The compression of gases may be achieved by use of a variety of machines such as turbo machines or positive displacement machines. The most compact and simple types of device are turbomachines. They are very compact relative to positive displacement types. The gas is pumped in a continuous flow at very high speeds through the machine, typically at speeds in the order of a large fraction of the speed of sound. In addition, these machines do not have rubbing surfaces (unlike positive displacement machines) and so contamination of the pumped gas by lubricant can be relatively easily avoided.

The use of turbomachines is virtually universal for high gas flow rates whereas for very low flow rates positive displacement machines are normally the only practical option.

However, for intermediate flow rates (typically 1-50 litres/sec), positive displacement machines tend to be large and heavy whilst the use of turbomachines is difficult for the reasons described below. If the volume flow rate for a turbomachine is in this range and the enthalpy rise is high (typically a pressure outletlinlet ratio of 2: 1 to 10: 1), the minimum shaft speed for the device tends to be high. Such conditions of gas flow rates and pressure ratios are typically required in the following applications: Compression of fuel gas for a micro, mini or small gas turbine engine Compression of air for a small fuel cell 'Compression of sub atmospheric air to atmospheric pressure in the case of a vacuum pump.

'Compression of air for a small microturbine engine (e. g. sub 20 kW) Compression of air for a small supercharged or turbocharged internal combustion engine.

The reasons for the difficulty in designing turbomachines for the range of flow rates and pressure ratios previously described will now be explained. For a given volume flow rate of gas and a given enthalpy rise to be achieved during compression, a turbomachine compressor must be designed to operate within a narrow speed range in order to achieve high efficiency. Typically this speed range must lie within around 30% of the optimum speed for the particular type of machine be it a centrifugal, a mixed flow or an axial type of turbomachine. The principles behind this are well known in the art and adequately described by the use of'specific speeds'.

For example, a centrifugal turbocompressor for use in compressing atmospheric air to a pressure of 500kPa (5 atmospheres) at an inlet volume flow rate of 5 litres per second would be required to operate at a speed of around 625,000 rpm. This is based on a single stage centrifugal compressor designed to operate at the lowest specific speed of 0.5. Furthermore, of all the standard types of turbomachine, including centrifugal, mixed flow and axial types, a centrifugal compressor operating at its highest efficiency has the lowest specific speed. This need for high speed creates a number of difficulties, specifically: It is difficult or impossible to find a mechanism to drive the compressor, particularly if this is to be by means of an electric motor; Providing bearings for the shaft is very difficult at such speeds; and The physical dimensions of the turbomachine are so small that it is difficult and expensive to manufacture and very difficult to create effective sealing to prevent leakage over the blades and in other places between the rotor and stator.

One way to improve these problems and reduce the necessary speed is spread the compression over a number of stages so that the pressure and hence enthalpy of each step is less. If compression is carried out over six stages, the shaft speed of the example above can be reduced to around 250,000 rpm. Such speeds are still far too high to be practical and it is difficult to mount this many stages on a shaft without encountering rotor dynamics problems. Rotor dynamics problems occur if a shaft is long relative to the speed required. This can lead to vibration of the shaft which may become unstable.

If the shaft speed chosen for the design is below the recommended minimum, windage losses (that is losses due to fluid friction and fluid pumping when a body rotates in a fluid) and flow leakage, around the interfaces between the moving rotor parts and stationary stator parts, become high leading to unacceptably low efficiency.

There are exceptions to this general rule, such as the regenerative helical flow machine (Sixsmith et al US Patent No. 4,325,672). However, whilst these machines have reduced windage losses, they still have problems with low efficiency since leakage is particularly high.

The problem of leakage flow can be reduced by shrouding the blades but this creates even more surface area for windage losses.

Therefore, according to the present invention there is provided a compressor comprising: a rotor having a plurality of nested rotor channels, each channel having an axial inlet and a radial outlet, and a stator comprising one or more stator channels for directing the gas from the outlet of one rotor channel to the inlet of another rotor channel.

The compressor is preferably generally circular and, apart from the innermost inlet, is arranged around one of the other inlets. The outlet is preferably generally annular, with the outlets arranged side by side axially along the rotor. To minimise the size and cost of the rotor, at least one of the walls of each rotor channel forms one of the walls of at least one other rotor channel.

The stator channels may be arranged to direct gas from the outlet of one rotor channel to the inlet of another, radially outer, rotor channel. However, they may also be arranged to direct gas from the outlet of one rotor channel to the inlet of another, radially inner, rotor channel. Each of the rotor channels are preferably provided with a plurality of vanes to accelerate the air entering the rotor and drive it radially out of the rotor. The rotor vanes are preferably thin but they may increase in thickness towards the radially outer periphery of the rotor. In an alternative construction, the rotor vanes are hollow. The rotor vanes are advantageously contained between two substantially radial walls.

Although the stator channels may be vaneless, preferably one or more of the stator channels comprises a plurality of stator vanes. The stator vanes may be thin although as with the rotor vanes, they may increase in thickness towards the radially outer periphery outer part of the stator. The stator vanes may also be hollow. The stator vanes are generally radial and may be pinched in.

Seals are generally required between the rotor and stator. These may be one or a combination of labyrinth seals, plain seals, stepped seals and angled seals. These may additionally include one or more of brush seals, sprung seals and weir seals.

The seals may be radial of axial or alternatively arranged at an angle of between 0 and 90 degrees relative to the longitudinal axis.

The rotor and stator are preferably axially relatively movable to allow the seals to be opened up particularly during start up.

The compressors may be arranged with their driving motor in any number of configurations. Preferably, the motor is mounted on the same shaft although it may be driven via a gearbox. The compressor and motor may be formed in a single unit which is sealed particularly when the pumped fluid is flammable or toxic.

The compressor may be mechanically separate from the motor, power being transferred by a magnetic coupling. Two or more compressors may be connected in series to increase the pumping efficiency of the system. These may be driven by a single motor.

A specific embodiment of the present invention will now be described in detail with reference to the drawings in which: Figure 1 shows a cross-sectional view of the radial regenerative compressor of the present invention connected to a drive motor and drive circuitry; Figure 2 shows a radial cross-section through a typical stage of a radial regenerative compressor of the present invention; Figure 3 shows a radial cross-section through a second embodiment of a radial regenerative compressor of the present invention; Figure 4 shows an alternative type of rotor vane according to the present invention; Figure 5 shows a radial cross-sectional view of the rotor vane of the radial regenerative compressor of figure 4; Figure 6 shows a cut view of the last stage of the compressor of figure 4 just before the gas enters the vanes; Figure 7 shows a radial cross-section through a third embodiment of a radial regenerative compressor of the present invention; Figure 8 shows an alternative construction of the third embodiment; Figure 9 shows a radial cross-section through a fourth embodiment of a radial regenerative compressor of the present invention; Figure 10 shows an alternative configuration of the present invention; Figure 11 shows the rotor of figure 1 with the gas seals visible; Figure 12 shows an alternative design for the seals; Figures 13 to 16 shows a number of further alternative designs for the seals; Figure 17 shows the shaft of the compressor mounted on an alternative type of bearing ; Figure 18 shows the shaft of the compressor driven by a motor by means of a non- contact magnetic; Figure 19 shows an arrangement in which two compressors of the present invention are driven by means of one electric motor; and Figure 20 shows a arrangement in which an electric motor and the compressor rotor are mounted on one shaft.

Figure 1 shows a schematic representation of a compressor system comprising a radial regenerative compressor 1 driven by a high speed motor 2 with associated drive. As can be seen from the cross sectional view of the compressor, the rotor 10 is mounted on a shaft 12 in bearings 13, shown as rolling element bearings, by way of example only.

The shaft is driven by means of the electric motor 2 causing the compressor rotor to rotate.

The rotor 10 of the compressor 1 comprises a series of nested channels provided with axial entrances and radial exits and having a plurality of blades or vanes arranged within. The rotor 10 is arranged generally within a stator 11 also comprising a number of nested channels, each arranged to receive gas from the outlet of one of the rotor channels and to carry the gas around to the inlet of a different rotor channel. When viewed in longitudinal cross-section as shown in figure 1, the stator channels having a generally spiral shape. Whilst the stator sections are shown having a square section, they may also have any suitable section e. g. a toroidal cross-section.

Gas enters the inlet 100 of the compressor 1 co-axial with the axis of the shaft and is drawn into the vanes or blades 101 situated in the first radial passage 105 of the rotor.

A typical section of the rotor blades is shown in figure 2, in which it can be seen that the rotor blades 101 are angled to accept the incoming gas at the correct angle. As the blades act on the gas, the gas is accelerated tangentially. The gas then passes generally radially out of the rotor albeit with a substantial tangential velocity. The gas then passes into the row of diffuser vanes 111 within a first radial channel 115 of the stator whereby a significant proportion of the kinetic energy in the gas is converted to a static pressure rise. The workings of the machine so far described are well known as a centrifugal or radial compressor and can be denoted as a first stage of compression.

However, in the construction of this invention, the gas is then guided in the channel 115 such that it returns to a new passage 106 in the rotor which contains another set of vanes 102 similar to those used in the first stage of the machine. The pressure of the gas is thus further increased by this second stage. The process is repeated through each of the stator channels 116,117 and rotor channels 107,108 respectively until the gas enters the final stator channel 118 and fed to the outlet 119 of the compressor shown as pipes in Figure 1 by way of illustration only.

For aerodynamic reasons, the vanes 111-114 in the passages 115-118 in the stator may be pinched-in. In this case, the axial width of the diffuser around the vanes is reduced.

Although figure 1 shows four stages of compression, the compressor may have any number of stages according to the desired characteristics of the machine.

Seals are required between the rotor and stator parts in order to reduce or prevent gas leakage wherever there is a significant pressure difference between adjacent passages.

The seals are not shown detail in figure but will be described in more detail later.

As indicated above, one of the major difficulties in running compressors at high efficiency is the problem of windage losses. In machines which have separate rotors stages, each rotor has external parts which are rotating in generally stationary gas. The motion of the rotor causes this gas to move drawing power from the rotor which is not useful. In the present invention, the surface area of the rotor adjacent to stationary surfaces is considerably reduced because the external parts of most of the rotor channels (i. e. all except the external ones) have their outer sides in a flow of gas which is travelling at high speed. This reduction in rotor surface area which is adjacent to stationary surfaces, due to the nested design of the present invention, ensures a considerable reduction in windage losses. In addition, rotor surfaces which are adjacent to other rotor surfaces produce substantially lower windage losses and much of the rotor surface is in this category. The windage of the invention is thus a small fraction of a compressor with separate rotor stages. An additional benefit is the compactness of the machine relative to a machine with separate stages. This also means that rotor dynamic problems are reduced and so it easier to maintain small clearances on the seals and hence reduce leakage.

Figure 2 shows a wedge type diffuser ring 20 comprising a plurality of vanes 101 etc., which can be used with this machine in order to decelerate the gas coming out of the rotor channels and recover kinetic energy into static pressure. It is possible to remove the vanes altogether and have a vaneless diffuser. Whilst the diffusion effect is not as great, the construction of the stator is simplified. As the walls of the channels are supported by the vanes, if the vanes are omitted, the casing must be supported using pillars or the like. It is preferable to place such pillars at the greatest radius possible such that the gas velocity and hence losses will be minimised.

Figure 3 shows a second embodiment of the present invention utilising an alternative cascade type diffuser arrangement of the diffuser blades 105-108.

Figure 4 shows an alternative type of rotor vane in which the vanes have a substantially axial leading edge. Figure 4 is a cross-sectional view of a radial regenerative compressor incorporating such a rotor vane. The vanes are thus extended such that a sectional view will cut the vanes as shown in figure 5. Figure 7 is a partial cut view showing the last stage of the compressor just before the gas enters the vanes.

According to a third embodiment of the present invention the vanes on the rotor may be replaced with generally wedge shaped sections 70 to form passages 71, referred to herein as partial passages, as shown in Figure 7. Here, the vanes are made excessively wide beyond what is required for structural reasons. In this way, the axial passage height can be increased for a given gas flow rate. This allows the gas to enter the passages more smoothly by setting the leading edge of the vanes at an appropriate angle but where the modified vanes 70 thicken quickly to create the partial passage 71 effect.

The flow is decelerated in the diffuser in the same manner as before. The advantage of this aspect of the invention is that the rotor axial passage height or heights can be increased making the machine easier to manufacture. The presence of the modified vanes 70 also makes the construction of the rotor easier since there is an unused space 73 to pass bolts though or material available which can be used in order to provide a welded joint.

Figure 8 shows an embodiment of the partial passage rotor in which material from the thick vanes has been removed leaving a hollow 72. This lightens the rotor and reduces centrifugal stresses in the material.

Figure 9 shows a further modification of the third embodiment in which the diffuser also has modified vanes 90 to provide partial passages 91 used for similar reasons as in the rotor. This modification of the diffuser vanes of the stator may be incorporated in other constructions including those where'normal'vanes are used in the rotor.

In the foregoing description and drawings, the passage of the gas has been described and shown in terms of the gas entering the compressor along its axis and passing through compressor sections which are progressively radially further out until it exits the compressor from the radially outermost compressor stage. However, it is envisaged that the compressor could be arranged in any number of alternative configurations where the gas does not proceed from one stage to the next radially outer stage.

Consecutive stages may be arranged in any order. For example, as shown in figure 10, the incoming gas enters the radially outermost (right hand in figure 10) stage first and passes to adjacent rotor stages radially inner and to the left of the previous stage. The gas is ultimately passed from the machine via an outlet arranged in the final channel of the stator, as shown in figure 10.

Figure 11 shows an enlarged version of the compressor shown in Figure 1 in which the gas seals 40 are clearly visible. In the construction shown, the seals are of the labyrinth type. The seals 40 are in an annular orientation, i. e. the outer peripheral edge of the rotor sections forma generally annular seal against the inner peripheral edge of the stator. An advantageous variant of this seal arrangement is shown in Figure 12. Here, the seals 41 at the rotor exit are in a radial orientation. Detail of the seals is shown where a typical labyrinth is illustrated. The number of fins on the labyrinth can be as many as can be afforded by the space available. However, for illustration purposes, only five fins are shown here. The seals 40,41 on the walls between the exit channels of the rotor are particularly important since the diameter of these seals is large relative to the seals 42,43 on the walls between the entry channels of the rotor. Hence the leakage area is likely to be greater on these seals 40,41 than the seals 42,43 on the walls between the entry channels. However, the seals shown in Figure 12 may be used for this region if desired.

Figure 13 shows a version of a weir seal 45 which can be used in place of a labyrinth seal at any place in the compressor. Here a liquid 46 is trapped in a groove 47 on the rotor 10 and a stationary radial fin 48 is provided on the stator 11 which dips into the liquid preventing gas from passing across the seal.

Figure 14 shows a version of a disc seal 50 which can be used in place of a labyrinth seal at any place in the compressor. On the stator 11 is mounted a ring of material 51, typically carbon which is sprung loaded 52 towards the rotor 10. This forms a tight seal when the rotor is stationary. However, when the rotor is in motion, the ring 51 lifts off from the rotor surface due to aerodynamic forces. Only a small clearance results thus gas leakage can be very small.

Figure 15 shows a variant of the labyrinth seal which incorporates steps 60 in order to further reduce leakage relative to the plain labyrinth seal shown in Figure 12. This type of seal can be implemented with ease when the seal orientation is radial.

Figure 16 shows a labyrinth seal with the addition of a brush seal 61 which can be used in place of a labyrinth seal at any place in the compressor. The brush seal 61 can be used either with a labyrinth seal as shown in Figure 16 or on its own.

For all the seal types, it may be beneficial to withdraw the seals during start up of the machine and allow them to go back in place once a stable speed has been reached.

During start up, vibration is often higher than during stable running. This vibration may cause damage or high wear to the seals. By withdrawing the seals during start up, such damage can be avoided and this allows for tighter clearances and lower leakage.

Withdrawal of the seals can be achieved by moving the rotor and stator axially relative to each other a small distance. In the case of seals substantially in the radial orientation, the stator containing the seals can be moved axially by means of the pressure delivered by the compressor. Thus when the machine is at a low speed, the seals are kept away from the rotors by means of a spring load but as the machine speeds up, pressure from the compressor overcomes the springs and moves the seals to a predetermined stop at the minimum clearance possible.

Figure 17 shows the shaft of the compressor mounted on an alternative type of bearing which is a fluid film type, typically oil lubricated, or gas bearings. Whatever the type of bearings used, the shaft onto which the rotor is mounted can be driven by the motor by any type of compliant coupling.

Figure 18 shows a design in which the coupling is a magnetic face type such that there is no physical contact between the motor shaft and the compressor shaft. If desired, this allows a membrane of non-conducting material to be placed in the gap between the coupling halves thus hermetically sealing the rotor shaft from the outside. This is particularly advantageous where the pumped fluid is a flammable fuel gas or is toxic. A cylindrical coupling may be used instead of a face type coupling.

Figure 19 shows an arrangement in which two compressor stages of the said machine are driven by means of one electric motor. This is advantageous if more compression stages are required in order to reach a certain pressure ratio and all these stages cannot be nested together for reasons of mechanical or geometric limitations.

The compressor can also be driven by means of an electric motor via a step up gearbox (not shown). The compressor can also be driven by a turbine such as would be the case in the application of a micro gas turbine engine (not shown).

Figure 20 shows a design in which the electric motor and the compressor rotor are mounted on a common shaft. The motor and compressor may be hermetically sealed as one unit. This is advantageous where the pumped fluid is a flammable fuel gas or is toxic.