section. Alternatively, the volume 29 may vary in crosssection.

A clearance exists in the form of a flow channel 36 between a portion of the bearing housing 3 and the compressor impeller 7 such that the volume 29 is not a sealed volume. The flow channel 36 also provides a clearance which is needed to allow the compressor impeller 7 to rotate. Fluid may flow through the flow channel 36 and, as a result, the volume 29 is capable of being pressurised by air compressed by the compressor impeller 7 via the flow channel 36.

The flow channel 36 comprises a recess 31 and an annular stepped outer portion 35. The recess 31 is a revolved cut, a profile of which is revolved about the turbocharger axis 33 into the bearing housing 3. The recess 31 forms an annular ledge if viewed from the compressor-side of the turbocharger axis 33. The annular recess 31 is shaped such that an amount of the air compressed by the compressor impeller 7 impacts an annular stepped outer portion 35, changes direction and flows between the annular recess 31 and the back face 27 of the compressor impeller 7 (via flow channel 36), into the volume 29. This flow of air into the volume 29 increases the pressure in the volume 29. The increase in pressure in the volume 29 reduces the chance of oil leaking from the bearing housing 3 into the compressor 2. This is achieved by creating a more positive pressure difference over the seals (PDOS) effect.

There are certain operating conditions wherein the oil pressure on the bearing housing side of the compressor sealing assembly 24 can exceed the pressure in the volume 29. Such conditions increase the likelihood of oil leakage into the compressor 2 from the bearing housing 3. The conditions under which oil leakage into the compressor 2 from the bearing housing 3 is most likely are low speed and low pressure ratio operations, including choked operation. Furthermore, an increased oil pressure in the bearing housing 3 due to blow-by or crank case gases can exacerbate the leakage of oil into the compressor 2. High engine RPM operation can also have the adverse effect of increasing the oil pressure in the bearing housing 3, thereby increasing the risk of oil leakage into the compressor 2 from the bearing housing 3. These conditions are as discussed below.

Low speed operation occurs when the turbocharger rotates at a low number of rotations per minute (RPM). An example of a low number of RPM may fall within the range of 5k to 25k (i.e. 5,000 to 25,000) RPM. While operating at these low speeds, the compressor impeller 7 may not be able to generate a pressure in the volume 29 which is higher than the oil pressure in the bearing housing 3. However, even when the turbocharger is rotating at a low RPM, including idle, condition (i.e. 5k to 25k RPM), the compressor impeller 7 is imparting work on the air and is therefore providing some air to the volume 29. The main problem at low speed operation is that, although air is pressurised in the compressor housing 6, air is not sufficiently pressurised in the volume 29. Accordingly, the volume 29 may not be sufficiently pressurised to prevent the leakage of oil from the bearing housing 3.

Low pressure ratio operation occurs when the turbocharger is operating in the choke condition. Pressure ratio is defined as the ratio of the annular outlet volute 14 pressure to the axial inlet 13 pressure. Choke occurs when the resistance to flow in the annular outlet volute 14 drops below normal levels. As a result of the low resistance, the annular outlet volute 14 flow experiences a very low back pressure. Accordingly, the low back pressure leads to an increase in the compressor output as the gas velocity increases in the compressor 2. The gas velocity can only increase until a near-sonic flow velocity is reached, when the flow approaches incompressibility. At this point, the compressor 2 operates inefficiently and with high annular outlet volute 14 temperatures (up to around 230°C). In such conditions, the compressor 2 flowrate is high but the increase in pressure between the axial inlet 13 and the annular outlet volute 14 is comparatively low, due to the inefficient operation of the compressor 2 in the choke condition.

Choke may be defined as an operating condition whereby compressor efficiency reduces below a certain value, for instance 60% efficiency, for pressure ratios lower than a certain value, for example 1.5. A visual representation of operation in the choke condition is highlighted in a compressor map illustrated in Figure 9.

Choke is undesirable for a number of reasons including low compressor efficiency, increased risk of component damage and excessively high operating temperatures. Choke is also undesirable because operating in the choke condition can give rise to low PDOS which can lead to the leakage of oil and other lubricant from the bearing housing 3 into the compressor 2. Operating in the choke condition can also cause bearing failure, for instance thrust bearing failure. When the turbocharger is operating under choke conditions, the turbocharger can be said to be “choked”.

To summarise, operating a turbocharger under choke conditions means the compressor 2 is operating in an inefficient band. The low efficiency operation of the compressor 2 results in a low pressure ratio. The pressure ratio is defined as the pressure of the air exiting the compressor 2 via the annular outlet volute 14 divided by the pressure of the air entering the compressor 2 via the axial inlet 13. As a result of the low pressure ratio, the pressure of the air which has been compressed by the compressor impeller 7 is at a comparatively lower pressure than it should be if operating under normal conditions. As this comparatively lower pressure air is the air which is supplied to the volume 29, it follows that the pressure in the volume 29 is also lower than desired. There is therefore a risk that the insufficiently pressurised volume 29 may not reduce the leakage of oil from the bearing housing 3 into the compressor 2 due to an insufficient PDOS.

The leakage of oil from the bearing housing 3 into the compressor 2 during both low speed and low pressure ratio operating conditions may be reduced or eliminated by selectively introducing pressurised air into the volume 29. Where this is done, the pressure is increased in the volume 29 and this substantially prevents the leakage of oil from the bearing housing 3 into the compressor 2.

The turbocharger assembly of Figure 1 includes a connecting channel 47 and valve assembly 38. The connecting channel 47 and valve assembly 38 selectively provide the volume 29 with compressed air from the compressor 2, when required. The connecting channel 47 and valve assembly 38 are shown in greater detail in Figures 25 and are discussed in detail below.

Figure 2 is an enlarged view of part of a compressor assembly of the turbocharger assembly of Figure 1.

Figure 2 includes the connecting channel 47 and valve assembly 38. The connecting channel 47 comprises two flow passages in fluid communication with one another: a valve bore 41 and a connecting bore 43. The connecting channel 47 further comprises a valve seat 39. The illustrated connecting channel 47 extends between an annular flow passage 37 and the volume 29. Alternatively, the connecting channel 47 may extend from another location downstream of the compressor impeller 7 to the volume 29.

The annular flow passage 37 is defined between the compressor housing 6 and a wall 45 of the bearing housing 3. As air passes through the compressor 2, it is first compressed by the compressor impeller 7 and then flows radially outward via the annular flow passage 37 before exiting the compressor 2 via the annular outlet volute 14. The axial gap between the wall 45 of the bearing housing 3 and a radial wall 6a of the compressor housing 6 may for example be around 3.8 mm, around 5 mm or around 10 mm. The axial gap may be between 3 mm and 15 mm (or some other size, depending upon the compressor impeller 7 geometry). If the axial gap were to be smaller than 3 mm there may not be sufficient room to accommodate opening of a moveable element 53 (described below).

The connecting channel 47 and valve assembly 38 selectively supply the volume 29 with high pressure air from the annular flow passage 37. The connecting channel 47 selectively provides fluid communication between the volume 29 and the annular flow passage 37.

In the illustrated embodiment, the valve seat 39, valve bore 41 and connecting bore 43 are all blind bores. They may be machined by, for instance, drilling into the bearing housing 3. The bores may feature a fillet, in contrast to the sharp corners as depicted, resulting from the drilling process. Alternatively, the bores may be manufactured by the use and subsequent removal of an intermittent destructible mould. For example, the bores could be manufactured using a lost-wax method. Alternatively, the bores may be manufactured as part of an additive manufacturing process. The bores may have a constant cross-section. Alternatively, the bores may have a non-constant crosssection. Alternative arrangements of the connecting channel 47 may be used. For example, the flow passages which make up the connecting channel 47 could be replaced with a single, angled flow channel. The connecting channel 47 may be machined into the compressor housing 6 instead of the bearing housing 3. This may be used in turbochargers which include a larger compressor housing 6 which extends into the region occupied by the bearing housing 3 in the illustrated embodiment.

The valve assembly 38 is positioned to allow the connecting channel 47 to be closed. The moveable element 53 of the valve assembly 38 allows fluid communication between the annular flow passage 37 and the volume 29 via the connecting bore 43 when in an open configuration. In a closed configuration, the moveable element 53 of the valve assembly 38 prevents the annular flow passage 37 from fluidly communicating with the volume 29 via the connecting bore 43. In Figure 2, the moveable element 53 is in the closed configuration, thereby preventing the annular flow passage 37 from fluidly communicating with the volume 29 via the connecting bore 43. In this position, the volume 29 is not pressurised by air flowing through the connecting channel 47 and valve assembly 38.

The valve assembly 38 is schematically depicted in Figure 3. The valve assembly 38 comprises the moveable element 53 and a spring 54. The moveable element 53 includes an outer face 49 (not visible in Figure 3). In some embodiments, when the moveable element 53 is in the closed configuration, the outer face 49 may align flush with the wall 45 of the bearing housing (see Figure 2). By flush alignment it is meant that there is no step between the outer face 49 and the wall 45 of the bearing housing 3. The outer face 49 has an area A! upon which a force is exerted by pressure in the annular flow passage 37. The moveable element 53 also has an inner face 51, the inner face 51 being annular in shape. The inner face 51 has an area A₂ upon which a force is exerted by pressure in the volume 29. The inner face 51 includes a groove 57, into which an O-ring 59 is located. When the moveable element 53 is in the closed configuration, the O-ring 59 contacts the valve seat 39 (see Figure 2) so as to seal the connecting channel 47 from the annular flow passage 37. Thus, the connecting channel 47 does not pressurise the volume 29 when the moveable element 53 is in the closed configuration. Alternative sealing arrangements may be used instead of locating the O-ring 59 in the moveable element 53. For instance, the O-ring 59 could be located in a groove in the bearing housing 3.

A valve stem 55 extends from the inner face 51 and is cylindrical in the illustrated embodiment. The valve stem 55 may be of an alternative geometry, such as cuboidal. Further alternatively, the valve stem 55 may include one or more constraining features to prevent rotation, such as a keyway. Where a cylindrical geometry is used for the valve stem 55, there may be no need to incorporate a constraining feature to prevent rotation of the valve stem 55 and so moveable element 53. This is dependent upon the use of a suitable geometry of the outer face 49 of the moveable element 53 (e.g. a circular outer face may be used).

The spring 54 attaches to the moveable element 53 at an outer face 55a of the valve stem 55. Alternatively, the spring 54 may attach directly to the inner face 51 of the moveable element 53. The method of attachment of the spring 54 is not important to the invention and is not considered in detail.

The spring 54 is just one possible example of a biasing member which could be used in the valve assembly 38. For instance, an alternative biasing member such as a diaphragm or pressurised chamber could instead be used. An alternative elastic member, for instance made of rubber, could also be used. A range of types of spring could also be utilised. For instance, a coil spring could be used. Other types of spring, such as a wave spring or a disc spring, could also be used.

As well as providing an attachment location for the spring 54, the valve stem 55 also constrains the moveable element 53. The valve stem 55 cooperates with a portion of the valve bore 41, when located in situ, to form a sliding linear guide. The valve bore 41 constrains the valve stem 55 (and the moveable element 53) such that the valve stem 55 and the moveable element 53 move substantially in only an axial direction. As explained above, the valve stem 55 may or may not permit rotation of the moveable element 53 depending on geometry of the valve stem 55 and/or the outer face 49 of the moveable element 53.

The function of the spring 54 is to bias the moveable element 53 outward from the valve seat 39, located in the bearing housing 3, as described in more detail below.

Although the illustrated embodiments depict a circular outer face 49 and annular inner face 51, these selections are non-limiting and alternative geometries may be used. For example, alternative outer face geometries which utilise a static pressure gradient which exists in the annular flow passage 37 may be desirable. A static pressure in the annular flow passage 37 is at its lowest value just radially outward of the compressor impeller 7. This is because a dynamic pressure of the flow is highest as the flow passes the compressor impeller 7, due to the compressor impeller 7 imparting work on the flow. The static pressure then increases to a highest value at the annular outlet volute 14. This is due, at least in part, to the recovery of dynamic head of the flow into static pressure. Between the two locations, a static pressure gradient exists. Particularly of note, a static pressure gradient therefore exists in the annular flow passage 37. The static pressure gradient exists in the annular flow passage 37 such that the static pressure increases radially outwards from the turbocharger axis 33. Figures 6-8 and accompanying description below provide further examples of the pressure gradients in a prior art compressor.

The static pressure gradient may influence the choice of geometry of the outer face 49 of the moveable element 53. For instance, it may be desirable to select an outer face 49 geometry with a greater distribution of area further radially outward from the turbocharger axis 33. Geometries such as a wedge or inverted triangle may be used. Such geometries could be used to influence the moveable element 53 behaviour with regard to when the moveable element 53 transitions from the open configuration to the closed configuration, or vice versa. Where the outer face 49 is circular or generally circular, it may not be necessary to constrain the moveable element 53 to prevent rotation. This is because the moveable element 53 is able to rotate and still close. Where the outer face 49 geometry is not circular, it may be necessary to constrain the moveable element 53 to prevent rotation. This may be achieved by, for instance, a keyway. The keyway may interact with a groove in the bearing housing 3 to rotationally constrain the moveable element 53. Alternatively, the keyway may interact with a groove in the compressor housing 6 to rotationally constrain the moveable element 53.

Figure 4 depicts the moveable element 53 in the open configuration, and Figure 5 depicts the moveable element in the closed configuration. Both Figures 4 and 5 also show the flow of air through the compressor 2 and the connecting channel 47, where appropriate.

Figure 4 depicts an enlarged view of part of the compressor assembly of Figure 1, with the compressor 2 operating in a low RPM (including idle) or low pressure ratio condition (pressure ratios of, for example, between 1.1 and 1.5). As previously explained, during low RPM and low pressure ratio operating conditions the compressor impeller 7 is rotating, is imparting work on the air and is therefore increasing the pressure of the air which enters the compressor 2 via the axial inlet 13. Air is compressed by the compressor impeller 7 and then flows radially outwards via the annular flow passage 37. Air then exits the compressor 2 via the annular outlet volute 14. There may be a flow of pressurised air which enters the volume 29 via the flow channel 36. However, the air flowing through the flow channel 36 is unable to sufficiently increase the pressure in the volume 29 so as to prevent the leakage of oil from the bearing housing 3 into the compressor 2. This is due to the compressor operating conditions which may include low RPM or low pressure ratio conditions. Low RPM (including idle) may be 5k to 25k RPM.

Figure 4 depicts the moveable element 53 biased away from the bearing housing 3. In this position, the inner face 51 does not sealingly engage the valve seat 39.

The spring 54 resiliently biases the moveable element 53 outwards towards the open configuration. The moveable element 53 is therefore biased away from the bearing housing 3 as the result of force arising from the pressure difference between the annular flow passage 37 and the volume 29, and the force exerted by the spring 54. When the moveable element 53 is biased away from the bearing housing 3, the pressure in the volume 29 may initially be similar to the pressure in the annular flow passage 37. Any increase in pressure at the annular flow passage 37 is not fully communicated to the volume 29 via the connecting channel 47 due to the constricted flow passage. As the compressor 2 increases in speed, the moveable element 53 remains in the open configuration until the force from the air pressure in the annular flow passage 37 is sufficiently high so as to overcome the force exerted by the spring 54 and the force from the pressure in the volume 29. Eventually the force balance is such that the moveable element 53 sealingly engages the valve seat 39. At the point whereby the moveable element 53 sealingly engages the valve seat 39, the pressure in the volume 29 is sufficiently high to substantially prevent the leakage of oil from the bearing housing 3 into the compressor 2. In other words, a sufficiently high PDOS is achieved and leakage of oil from the bearing housing 3 into the compressor 2 is reduced or prevented. A desirable PDOS value may be around 60 mbar. The force balance of the moveable element 53 is discussed in greater detail below. Also of note, the outer face 49 area A! is greater than the inner face 51 area A₂.

Figure 5 depicts an enlarged view of part of the compressor assembly under desired PDOS conditions. In other words, Figure 5 depicts part of the compressor assembly wherein the pressure of the volume 29 is sufficiently high so as to substantially prevent the leakage of oil from the bearing housing 3 into the compressor 2.

Figure 5 depicts the moveable element 53 sealingly engaging the valve seat 39. This occurs when the compressor impeller 7 is rotating at a desired operating condition (i.e. not low RPM [including idle, i.e. 5k to 25k RPM], or low pressure ratio) and at a sufficiently high speed such that the volume 29 is pressurised via air flowing through the flow channel 36.

Under positive PDOS conditions, as shown in Figure 5, the annular flow passage 37 is therefore sealed from the connecting channel 47. By isolating the annular flow passage 37 from the connecting channel 47 when the pressure in the volume 29 is sufficiently high, the efficiency of the compressor 2 can be improved. Compressor efficiency is improved by preventing the flow of pressurised air from the annular flow passage 37 into the volume 29, via the connecting channel 47, when not required. The pressurised air is, instead, supplied to the engine intake from the compressor 2 via the annular outlet volute 14. This is achieved whilst still providing a mechanism to increase the PDOS when required, thereby reducing the leakage of oil into the compressor 2 from the bearing housing 3.

Particularly of note, Figure 5 depicts the outer face 49 of the moveable element 53 as being aligned flush with a wall 45 of the bearing housing 3. This is beneficial because, under desirable PDOS conditions, the pressurised air flowing through the annular flow passage 37 is substantially unaffected by the presence of the moveable element 53. In other words, when in the closed configuration, the moveable element 53 does not significantly affect the flow velocity, flow direction or flow regime (to name just some flow characteristics) of the compressor flow stream.

Another significant benefit provided by the present invention is that by pressurising the volume 29 by way of the connecting channel 47 and valve assembly 38, the annular stepped outer portion 35 of the flow channel 36 can be removed from the turbocharger. As explained further above, the flow channel 36 is located between the back face 27 of the compressor impeller 7 and the bearing housing 3. Removing the annular stepped outer portion 35 may provide an increase in turbocharger efficiency of the order of around 1-2%. In existing turbochargers, the annular stepped outer portion is required to intrude into the flow stream leaving the compressor impeller. This is so that a portion of the pressurised air, having left the compressor impeller, flows through the flow channel and pressurises the volume. It is noted that a clearance may remain to allow the compressor impeller 7 to rotate and, as such, the flow channel 36 may remain in some form. However, the invention permits the flow channel 36 to be reduced in size, thereby increasing the efficiency of the turbocharger. Implementing the invention may therefore allow the design of the flow channel 36 to be optimised for turbocharger efficiency without a compromise being required by a stepped partition. The invention allows the geometry of the flow channel 36 to be optimised without the need to have flow from the compressor impeller 7 impact the annular stepped outer portion 35 to direct the flow into the volume 29.

The flow channel 36 places the volume 29 in permanent fluid communication with the annular flow passage 37. The flow channel 36 is therefore also a source of leakage of pressurised air from the volume 29 back into the compressor 2. Exchanging the existing flow channel 36 for the connecting channel 47 and valve assembly 38 gives rise to an improvement in efficiency as the connecting channel 47 and valve assembly 38 only bleed pressurised air from the annular flow passage 37 when required. When not required, the moveable element 53 of the valve assembly 38 moves to a closed configuration and stops removing air from the annular flow passage 37. This air can, therefore, be utilised in the engine instead of being used in the volume 29.

The pressure of the compressed air flowing through the annular flow passage 37 exerts a force F₁ over the area A₁ of the outer face 49 of the moveable element 53. The force exerted by the annular flow passage air urges the moveable element 53 toward the closed configuration. It is noted that the force F exerted by a fluid at a pressure P over an area A is equal to:

F = P * A

The force urging the moveable element 53 towards the closed configuration is 5 therefore equal to:

Fi = Ρ,Α,

Similarly, the compressed air in the volume 29 and the connecting channel 47 exerts a force over the area A₂ of the inner face 51 of the moveable element 53 when the moveable element 53 is in the closed configuration. The force exerted by the air in the volume 29 and connecting channel 47 urges the moveable element 53 toward the open configuration. The force F₂ exerted over the area A₂ by the pressure P₂ of the air urging the moveable element 53 towards the open configuration is equal to:

F₂ = F₂4₂

When the moveable element 53 is stationary and in the closed configuration, such that it is not sliding about the valve bore 41, and initially ignoring the forces exerted on the moveable element 53 by the spring 54, it must be the case that:

Ti >F₂ and therefore:

P-iA-l > P₂A₂

As well as the pressure forces exerted on the faces of the moveable element 53, the spring 54 also exerts a force on the moveable element 53. The spring 54 urges the moveable element 53 away from the valve seat 39, into the open configuration. The moveable element 53 is urged toward the open configuration to ensure that the annular flow passage 37 is in fluid communication with the volume 29 via the connecting channel 47 until turbocharger conditions are such that the volume 29 is sufficiently pressurised. The spring 54 should therefore exert sufficient force upon the moveable element 53 to ensure the moveable element 53 moves to the closed configuration when the volume 29 is sufficiently pressurised.

It is noted that springs have a characteristic spring constant k, a variable having a value measured in Newtons per metre (/Vm^_1). In accordance with Hooke’s law, the force required to extend a spring of spring constant k by x metres is:

F = k * x

To compress a spring by the same x metres, the same magnitude of force is required but in the opposite direction. With a spring either extended or compressed, the spring will exert a restoring force to return to its original length.

Returning to the earlier force expression, when the moveable element 53 is in the 10 closed configuration it is the case that:

PiA^ 2: P2A2

Now including the spring force into the above yields:

ΡιΑγ > P₂A₂ + kx

Rearranging the above:

kx < P₁A₁ — P₂A₂

The spring 54 can then be selected based on the pressure generated when the compressor 2 is operating in a given condition. For example, the spring 54 may be selected based on the pressure generated when the compressor 2 is operating in a low RPM, including idle, condition. Alternatively, the spring 54 may be selected when the compressor 2 is operating in any one of a number of typical conditions. Under the idle condition, the pressure in the annular flow passage 37 is approximately equal to the pressure in the connecting channel 47 and the volume 29. With the length x by which the spring 54 is compressed being dictated by geometry of the valve assembly 38, the following equation can be used to estimate a value for the spring constant k\

Ί Π/-Λ Λ \ ^_ A₂) kx < PiAi - A₂) -> k < 25

Implementing the above equation in a real world example, for a moveable element having an outer face 49 diameter of 25 mm and a valve stem 55 diameter of 14 mm, the outer face area A₁ and inner face area will be 4.91 E-4 m² (i.e. 0.000491 m²) and 3.37E-4 m² respectively. In practice, A will be smaller than the inner face area value if the moveable element 53 incorporates the O-ring 59. This is because pressure from the volume 29 will only act upon a portion of the inner face 51 of the moveable element 53 (A₂) as some of the area of the inner face 51 will be radially outside the O-ring 59 and thereby sealed from the volume 29. For simplicity, for the purposes of this worked example, 4₂ is assumed to be equal to the inner face area. For a pressure value P of

1.288E5 Pa, and for a spring extension x of 5 mm, the spring constant k can be calculated as follows:

1.288//5(4.91// - 4 - 3.37£ - 4) '-5-k < 3.97 N/mm

A suitable value of spring constant k for this example would therefore be 3 N/mm. Also for this example, a suitable axial gap between the wall 45 of the bearing housing 3 and the radial wall 6a of the compressor housing 6 would be 5 mm or more. This is to allow axial clearance for the moveable element 53 to extend by 5 mm. The closing force required to close the moveable element 53 would be 15 N in this instance.

Implementing the above spring constant equation in a different real world example (for a larger frame turbocharger), for a moveable element having an outer face 49 diameter of 50 mm and a valve stem 55 diameter of 19 mm, the outer face area A₁ and inner face area (taken to be equal to A₂) will be 1.96E-3 m² and 1.68E-3 m² respectively. For a pressure value P of 1.764E5 Pa, and for a spring extension x of 10 mm, the spring constant k can be calculated as follows:

1.764^5(1.96^ - 3 - 1.68£ - 3)

Ϊ0 k < 4.94 N/mm

A suitable value of spring constant k for this example would therefore be 4 N/mm. Also for this example, a suitable axial gap between the wall 45 of the bearing housing 3 and the radial wall 6a of the compressor housing 6 would be 10 mm or more. This is to allow axial clearance for the moveable element 53 to extend by 10 mm. The closing force required to close the moveable element would be 40 N in this instance.

A radially outermost point of the moveable element 53 may be aligned with the radially outermost point of the radial wall 6a of the compressor housing 6. In other embodiments, the radially outermost point of the moveable element 53 may be radially inwards of the radially outermost point of the radial wall 6a of the compressor housing 6.

In order to define a range of suitable geometries of connecting channel 47 for different turbochargers, it is possible to define a range of ratios of a circumferential flow path area to connecting channel 47 cross-sectional area.

The circumferential flow path area of the compressor 2 is defined in part by the circumference of the annular flow passage 37 just radially inward of the moveable element 53. When the circumference is swept axially such that the circumferential area spans the axial gap across the annular flow passage 37, the circumferential flow path area is defined. The circumferential flow path area is the area through which air flows as it passes through the annular flow passage 37, at the point just radially inward of the moveable element 53.

In a first non-limiting example, an outlet of the annular flow passage 37 has a radius of around 47 mm. With a moveable element outer face 49 diameter of 25 mm taken away from the first radius, the radius just radially inward of the moveable element 53 is around 22 mm. When swept across the annular flow passage 37 axial gap of 5 mm, the circumferential flow path area is around 690 mm². For a corresponding connecting channel 47 cross-sectional area of 3.15 mm², the ratio of the circumferential flow path area to the connecting channel 47 cross-sectional area is around 220:1.

In a second non-limiting example for a larger turbocharger, the outlet of the annular flow passage 37 has a radius of around 174 mm. With a moveable element outer face 49 diameter of 50 mm taken away from the first radius, the radius just radially inward of the moveable element 53 is around 124 mm. When swept across the annular flow passage 37 axial gap of 10 mm, the circumferential flow path area is around 7790 mm². For a corresponding connecting channel 47 cross-sectional area of around 23 mm², the ratio of the circumferential flow path area to the connecting channel 47 crosssectional area is around 340:1.

Figures 6 and 7 are graphs which show data from computational fluid dynamics simulations for a prior art compressor which does not incorporate the present invention. Figure 6 is a plot of data when the compressor is operating at 70,000 RPM (i.e. 70k

RPM) and in Figure 7 the compressor is operating at 112k RPM. The graphs are included merely to further describe the presence of a positive pressure gradient between the annular flow passage 37 and the volume 29, which can be utilised by the present invention. To aid in the explanation, reference numerals from Figure 2, which does incorporate the present invention, are used.

Plotted on the y axis for both plots is the difference in static pressure, measured in pascal (Pa), between an outlet of the annular flow passage 37 and a surface of the oil seal ring 18a adjacent the volume 29. The difference in static pressure is equal to the static pressure at the oil seal ring 18a surface subtracted from the static pressure at the outlet of the annular flow passage 37. The outlet of the annular flow passage 37 is defined as the radially outermost point of the annular flow passage 37 where the annular flow passage 37 meets the annular outlet volute 14. In other words, the outlet of the annular flow passage 37 is the radial position within the annular flow passage 37 which is adjacent the radially outermost point of the radial wall 6a of the compressor housing 6.

Plotted on the x axis for both plots are increasing values of mass flow parameter (MFP). The MFP is a common variable used in compressor maps which takes into account mass flowrate, temperature and pressure of a fluid.

Each of Figures 6 and 7 have three separate plots, 48a-c and 52a-c respectively. Each of the plots (a-c) represents a different diffuser radius ratio (DRR) value. The DRR values used are 1.35 (48a and 52a), 1.59 (48b and 52b) and 1.83 (48c and 52c). The DRR values are the ratios of the radial positions of the annular stepped outer portion 35 of the flow channel 36 to the outlet of the annular flow passage 37. That is to say, for a DRR of 1.35 whereby the annular stepped outer portion 35 is 50 mm radially outward from the turbocharger axis 33, the outlet of the annular flow passage 37 is 67.5 mm radially outward from the turbocharger axis 33. As such, the different plots represent different compressor geometries.

Figures 6 and 7 demonstrate that there exists a positive pressure difference or gradient between the outlet of the annular flow passage 37 and the oil ring seal 18a surface. Due to the location of the oil ring seal 18a surface, the pressure at the oil ring seal 18a surface is representative of the pressure inside the volume 29. Figures 6 and 7 therefore illustrate how the present invention can be implemented to provide a positive PDOS where required by utilising the available pressure gradient. Figures 6 and 7 also demonstrate how the invention can operate over the compressor map (i.e. in a range of different compressor operating conditions).

Using the results of the simulations it is possible to estimate the maximum diameter of the connecting channel 47 required to keep mass flow leakage through the connecting channel 47 to a given percentage of the overall mass flow through the compressor when the compressor operates at a given MFP value. Variables in the below “choked mass flow” equation are as follows: m is equal to the mass flow rate of air through the connecting channel 47, A, is the maximum cross-sectional area of the connecting channel 47, γ is the specific heat ratio of the fluid, Ft is the gas constant, p₀ and T_o are the pressure and temperature at the outlet of the annular flow passage 37 and M is the mach number of the fluid flow. In the worst case of choke flow, M is equal to 1. That is to say, the flow velocity is equal to the speed of sound, m, the mass flow rate of air through the connecting channel 47, is calculated as a percentage of the mass flow which “leaks” from the main compressor flow path through the connecting channel 47. The choke flow equation is provided below. Connecting channel cross-sectional area 4(can be calculated by rearranging the equation:

Po * aK /+1

2(/-1)

Inputting the variables from the table below into the above equation (and selecting an MFP value, 25, and DRR value, 1.35), it is possible to calculate a value of A,. A corresponding diameter, D, of the connecting channel 47 can be calculated from the cross-sectional area A_t. The calculation shows that a diameter of 2 mm for the connecting channel 47 would provide 1% leakage in a worst case scenario. By 1% leakage it is meant that 1% of the overall mass flow of air through the compressor 2 flows through the connecting channel 47. By providing 1% leakage in the worst case scenario, the performance of the compressor 2 and so turbocharger overall is maintained. That is to say, performance is not noticeably impacted by implementing the invention. This is the case for the pressure at the outlet of the annular flow passage 37 being 1.288 bar or lower. For physically larger turbocharger and so compressor sizes, the diameter could be larger whilst maintaining a maximum of 1% leakage as the MFP at which the larger turbocharger would operate would be higher. As a result, drilling or adding other features would become more feasible. The numerical values in this example are not intended to be limiting and are provided merely as a worked example.

Variable	Value	Unit
MFP	25	kg/s*sqrt(K)/Mpa
Leakage	1%
stacie	0.14386	kg/s
m	0.00144	kg/s
M	1
Y	1.4
R	287	J*Kg/K
Po	1.288	bar
To	324.7	K
A_t	3.15	2 ΙΊΊΙΊΊ
Diameter (D)	2	mm

As well as the above example, a further example is provided in the table below. The second example relates to a larger frame turbocharger and so compressor. As a result, the MFP value of the following example is higher and consequently the maximum cross-sectional area of the connecting channel 47 is also larger, whilst still maintaining a maximum 1% leakage. Once again, the numerical values are not intended to be limiting and are provided merely as a worked example.

Variable	Value	Unit
MFP	250	kg/s*sqrt(K)/Mpa
Leakage	1%
Wlstaqe	1.43859	kg/s
m	0.01439	kg/s
M	1
Y	1.4
R	287	J*Kg/K
Po	1.764	bar
To	324.7	K
A,	23	2 ΙΊΊΙΊΊ
Diameter (D)	5	mm

Figure 8 is a graph which shows conditions within a prior art compressor. Although the graph relates to a prior art compressor, macroscopic values in Figure 8 do not change significantly when the invention is implemented as the invention has only a microscopic local impact upon the flow. As such, Figure 8 is generally representative of the compressor 2 which incorporates the invention. Reference numerals are used accordingly to explain the positions within the compressor 2 where the pressure values are taken. The graph shown in Figure 8 is plotted using simulation data. Entropy of the air flow through the compressor 2 is shown on the x axis and increases from a zero value at the y axis intercept. Temperature of the air flow through the compressor 2 is plotted on the y axis. The labelled points 60, 62, 64, 66 demonstrate how the conditions vary as the air flows through the compressor 2. The first point 60 represents conditions at the impeller inlet i.e. between the axial inlet 13 and the blades 7a of the compressor impeller 7 (see Figure 1). The second point 62 represents conditions at the impeller outlet i.e. just downstream of the compressor impeller 7. The third point 64 is taken at a diffuser outlet i.e. the outlet of the annular flow passage 37. The final point 66 is taken at a stage outlet i.e. at an exit point of the compressor 2 downstream of the annular outlet volute 14 (the exit point is not visible in Figures 1 or 2).

The graph shows how the temperature of the air increases as it passes through the compressor 2. The graph also shows the gain in static pressure at subsequent points 62, 64, 66 relative to the impeller inlet condition of the first point 60. The gain in total pressure is also shown at subsequent points 62, 64, 66 relative to the impeller condition of the first point 60. Figure 8 also demonstrates the presence of a static pressure gradient in the annular flow passage 37 which can be utilised by the invention.

Figure 9 is a generic compressor map for a prior art turbocharger which displays performance characteristics of a compressor when operating in different conditions. Flow parameter (the mass flow parameter as mentioned above) is plotted on the x axis and pressure ratio is plotted on the y axis. The thicker lines within the compressor heart are compressor speed lines. The thinner lines within the compressor heart are efficiency lines. Regions where a compressor is likely to suffer from surge or choke are indicated. Similarly, a region of excess rotor speed is also indicated. The highlighted region is a target region for the invention. When the compressor operates in the highlighted region, oil leakage and thrust bearing failure could be a problem if not for incorporation of the invention. This is because the pressure in a volume behind a compressor impeller is too low. That is to say, by incorporating the invention, a more positive PDOS can be maintained in the compressor when operating in the highlighted region. A more positive PDOS can be maintained by selectively introducing pressurised air into a volume behind a compressor impeller.

A further advantage of the present invention is that it can be used to assist in the balancing of thrust loads on the shaft 8. Thrust loads are forces which act in a direction parallel to the length of a shaft. Thrust loads are undesirable due to the risk of damage to roller bearings and other components, to name just some examples.

As explained above and with reference to Figure 1, the turbine impeller 5 and compressor impeller 7 are coaxially mounted on the shaft 8. During normal turbocharger operation, thrust loads will typically urge the shaft 8 towards the compressor 2 side of the turbocharger. This is due to the loading on the turbine impeller 5 as exhaust gases are expanded through the turbine impeller 5. Furthermore, the highest thrust loads experienced during turbocharger operation occur when the compressor 2 is operating in the surge condition, whereby the pressure ratio across the compressor 2 is high. As such, the thrust bearing assembly 16, and so thrust bearing, is incorporated in the turbocharger to oppose the thrust loads acting towards the compressor 2 side.

When the compressor 2 flow is choked (i.e. the compressor 2 is operating in the choked condition), thrust loads urge the shaft 8 towards the turbine 1 side of the turbocharger. This is due to a comparatively low pressure ratio across the compressor 2 in combination with relatively high flow rates of air through the compressor 2. For the reasons mentioned above, turbochargers are designed to effectively counteract thrust loads in the compressor 2 direction. However, thrust loads in the turbine 1 direction risk damage to the turbocharger by way of, for example, thrust bearing failure.

The invention provides a way of selectively pressurising the volume 29 behind the compressor impeller 7. During operation, pressurised air in the volume 29 will exert a force over the area of the back face of the compressor impeller 7. Such a force will have the effect of urging the shaft 8 towards the compressor 2 side of the turbocharger. Therefore, if the volume 29 is pressurised when the shaft 8 is experiencing thrust loading towards the turbine 1 side, the force exerted on the back face of the compressor impeller 7 will help to counteract the thrust loading. The invention therefore assists in balancing axial loading upon the shaft 8. This is achieved by counteracting thrust loading of the shaft towards the turbine side of the turbocharger.

As the position of the moveable element 53 is dependent at least in part upon the operating condition of the compressor 2 (for instance, choke), the moveable element 53 could also be used to affect the flow in only these certain operating conditions. Similarly, the valve seat 39 may be used to only influence the flow in certain operating conditions.

The housing may comprise a compressor housing 6, bearing housing 3 and a turbine housing 4.

The connecting channel 47 may open into the volume 29 behind the compressor impeller 7 at a range of positions. For instance, the connecting channel 47 may open into the volume 29 in a radial direction and at any angular position. Alternatively, the connecting channel 47 may open into the volume at an angle which is offset from a radial direction.

The connecting channel 47 may be a bore machined through the bearing housing 3. Alternatively, the connecting channel 47 may be a bore machined through the compressor housing 6. The connecting channel 47 may comprise a number of individual bores. Further alternatively, the connecting channel 47 may be a separate component. For example, the connecting channel 47 may be a pipe. The connecting channel 47 may have a continuous cross-section. Alternatively, the connecting channel 47 may vary in cross-section.

There may be a seal between the compressor impeller 7 and the bearing housing 3. For example, a labyrinth seal may be used. Alternatively, another variety of seal may be used. Sealing arrangements may substantially isolate the volume 29 from the annular flow passage 37 and/or other locations in the compressor 2 downstream of the compressor impeller 7 when the moveable element 53 is in the closed configuration. Isolating the volume 29 may help to maintain an increased pressure in the volume 29. The increased pressure may be maintained by restricting the leakage of the pressurised air from within the volume 29 out through the compressor 2. Maintaining an increased pressure in the volume 29 may contribute to a more positive PDOS which is desirable for sealing purposes. This may, in turn, reduce oil leakage from the bearing housing 3 into the compressor 2.

Alternatively, instead of using an O-ring to substantially seal the moveable element 53 5 against the valve seat 39, other seals may be used such as face seals, lip seals or gaskets.

The described and illustrated embodiment is to be considered as illustrative and not restrictive in character, it being understood that only a preferred embodiment has been shown and described and that all changes and modifications that come within the scope of the inventions as defined in the claims are desired to be protected. In relation to the claims, it is intended that when words such as a, an, at least one, or at least one portion are used to preface a feature there is no intention to limit the claim to only one such feature unless specifically stated to the contrary in the claim. When the language at least a portion and/or a portion is used the item can include a portion and/or the entire item unless specifically stated to the contrary.

Optional and/or preferred features as set out herein may be used either individually or in combination with each other where appropriate and particularly in the combinations as set out in the accompanying claims. The optional and/or preferred features for each aspect of the invention set out herein are also applicable to any other aspects of the invention, where appropriate.

Claims

CLAIMS:

1. A turbocharger comprising: a turbine;

a housing;

2. A turbocharger according to claim 1, wherein the outer face of the moveable element aligns flush with a wall of the housing when the moveable element is in the closed configuration.

3. A turbocharger according to any preceding claim, wherein the outer face of the moveable element has a generally circular cross-section.

4. A turbocharger according to any preceding claim, wherein the moveable element is a mechanical plunger.

5. A turbocharger according to any preceding claim, further comprising an O-ring arranged to seal the moveable element against the housing when the moveable element is in the closed configuration.

6. A turbocharger according to any preceding claim, wherein the ratio of the crosssectional area of the outer face of the moveable element to the cross-sectional area of an inner face of the moveable element is between 1.1:1 and 1.6:1.

7. A turbocharger according to any preceding claim, wherein the biasing member is a coil spring.

8. A turbocharger according to claim 7, wherein the coil spring has a spring constant of less than around 6 N/mm.

9. A turbocharger according to claims 7 or 8, wherein the coil spring has a spring constant of at least around 3 N/mm.

10. A turbocharger according to any preceding claim, wherein the connecting channel includes a change of direction.

11. A turbocharger according to any preceding claim, wherein a valve seat of the connecting channel includes a stepped profile.

12. A turbocharger according to any preceding claim, wherein a smallest crosssectional area of the connecting channel is at least around 2 mm2.

13. A turbocharger according to claims 1 to 11, wherein a smallest cross-sectional area of the connecting channel is less than around 20 mm2.

14. A turbocharger according to any of claims 1 to 11, wherein the ratio of a circumferential flow path area of the compressor to a smallest cross-sectional area of the connecting channel is between 150:1 and 410:1.

15. A turbocharger according to claim 14, wherein the ratio of the circumferential flow path area of the compressor to the smallest cross-sectional area of the connecting channel is between 200:1 and 360:1.