GB2575478A

GB2575478A - A centrifugal compressor

Info

Publication number: GB2575478A
Application number: GB1811348.0A
Authority: GB
Inventors: Thomas Tarver Benjamin; Harley Peter
Original assignee: Dyson Technology Ltd
Current assignee: Dyson Technology Ltd
Priority date: 2018-07-11
Filing date: 2018-07-11
Publication date: 2020-01-15
Also published as: GB201811348D0; WO2020012186A1

Abstract

A centrifugal compressor with an impeller (201, fig. 2b), a diffuser passage 401 and a volute 213, wherein the ratio of the impeller’s outer diameter to the diffuser’s impeller’s outer diameter ranges from 0.85 to 0.95. Such ratios aim to provide a dimensionally more compact compressor with a relatively shorter diffuser passage, striking a balance between packaging constraints and degree of pressure recovery by the diffuser. Preferably the ratio is more than 0.88 and less than 0.92, more preferably ranging from 0.89 to 0.91. Preferably the impeller’s outer diameter ranges from 17 to 19 cm, more preferably from 17.9 to 18.3 cm. A vehicle with a ventilation system comprising such a compressor and suitable air ducts, drawing air from outside the passenger cabin to the compressor and then discharging it to the passenger cabin, is also claimed.

Description

A CENTRIFUGAL COMPRESSOR

Field of the Invention

The present invention relates to a centrifugal compressor, and to a vehicle including the centrifugal compressor.

Background of the Invention

Centrifugal compressors typically comprise an impeller, a diffuser passage, and a volute.

It is known that the performance of a centrifugal compressor is at least partly determined by the absolute and relative dimensions of the stages of the compressor. For example, an impeller with a greater diameter will typically be capable of pumping a greater mass of air per rotation than a lesser diameter impeller, a diffuser stage with a greater radial length will result in increased fluid diffusion and so improved recovery of static pressure, and the internal volume and geometry of the volute will influence compressor efficiency and operating range.

Consequently, a high-volume, high-pressure and/or high-efficiency centrifugal compressor may disadvantageously tend to have relatively great outer dimensions.

Summary of the Invention

According to a first aspect of the present invention there is provided a centrifugal compressor comprising: a centrifugal impeller; a volute for collecting fluid expelled by the impeller; and a diffuser passage extending between the impeller and the volute. The ratio of the impeller outer diameter to the diffuser passage outer diameter is in the range of 0.85 to 0.95.

Rotation of the impeller draws in fluid in an axial direction, and the turn of the impeller blades accelerates the flow radially towards a peripheral outlet. The high velocity flow is admitted to the diffuser passage where the fluid is diffused resulting in recovery of static pressure. From the diffuser passage the diffused fluid is discharged to the volute which collects the fluid and directs the flow to an outlet port for supply to a downstream stage.

It is understood that the absolute and relative dimensions of the impeller and diffuser passage affect the performance characteristics of the impeller. For example, a larger diameter impeller will typically provide more efficient pumping than a smaller diameter impeller, and a longer diffuser passage will typically result in increased fluid diffusion and so greater pressure recovery. Thus, compressor performance may typically be improved where the absolute dimensions of the impeller and diffuser passage are increased.

However, practical considerations, for example, packaging constraints, will typically limit the exterior diameter dimension of the compressor, and similarly limit the diameter of the impeller and the diffuser passage. Thus, a balance must be found between the diameter of the impeller and of the diffuser passage.

An impeller outer diameter to diffuser passage outer diameter ratio in the range of 0.850.95 has been found to advantageously balance impeller efficiency and flow rate with diffusion in the diffuser passage.

The ratio of the impeller outer diameter to the diffuser passage outer diameter may be at least 0.88, and may be less than 0.92. Further, the ratio of the impeller outer diameter to the diffuser passage outer diameter may be in the range of 0.89 to 0.91. Ratios in these ranges have been found to more advantageously balance impeller efficiency and mass flow rate with diffusion in the diffuser passage. Ratios in the range of 0.89 to 0.91 have been found to most advantageously balance these characteristics for at least one operating point.

The outer diameter of the impeller may be in the range of 17 centimetres to 19 centimetres, and optionally in the narrower range of 17.9 centimetres to 18.3 centimetres. Impeller outer diameters in the specified range have been found to exhibit a good balance between size and performance. Specifically, impellers with the specified diameters have been found to advantageously achieve a good mass flow rate, good pressure rise, and good efficiency.

According to a second aspect of the present invention there is provided a vehicle comprising a passenger cabin for accommodating passengers and a passenger cabin ventilation system including the centrifugal compressor according to any one of the preceding statements.

A centrifugal compressor having the aforementioned characteristics has been found to perform well as a cabin air ‘blower’ in a vehicle ventilation (HVAC) system. In particular, it has been found that the described centrifugal compressor is advantageously capable of providing relatively high volume and high pressure flows of air, has a relatively high operating range, and at certain typical flow rates operates more efficiently than conventional cabin ‘blowers’, for example, ‘squirrel-cage’ blowers. In particular, because a centrifugal compressor according to the first aspect of the present invention may be more dimensionally compact than conventional centrifugal compressors, and because in a preferred embodiment the compressor may be less susceptible to mechanical shock, the centrifugal compressor of the present invention may be better suited to incorporation in a vehicle HVAC system than conventional centrifugal compressors.

The vehicle may comprise ducts for ducting air from outside the passenger cabin to an inlet of the compressor, and for ducting air from an outlet of the compressor to the passenger cabin. Advantageously this allows air in the passenger cabin to be replaced with atmospheric air, thus improving air quality in the cabin.

Brief Description of the Drawings

In order that the present invention may be more readily understood, embodiments of the invention will now be described, by way of example, with reference to the accompanying drawings, in which:

Figure 1 is a schematic side view of a vehicle in the form of a passenger car according to an exemplary embodiment of one aspect of the present invention;

Figures 2a and 2b are perspective and exploded views respectively of the example centrifugal compressor previously identified in Figure 1;

Figure 3a shows the impeller of the centrifugal compressor in a perspective view, and Figure 3b shows the impeller in a cutaway perspective view. Figures 3c and 3d show top plan and side elevation views respectively of the impeller, and Figure 3e shows a side cross-sectional view of the impeller;

Figures 4a and 4b are first and second mutually orthogonal side cross-sectional views of the centrifugal compressor; and

Figures 5a and 5b are cross-sectional views of the centrifugal compressor along the plane depicted by the line C—C identified in Figure 4b.

Detailed Description of the Invention

Referring firstly to Figure 1, passenger car 101 comprises a body structure 102 defining internally a passenger compartment 103 for accommodating passengers, and a ventilation system, indicated generally at 104, for ventilating the passenger compartment.

Ventilation system 104 comprises an air-handling unit 105 installed in the passenger compartment, and ducts for ducting air from the vehicle exterior to the passenger cabin space. In the example air handling unit 105 comprises a blower in the form of centrifugal compressor 106, and an air filter 107. A first duct 108 is provided with a first end open to an exterior of the vehicle and a second end in communication with an inlet of the compressor 106 for ducting air from an exterior of the vehicle to the compressor. A second duct 109 is provided with a first end in in communication with an outlet of the compressor via the air filter 107, and a second end open to the passenger cabin space, such that air discharged by the compressor 106 passes through the air filter 107 and is admitted to the passenger cabin space.

Centrifugal compressor 106 may thus be operated as a ‘blower’ to draw in fresh-air from the exterior of the vehicle, which may then be filtered by the air filter 107 to remove particulate and/or gaseous contaminants, before being admitted to the passenger cabin space to ventilate the cabin.

Referring secondly to Figures 2a and 2b, in the example centrifugal compressor 106 comprises an impeller 201, a motor unit 202, and a two-part housing comprising upper and lower parts 203, 204. In the assembled, operational, condition shown in Figure 2a the impeller 201 is rotatably mounted within the housing 203, 204 for relative rotation thereto and may be driven to rotate by the motor 202.

Impeller 201 is a closed centrifugal impeller comprising a plurality of three-dimensional blades, indicated generally at 206, an upper shroud 207, and a lower shroud 208. An axial inlet opening 209 and a peripheral outlet opening 210 are defined by the upper and lower shrouds 207, 208. The construction of impeller 201 will be described in further detail with reference to Figures 3a to 3e.

The upper part 203 of the housing defines a generally frustoconical wall 211 which in the assembled condition covers and circumferentially surrounds the impeller upper shroud 207. The central section 211 defines an intake opening 212 coaxial with the impeller 201 for intake of air to the impeller inlet 209. The upper part 203 further defines a generally toroidal volute 213 arranged to extend circumferentially about the impeller 201, having an inlet opening 214 in communication with the peripheral outlet opening 210 of the impeller 201, and having an outlet opening 215 through which fluid may be discharged from the compressor. In the example the upper and lower housing parts 203, 204, respectively are formed of a hard plastics material, such as nylon or acrylonitrile butadiene styrene (ABS).

The upper and lower parts 203, 204, of the housing together further define a generally annular diffuser passage (not visible in Figure 2a or 2b) extending radially between the impeller and the volute for communicating the outlet 210 of the impeller with the inlet 214 of the volute. The diffuser passage will be described in further detail with reference to Figures 4a and 4b.

Motor unit 202 comprises principally an electric motor 216 and a sole plate 217. Electric motor 216 is conventional in construction, and has a driven shaft 218 on which the impeller 201 is mounted defining an axis of rotation of the impeller. In the example, the motor 216 is configured to be capable of rotating the impeller 201 at a range of different rotation rates between approximately 3000 revolutions per minute (rpm and approximately 5000 rpm. In the example the electric motor 216 is a direct current (DC) electric motor, but could alternatively be an alternating current (AC) electric motor, or indeed an alternative form of motor operable to rotate the impeller 201.

Electric motor 216 is hard-mounted to the sole plate 217, which in the example is formed of a similar plastics material to the plastic of the housing 203, 204. Sole plate 217 is soft-mounted to the lower part 204 of the housing. The soft-mounting of the motor unit to the housing 204 serves to reduce transmission of vibration from the motor 216 to the housing 204, and so reduce onward transmission of vibration to the body structure of the vehicle.

Referring next to Figures 3 a to 3 e, as previously described, in the example impeller 201 comprises axially spaced generally frustoconical upper and lower shrouds 207, 208 respectively, and a plurality of blades 206. An axial inlet opening 209 exposing the leading edges of the blades and a peripheral outlet opening 210 exposing the trailing edges of the blades are defined between the lower and upper shrouds.

The lower shroud 208 has an aerodynamic upper surface 301. The upper shroud 207 has an aerodynamic lower surface 302. The plurality of blades 206 extend between the surfaces 301, 302, from respective root ends 303 joined to the surface 301 to respective tip ends 304 joined to the surface 302. Thus, in use, fluid, in the example air, may flow through the generally frustoconical passage defined between the surface 301 and the surface 302 acted on by the blades 206.

In the example a lower surface of the lower shroud 208, forming an underside surface of the impeller, is concave axially upwards of the impeller such that a recess 313 is defined to the underside of the impeller. As shown in later Figures, in the assembled condition the motor 216 is partly received in the concave underside profile of the lower shroud 208. As a result the overall axial height of the impeller-motor pair is reduced.

The lower shroud 207 comprises a hub portion 305 defining centrally an axially extending bore 306 which receives the shaft 218 of the motor 216 on which the impeller rotates. The plurality of blades are circumferentially spaced about the hub and extend radially outwards to respective tip ends. Except for the bore 306, the area of the hub is substantially closed.

The area of the inlet opening 209 is thus generally annular and equal to the area of the opening defined by the upper shroud 207, minus the area of the closed hub 305 at the inlet opening defined by the lower shroud 208. In the example, the lower shroud 208 has an outer diameter DI at the inlet opening of approximately 3 centimetres (cm), and the upper shroud has an inner diameter D2 at the inlet opening of approximately 11 cm. The area of the inlet opening 209 of the impeller 201 is thus approximately 88 square centimetres (cm ).

The area of the outlet opening 210 is a function of the circumference, and so the outer diameter D3, of the impeller and the axial spacing between the upper and lower shrouds 207, 208, at the outer diameter of the impeller, i.e. the distance H between the upper surface 311 of the lower shroud 208 and the lower surface 312 of the upper shroud 207.

In the example, the outlet opening 210 has a height dimension Hl of approximately 1.9 cm, and the impeller has an outer diameter dimension of approximately 18.1 cm. The area of the outlet opening 210 of the impeller 201 is thus approximately 108 cm .

Thus, in the example, the ratio of the inlet opening area to the outlet opening area is approximately 0.81.

Referring particularly to Figure 3b, in the example one group of the plurality of blades are ‘full’ blades, such as blade 307, and another group are ‘splitter’ blades such as blade 308. In the example the impeller comprises nine full blades 307 and nine splitter blades 308. The full blades each extend substantially the full axial height of the impeller from the upper surface 301 of the lower shroud 208 to the inlet opening 209. The splitter blades extend approximately half the axial height from the surface 301 towards the inlet opening 209. The principal reason for the splitter blades being cut short is to reduce the blockage to the impeller inlet opening 209 presented by the leading edges of the blades, to thereby reduce the likelihood of inlet stall occurring under high flow conditions. The plurality of blades are arranged alternately such that a splitter blade extends mid-pitch between adjacent full blades.

The plurality of blades have a relatively complex three-dimensional profile. For example, each of the full blades comprises generally three distinct stages:

A leading edge 309 of each of the blades 307 is angled forwardly in the direction of rotation of the impeller, so as to function to scoop fluid into the blade channel and move the fluid axially. In the example, the pitch angle of the leading edge of each of the full blades develops from the root end 303 to the tip end 304, such that the blade undergoes a twist from a lower pitch at the root end to a greater pitch at the tip end. In the example, each of the full blades 307 is formed such that the pitch of the leading edge at the root end is approximately 20 degrees and develops towards the tip end which has a pitch angle of approximately 60 degrees, each taken relative to a plane aligned with the rotational axis of the impeller.

A central portion 310 of each blade is shaped to gradually turn the direction of fluid flow as the impeller rotates from the axial direction to the radial direction, and to radially accelerate the fluid flow to impart a high radial velocity component to the flow. As will be described with reference to later Figures, it is typically desirable that the fluid flowing through the blade channels is turned relatively gently from the axial to radial direction of flow so as to avoid flow separation for the blade surfaces. Central portion 308 of the blades is thus preferably shaped to turn the fluid flow gently.

A final, radially outermost, portion 311 of each blade is adapted to further accelerate the velocity of the flow radially and discharge the flow cleanly from the blade trailing edge 312 to the diffuser passage. The trailing edge of the blade is similarly complex, provided with a pitch relative to the imaginary plane aligned with the rotational axis, and a back-taper or back-sweep angle relative to the radial direction. In the example, the trailing edge has a pitch angle of approximately 30 degrees.

The splitter blades are similar in profile to the full blades, except that the axial height of the splitter blades is reduced such that the leading edge of each splitter blade is downstream compared to the leading edges of the full blades. Further, given that the splitter blades are not required to induce airflow into the impeller, the leading edge of each splitter blade is not pitched so greatly as the leading edges of the full blades.

In the example, the upper and lower shrouds 207, 208 are formed integrally with the blades 206, such that the impeller has a unitary structure. As will be described with reference to later Figures, integral shrouds are advantageous in applications where maintaining a close spacing between a static shroud and the blade tips of a relatively movable impeller would be difficult, for example, where the impeller is expected to be subjected to mechanical shock during operation.

Difficulties can however be encountered in manufacturing an impeller of the type with integral upper and lower shrouds, inasmuch that it can be difficult to mould or machine the channels between the blades and the shrouds. To overcome these difficulties, in the example the impeller 201 is manufactured using a known two-stage process, in which the lower shroud 208, and blades 206, are firstly formed using an injection moulding technique, and secondly the preformed upper shroud 207 is joined, for example by ultrasonic welding, to the tip ends of the plurality of blades. In the specific example the impeller 201 is formed of a hard plastics materials such as nylon or acrylonitrile butadiene styrene (ABS). It should be understood however that the choice of material and manufacturing technique is a selection dictated by the expected operational demands placed on the impeller, for example, operating temperatures and rotational speeds. Thus, in dependence on the intended application, the impeller could alternatively be formed of a different material or using an alternative manufacturing technique, for example, milling or casting techniques on a metal such as steel or titanium

Referring next to Figures 4a and 4b, the compressor is shown in the assembled condition previously depicted in Figure 2a, in which the impeller 201 is located in the housing 203, 204.

In the example the impeller is mounted to the shaft 218 of the motor by an interference fit between the bore 306 of the impeller and the shaft which rotationally couples the impeller to the shaft such that the impeller may be driven to rotate by the motor.

The impeller 201 is covered and circumferentially surrounded by the wall 211 of the housing part 203, such that the intake opening 212 of the housing is approximately aligned with the impeller inlet opening 209.

As will be understood, in operation, fluid, in the example air, is drawn through the intake opening 212 in an axial direction, and enters the impeller inlet opening 209. The turn of the impeller blades accelerates the flow radially towards the peripheral outlet 210, imparting a high kinetic energy to the air. At the outlet of the impeller the flow has both high radial and tangential velocity components. From the outlet 210 the high velocity fluid flow is admitted to the diffuser passage.

Diffuser passage 401 is arranged circumferentially about the impeller and communicates the impeller 201 with the volute 213. The diffuser passage 401 is bounded axially by the upper and lower walls 402, 403 of the housing 203, 204.

The diffuser passage 303 is open at a first end to the outlet 210 of the impeller, and extends radially therefrom to a second end open to the volute inlet 214. The diffuser passage 401 thus fluidly connects the impeller outlet 210 with the volute inlet 214 such that fluid expelled by the impeller passes through the diffuser passage before entering the volute.

In the example, the diffuser passage 401 has a ‘vaneless’ construction. Vaneless diffusers are well known in the prior art, and are contrasted to ‘vaned’ diffusers. The example vaneless diffuser passage has the advantage of relatively greater reduction in the radial velocity component of the airflow than could typically be achieved using a vaned diffuser.

The axial height H2 of the diffuser passage, that is the axial distance between the upper and lower walls 402, 403, is approximately equal to the axial height Hl of the peripheral outlet opening 210 of the impeller 201, i.e. the axial distance between the lower surface 302 of the upper shroud 207 and the upper surface 301 of the lower shroud 208. Matching the height of the diffuser passage to the height of the impeller outlet opening advantageously minimises turbulence and other undesirable aerodynamic effects that are inevitably induced as the airflow transitions from the impeller to the diffuser passage.

As will be understood, the primary function of the diffuser is to reduce the velocity of the fluid leaving the impeller to achieve an increase in static pressure of the fluid. Diffusion of the fluid occurs along the diffuser passage principally because the area of the passage increases as a function of its radius. The increasing area of the diffuser reduces the velocity of the fluid, and correspondingly increases the static pressure of the flow.

From the diffuser passage the fluid is discharged into the volute 213. As will be understood, the primary function of the volute is to collect the fluid from the diffuser and direct it to a single outlet. Depending on the geometry of the volute however further diffusion, principally of the tangential components of the velocity, may occur in the volute.

Referring to the Figures, in the example volute 213 is generally toroidal and arranged circumferentially about the impeller 201 and the diffuser passage 401.

The inlet opening 214 of the volute is open to the outlet of the diffuser passage 401 around substantially the entire circumference of the diffuser passage, such that fluid from the impeller may be admitted to the volute about the full circumference of the diffuser passage.

Volute 213 is generally ‘D’-shaped in cross-section, having a flattened outer side wall 404 and a curved inner side wall 405 that is convex from the centroid axis of the volute towards the rotational axis of the impeller.

The axial and radial dimensions, and correspondingly the cross-sectional area of the volute, increase circumferentially towards the throat end of the volute. The increasing cross-sectional area of the volute accommodates the circumferentially increasing volume of fluid admitted to the volute.

An inlet portion 405 of the volute, open to the volute inlet 214, extends circumferentially about the diffuser passage 401 bounded by a lower wall 406 which joins continuously with the lower wall 403 of the diffuser passage 401.

From the inlet portion 405 the volute extends axially upwardly in a direction generally parallel to the rotational axis of the impeller, and an upper portion 407 of the volute extends radially inwardly over the diffuser passage 401 and the impeller 201 along at least part of the circumference of the volute.

The radial distance by which the volute 213 extends over the impeller 201 increases circumferentially of the volute. Referring to the Figures, at an end 408 where the crosssectional area of the volute is lowest, the inner diameter of the volute defined by the inner side wall 405 is greater than the outer diameter of the diffuser passage 401 for the full axial height of the volute, such that the entire area of the volute has a diameter greater than the outer diameter of the diffuser passage.

In comparison, approaching the throat end 409, the inner side wall 405 of the volute is curved radially inwardly towards the rotational axis of the impeller to an extent that an upper portion 407 of the volute extends radially over the full radial length of the diffuser passage 401 and over approximately the outer 40 percent of the diameter of the impeller 201.

As may be recognised, that the volute, unconventionally, extends radially inwardly over the impeller, allows the internal volume of the volute to be increased for any given volute outer diameter as compared to a conventional volute which does not extend radially over the impeller. This configuration thus allows for the outer diameter of the volute to be reduced as the corresponding decrease in volute volume may be accommodated by the decrease in volute inner diameter, i.e. an increase in the extent to which the volute extends radially inwards. In this configuration the volute may thus occupy the area axially above the impeller which, for conventional compressor designs, may be redundant space.

As previously described, the upper part 203 of the housing comprises a generally tubular wall 211 which covers and circumferentially surrounds the upper shroud 207 of the impeller 201. At an upper end the wall defines the intake opening 212 coaxial with the impeller 201 for directing fluid to the impeller inlet 209. From the intake opening 212 the wall extends axially downwardly to the volute inlet 214 where the wall 211 joins the inner side wall 307 of the volute 213.

By extending from the volute inlet 214 to the intake opening 212, the wall 211 acts to close the radial gap between the periphery of the impeller 201 and the diffuser passage inlet, such that the requisite gap between the stationary housing and the rotatable impeller is moved to adjacent the inlet opening 209 of the impeller. Consequently fluid within the diffuser passage 401 and the volute 213 is better contained within the housing and can less easily escape through the gap between the volute wall and the periphery of the relatively movable impeller shroud.

In the example, the wall 211 is concave towards the rotational axis of the impeller such that an inner surface of the side wall conforms closely to the outer profile of the impeller upper shroud 207 with a uniformly narrow gap 409 defined between the two surfaces extending the length of the impeller upper shroud 207. The close conformation between wall 211 and the impeller upper shroud 207 further inhibits escape of fluid in the volute or diffuser passage.

Referring to Figures 5a and 5b, the relative dimensions of the stages of the compressor will be described.

As described, the performance of each stage of the compressor and of the compressor overall is influenced by the absolute and relative dimensions of the impeller 201, the diffuser passage 401, and the volute 213.

Referring in particular to Figure 5b, it will be noted that the impeller 201 of the example compressor has a relatively great outer diameter D3, as compared to the outer diameter D5 of the diffuser passage and of the volute D6.

In the example, the impeller 201 has an outer diameter of approximately 18.1 centimetres. As previously described, the relatively large diameter of the impeller has certain performance advantages. In particular, the large impeller diameter advantageously allows a relatively high pressure rise to be achieved at a relatively low rate of rotation of the impeller. Consequently, the required flow rates and pressures may be achieved at a reduced rotational rate of the impeller compared to a lesser diameter impeller. In the example this advantageously reduces the power draw of motor 216.

Moreover, it has been found that an impeller with a diameter of approximately 18.1 centimetres is particularly suited to an application in which the compressor is used as an HVAC blower in a vehicle, for the reason that it can achieve mass flow, pressure rise and efficiency rates typically required of a vehicle HVAC blower, without being excessively large in dimension. Impellers with an outer diameter in the range of approximately 17 centimetres to approximately 19 centimetres have been found to be particularly suited to such an application.

It has been found further that for many applications, the relatively large diameter impeller may be usefully paired with a relatively short diffuser passage. Thus, an acceptably compact overall outer diameter dimension may be maintained even using a relatively large impeller.

As is known, the degree of diffusion of the fluid in the diffuser stage is influenced by, amongst other factors, the radial length of the diffuser passage, that is the distance between the diffuser passage inlet at the inner diameter and the diffuser passage outlet at the outer diameter. It is generally expected that a decreased length will generally result in a lower degree of fluid diffusion in the diffuser, and so lessen pressure recovery.

It has been found however that an advantageous balance of flow rate, pressure rise, rotational rate, and efficiency is found where the ratio of the outer diameter of the impeller to the outer diameter of the diffuser passage is in the range of approximately 0.85 to approximately 0.95, that is to say where the impeller outer diameter is in the range of approximately 85 percent to approximately 95 percent of the diffuser passage outer diameter.

In the example the diffuser passage is configured to have an inner diameter D4 of approximately 19 centimetre, and an outer diameter D5 of approximately 20 centimetre. As compared to the impeller outer diameter of approximately 18.1 centimetre, the ratio of the impeller outer diameter D3 to the diffuser passage outer diameter D5 is approximately 0.9. Values close to this relationship have been found to advantageously balance flow rate, pressure rise and efficiency characteristics.

In the example, the volute has an average outer diameter dimension D6 of approximately 25.7 centimetre. It has been found that the volute should ideally have an outer diameter that is at least 140 percent of the impeller outer diameter to optimise efficiency.

The relatively small outer diameter of the diffuser passage 401 and the volute 213 advantageously allow the relatively large diameter impeller to be packaged in a smaller space than would be possible for conventional compressor designs. As previously described, the decreased outer diameter of the volute is accommodated for by the upper portion of the volute being shaped to extend radially inwardly over the impeller to thereby maintain volute volume. The relatively low outer diameter of the volute in comparison to the large outer diameter of the impeller has the advantage that the compressor is dimensionally compact yet still achieves good performance. The dimensions of the specific example described herein have been found to be particularly suitable where the compressor is employed as a vehicle HVAC blower.

Referring finally to Figure 6, as described previously, it is known that the operational characteristics of a centrifugal compressor, for example, the maximum flow rate, pressure rise and pumping efficiency, are functions of, inter alia, the various geometries of the centrifugal impeller. It is generally understood in this respect that increased flow rates and/or pressure rises may be achieved by increasing the outer diameter or the rotational rate of the impeller.

However in many applications packaging constraints or other practical considerations will typically limit the maximum size or rotational rate of the impeller. Use as an automotive HVAC blower is one such application.

The present inventors have investigated varying the relative areas of the impeller inlet and outlet openings as a means to improve impeller performance characteristics for a given impeller outer diameter. It has been found in this respect that the ratio of the inlet opening area to the outlet opening area of the impeller has a particularly significant effect on the compressor performance. As may be expected, it has been confirmed that increasing the inlet opening area increases the maximum flow rate achievable before inlet stall and/or other flow limiting aerodynamic conditions occur. That is to say, it has been observed that increasing the area of the inlet opening in comparison to the area of the outlet opening may generally increase the flow rate of an impeller for a given impeller outer diameter and volute/housing geometry, whilst substantially maintaining pressure rise.

However, it has been found that the efficiency of the impeller is highly dependent on the inlet-outlet opening area ratio. Specifically, it has been found that, for at least one operating point, the efficiency of the compressor is negatively affected if the inlet-outlet opening area ratio is too low or too high. Moreover, it has been found that the effect on performance of an inlet-outlet opening area ratio that is too low may tend to be even more severe than for a ratio that is too high.

Figure 6 shows a schematic plot of the efficiency of a centrifugal compressor as a function of the ratio of the impeller inlet opening area to outlet opening area. The X-axis defines the inlet-outlet opening area ratio, whilst the Y-axis defines the isentropic efficiency of the test compressor. The efficiency data reflects the efficiency of the compressor when operating to achieve a flow rate of approximately 150 litres per second (1/s) and a pressure rise of approximately 1500 Pascal (Pa). In the experiments the impeller is rotated at a rate of approximately 5000 rpm. In the tests the compressor had a construction substantially the same as that of compressor 106 as described herein with reference to Figures 1 to 5, save for the geometry of the impeller which was varied. Specifically, the geometry of the diffuser passage and volute were that of compressor 106.

Referring to the Figure, a sharp increase in the efficiency of the compressor is observed by increasing the inlet-outlet opening area ratio from approximately 0.74 to approximately 0.83. However, further increasing the inlet-outlet opening area ratio beyond this relationship results in a, more gradual, reduction in efficiency.

It has thus been found, as demonstrated in the Figure, that the impeller operates particularly efficiently at the test operating point where the ratio of the inlet opening area to the outlet opening area is in the range of approximately 0.78 to approximately 0.96, and even more preferably in the range of 0.79 to 0.89. It can be seen that the drop10 off in efficiency for area ratios outside of this range is relatively severe. Moreover, it can be seen that at the operating point the highest efficiencies occurred for an inlet opening area to outlet opening area ratio in the range of 0.80 to 0.85.

Claims

1. A centrifugal compressor comprising:

a centrifugal impeller;

a volute for collecting fluid expelled by the impeller; and a diffuser passage extending between the impeller and the volute;

wherein the ratio of the impeller outer diameter to the diffuser passage outer diameter is in the range of 0.85 to 0.95.

2. The centrifugal compressor of claim 1, wherein the ratio of the impeller outer diameter to the diffuser passage outer diameter is at least 0.88.

3. The centrifugal compressor of claim 1 or claim 2, wherein the ratio of the impeller outer diameter to the diffuser passage outer diameter is less than 0.92.

4. The centrifugal compressor of any one of claims 1 to 3, wherein the ratio of the impeller outer diameter to the diffuser passage outer diameter is in the range of 0.89 to 0.91.

5. The centrifugal compressor of any one of claims 1 to 4, wherein the outer diameter of the impeller is in the range of 17 centimetres to 19 centimetres.

6. The centrifugal compressor of any one of claims 1 to 5, wherein the outer diameter of the impeller is in the range of 17.9 centimetres to 18.3 centimetres.

7. A vehicle comprising a passenger cabin for accommodating passengers and a passenger cabin ventilation system including the centrifugal compressor of any one of claims 1 to 6.

8. A vehicle according to claim 7, comprising ducts for ducting air from outside the passenger cabin to an inlet of the compressor, and for ducting air from an outlet of the compressor to the passenger cabin.