COMPRESSION METHOD AND MEANS

This application relates to the field of gas pumping and compression.

Background of the invention

Gas compression devices used in refrigeration, air conditioning and industry consume a large portion of electrical power generated. An increase in gas pumping efficiency will result in reduction of carbon dioxide emissions. Proposals to sequester carbon dioxide at pressure underground or in the ocean depths are dependent on using compression methods that are efficient and can also overcome problems such as phase change and the material erosion of compressor parts when compressing impure gas mixture. Small changes in compressor efficiency may determine whether carbon sequestration is commercially viable.

Efficient compression requires that as little kinetic energy as possible is imparted to the gas molecules. This implies that a gas packet should move as slowly as possible through the compressor, without sudden accelerations. The direction of motion should preferably be in a straight line. There should be no sudden changes of volume that might lead to shock formation at high speeds. Since gas noise is initiated by the kinetic energy of the gas, a compressor that compresses smoothly and gently will be quiet in operation.

Heat of compression spreading back by leakage or thermal conductance to the intake gas or to less compressed gas in the compression chamber is a cause of inefficiency in compressors requiring additional work energy equal to any increased heat acquired by the gas both before entering and within the compressor. Some of the heat of compression flows into the walls of the

compressor. Because this heat cannot be removed quickly enough, the walls remain hot, heat remains in the gas and the work required rises. It is desirable, therefore, that as gas is compressed and heated some or all of the heat of compression should be removed from the gas as it moves through the compressor. Conventionally this is done by cooling the compressor walls or by water injection. However in piston and cylinder compressors the gas is compressed into a volume defined by unchanging but decreasing surface, therefore there is little possibility of removing the heat of compression during the process. The higher the rise the greater is the loss of efficiency.

Known types of compressors typically suffer from problems which tend to reduce efficiency, including but not limited to those described herein, namely:

• impartation of large amounts of kinetic energy to the gas being compressed • sudden acceleration of gas leading to high noise levels and energy losses

• high gas flow speeds leading to frictional heating of the gas being compressed, leading to an increased work requirement

• heat of compression feeding back to the intake charge, leading to an increased work requirement • variable internal surface area leading to a reduced ability to remove heat of compression from the gas being compressed

• high rubbing speeds between internal components leading to wear and frictional losses

• low inter-stage compression rise • large physical size relative to gas processing rate

Brief description of invention

The invention is set out in the claims.

In an embodiment of the invention there is provided a compressor comprising a cylinder and a rotor, whereby the rotor traverses the internal circumference of the cylinder and a pinch point is formed at the closest point of the rotor periphery to the internal wall of the cylinder. The rotor traverses the internal circumference of the cylinder such that the pinch point moves at high, preferably supersonic speed. In an embodiment, the rotor rolls around the internal circumference of the cylinder such that the speed of the rotor surface, relative to the cylinder wall, is low or zero, thus reducing wear and factional heating of the components and of the gas to be compressed, termed herein "rolling", thus aiding compressor efficiency. Optionally, a strip valve arrangement on the rotor surface allows entry of gas into the chamber formed between rotor and cylinder. Optionally, a strip valve arrangement on the cylinder wall allows exit of gas from the chamber and optionally incorporates actuation means to control its opening position.

In another embodiment of the invention there is provided a compressor comprising a cylinder and a rotor, whereby the rotor traverses the internal circumference of the cylinder and a pinch point is formed at the closest point of the rotor periphery to the internal wall of the cylinder. The rotor moves such that the pinch point moves at high, preferably supersonic speed. The rotor rotates around the internal circumference of the cylinder such that a fixed point on the rotor periphery is maintained adjacent to the pinch point - termed herein "rotating". Optionally, ports in the rotor allow entry and exit of gas via passages communicating with the axial ends of the cylinder.

In a further embodiment, the rotor orbits around the cylinder such that there is no rotation of the rotor itself- termed herein "orbiting".

Embodiments of the invention incorporate valve means for allowing compressed gas into the chamber and for allowing compressed gas out of the chamber.

Embodiments of the invention incorporate an arrangement for adjusting the running clearance between the rotor surface and the inner circumference of the cylinder.

Embodiments of the invention incorporate an arrangement for counterbalancing the rotating mass of the rotor and associated rotating parts. Such counterbalancing means can optionally be adjustable.

Embodiments of the invention incorporate rotor surface features in order to increase compressor efficiency.

Embodiments of the invention will now be described, by way of example, with reference to the figures which are as follows:

Fig. 1 - Schematic view of compressor housing and rotor Fig. 2a to 2g - 'Rolling' rotor operation Fig. 3a to 3g — 'Orbiting' rotor operation Fig. 4a to 4g - 'Rotating' rotor operation Fig. 5 - Strip valve arrangement Fig. 6 - Rotor port arrangement Fig. 7 - Rotor surface features Fig. 8 - Spiral duct embodiments

Fig. 9 - Rotor balancing/drive arrangement Fig. 10 - Strip valve and balancing arrangement Fig. 11 - Strip valve actuation arrangement

Detailed description

As shown in Fig. 1, the present invention provides a compression method that has the desired characteristics of smooth compression and internal cooling of the gas. This method employs a cylindrical chamber (10) and rotor or orbiter (20) to create a moving duct or chamber (40) of unchanging geometry and size, whose walls converge relative to a static gas packet drawn into the moving duct (40). The duct (40) walls converge at a lower speed than the point of closest approach of the walls [hereinafter called the pinch point (50)] moves along the duct (40). In preferred operation the closing speed of the walls is subsonic and the speed of the pinch point (50) is supersonic. As the pinch point (50) advances, the volume in which gas is at highest pressure/temperature also advances to areas of the walls that have been cooled since last being adjacent to the high temperature gas. When such a compressor is operating with the pinch point (50) moving at supersonic speeds, information about the pressure rise caused by narrowing of the duct (40) cannot propagate forward and push the gas forward. This enables high pressure to co-exist, at the narrowing end of the duct (40), with low pressure elsewhere in the duct (40) because the volumes are physically separated by the pinch point (50) and the pressure information barrier (40) produced by the supersonic advance of the pinch point (50). This provides a compressor that has the high pressure ratio capability of positive displacement compressors combined with the smooth pulse-less outflow of centrifugal and axial machines.

The invention can be realised in various embodiments by employing a duct (40) created between an inner circumference of a cylinder (10) and a shaped wall (20) moving within the cylinder (10) so as to form a narrowing of the duct (40) at the point of closest approach of the two members (50).

Three embodiments demonstrating variations on the movement of the rotor (20) within the cylinder (10) will now be described.

As shown in Figs. 2a to 2g, as can be used in a first class of embodiments described below, a 'rolling' rotor (20) rolls around the inner circumference of the cylinder (10) as the rotor (20) traverses the inner circumference of the cylinder (10). The orientation of the rotor (20) is shown by respective arrows A, B, C in Fig. 2a. The sequence of six illustrations shown consecutively in Figs 2b to 2g illustrates (see arrow A in each) how the orientation of the rotor (20) changes with respect to the cylinder (10) as the rotor (20) rolls around the inner circumference of the cylinder (10). The rotor changes orientation as it rolls such that the speed of the rotor (20) surface, relative to the surface of the inner circumference of the cylinder (10) is substantially low or zero. The rotor (20) can be arranged to substantially contact the inner surface of the cylinder (10) or the two surfaces can be spaced slightly apart. The rotor (20) can be arranged to roll by means of contacting the inner surface of the cylinder (10) or can be rotated by other means such as gears or by entrainment by the gas being compressed. This feature results in a substantially low or zero rubbing speed between the surface of the rotor (20) and the inner surface of the cylinder (10), which in turn results in improved wear performance of those surfaces. Other results of this feature are lower frictional losses, lower kinetic energy imparted to the gas being compressed (lower entrainment) and lower frictional heat imparted to the gas being compressed. These results all contribute to greater efficiency of the compressor.

As shown in Figs. 3a to 3g, as can be used in the first class of embodiments described below, an orbiting rotor (20) does not change orientation with respect to the cylinder (10) as the rotor (20) traverses the internal circumference of the cylinder (10). Fig 3a shows sequential position 20a, 20b, 20c and corresponding orientations with arrows A, B, C. Figs 3b to 3g show the sequential rotor positions and corresponding orientation A. An orbiting rotor (20) results in a greater relative speed between the surface of the rotor (20) and the inner surface of the cylinder (10) than with the rolling rotor (20) of Fig. 2, but a lower relative speed than with a rotating rotor (20) as will be described in the following paragraph. Efficiency losses when an orbiting rotor (20) is employed tend therefore to be in a range between those of the rolling rotor (20) and those of the rotating rotor (20).

As shown in Figs. 4a to 4g, as can be used in the first class of embodiments or a second class of embodiments described below, the rotating rotor (20) changes orientation as the rotor (20) traverses the internal circumference of the cylinder (10), in such a way that a fixed point on the rotor (20) surface A, B, C in the sequential positions 20a, 20b, 20c in Fig 4a is adjacent to the pinch point (50). The movement of point A can be seen in the sequential position shown in Figs 4b to 4g. A rotating rotor (20) results in a greater relative speed between the surface of the rotor (20) and the inner surface of the cylinder (10) than either the rolling rotor (20) of Fig. 2 or the orbiting rotor (20) of Fig. 3. Efficiency losses when a rotating rotor (20) is employed tend therefore to be in a range which is higher than those of the rolling rotor (20) of Fig. 2 or the orbiting rotor (20) of Fig. 3. An advantage of the rotating rotor (20) of Fig. 4 is that a greater range of valve arrangements can be practically used than with the other two rotor (20) types. A compressor incorporating the rotating rotor (20) can be

made with fewer moving parts than a compressor incorporating the other two types of rotor.

As shown in Fig. 5, in a first class of embodiments, the duct (40) is a chamber formed between two cylinders, one relatively static (10) and acting as a stator and the other (20) acting as a rotor — rolling, orbiting or rotating it within it. Using a valving mechanism described below, gas is drawn into the duct (40) by a rarefaction caused by the widening of one end the duct (40) (i.e. when the rotor is adjacent an opposing side of the stator). It passes through inlets in the walls of either of the cylinders (10, 20) or of the end walls and is expelled at higher pressure at the other end of the duct (40) after it has been compressed by a relative narrowing of the duct (40) caused by the orbiting component (20) approaching the stator wall. By mounting blades inside the rotor (20), a degree of pre-compression can be achieved. In such an embodiment the rotor (20) may have a rolling or rotating surface or may orbit without rotation.

In a device built according to such a first class of embodiments, as shown in Figs 5 and 10, there is provided a cylindrical rotor (20), within a cylinder (10). The rotor (20) is provided with a surface channel (210), of depth equal to the thickness of strip (220) that fits within the channel (210). The strip (220) is of larger circumference than the rotor (20) circumference, so that when the strip (220) is pressed onto the rotor (20) it forms a gas tight seal. However because the strip (220) is of larger circumference than the rotor portion (20), the strip (220) will always protrude above the rotor (20) surface circumference, away from the final point by virtue of the squeezing force exerted on it there, allowing gas to flow through openings (230) in the base of the channel (210) to outside the rotor (20).

The cylinder (10) is similarly provided with a channel (240) and strip (250) on the outside, allowing gas to pass from the inside of the cylinder (10) to ducting means on the outside. This outlet strip (250) may be provided with reinforcement across its width to support it against high gas pressures.

In operation the rotor (20) orbits, preferably at a speed that results in the pinch point (50) between the rotor (20) and cylinder (10) rotating at supersonic speed. As the pinch point (50) rotates, low pressure follows behind (in terms of the direction of rotation) the pinch point (50), pulling the strip valve (220) away from the rotor (20) and continuously inducing gas into the chamber (40). At the other end of the chamber (40) the converging surfaces of rotor (20) and cylinder (10) compress previously inducted gas and force it out of the chamber (40) through the exit strip (250), which is forced to and held in an open position by the pressure of gas in front of the pinch point (50). To prevent the exit strip (250) overlapping the pinch point (50) and allowing gas to escape from the high pressure volume into the low pressure volume, the exit strip (250) may be actuated by mechanical, electrical or magnetic means to control the distance of its opening (270) from the pinch point (50). As shown in Fig. 11, a cam (261) on the drive shaft (660) operates a pushrod (260) which operates to lift the exit strip (250). This actuation is also helpful for controlling start and shutdown conditions and to give a degree of capacity control. In rolling or rotating rotor embodiments (see below) where the high pressure exit side of the pinch point (50) can be separated by some distance from the low pressure inlet side, actuation of the strip (250) may be used to restrict the area of the outlet by moving the opening (270) partially past the pinch point (50) and so controlling the pressure ratio of the device. Hence, in an embodiment, a strip is deformable by mechanical actuation, in particular by an actuator such as a cam and pushrod coupled to the rotor, for example the rotor drive shaft.

As shown in Fig. 6, a blind passage (275) or passages are provided within the rotor, open on the axial face and terminating adjacent the inside of the rotor surface. This passage (275) communicates with the axial face of the rotor (20), so that cooling fluid may be circulated behind the rotor (20) circumferential surface. The walls and end plates of the chamber (10) are additionally provided with passages for the circulation of cooling fluid. Finned means (276) may be provided to increase the heat flow from the chamber (10) walls to be cooled into the cooling fluid.

In operation the rotor (20) is rotated with the inlet conduit (330) leading, so that the duct (40) rotates with the rotor (20) at a speed preferably in excess of the local speed of sound. Appropriate curvature of the inlet conduit (330) passage way causes gas to be drawn from an axial face of the rotor (20) into the conduit (330) in a substantially radial direction. As the duct (40) - i.e. the space between rotor and stator — rotates around the stator, the gas is confined to a converging duct (40) formed between the surface of the rotor (20) and the cylinder (10) wall. The supersonic speed of the approaching pinch point (50) does not give time for information about increasing pressure to propagate upstream. The gas is steadily compressed until, as the pinch point (50) reaches the gas, it is permitted to escape at high pressure through the outlet (340) and through passages within the rotor (20) to a radial end of the rotor (20) from whence it is ducted out of the device. During compression in the duct (40) the gas temperature increases. The heat of compression is transferred continuously, both through the wall of the rotor (20) into the cooling fluid circulating behind the wall and through the chamber walls (10) into the cooling fluid (277) circulating there.

As shown in Fig. 7 the surface of the rotor (20) may be provided with spiral grooves (400) and/or passages (410) to conduct high pressure gas that passes

the pinch point (50) or along the axial ends of the chamber back to a selected or controlled point (420) in the duct (40). This gas is cooled on its passage back to the chamber and this is more advantageous for the efficiency of the device than allowing it to re-emerge at the inlet end of the chamber (330). In complex devices it would be possible to bleed this gas through micro pores (430) in the rotor surface to promote laminar flow.

In this second class of embodiments the device may include a rotor (20) with the converging duct (40) formed between the cylinder (10) wall and the surface of the rotor (20) (as shown in Fig. 6) or the converging duct (40) may be formed between the cylinder wall (10) and a channel of reducing cross-section (330) on the rotor (20) where the rotor is concentric with the stator (as shown in Fig. 8).

Referring, for example, to Figs 7 and 8, as an aim of the invention is to avoid accelerating gas, it is important that for a given cross-section of duct (40) the ratio of stationary to moving duct (40) surfaces should be as high as possible. In embodiments using a channel (330) within the rotor (20), the duct (40) may be formed by a groove (580) which winds spirally down the circumferential face of the rotor (20) so that all parts of the duct (40) including the high pressure/high temperature end of the duct (40) are continuously exposed to fresh surface areas to conduct heat from the duct (40). In such an embodiment heat transfer up the cylinder stationary wall (10) may be reduced by flanges (350) behind the surface (see Fig 6). The rotor (20) may be further cooled by internal fluid flow along the sides of the duct (40) and side of rotor (20).

As shown in Fig. 8, a second class of embodiments employ a rotating rotor (20). In such a device there is provided a rotor (20) within a cylinder (10), the rotor (20) being profiled so that a substantial part of the rotor (20)

circumferential surface remains in rotatably close proximity to the inner wall of the cylinder (10) as it traverses the inner wall. The remaining circumferential surface of the rotor (20) is shaped or cut out so as to create a duct or groove (40) with a narrowing end (530) between it and the cylinder (10) wall. A wider end of the duct (540) is provided with an inlet conduit (520) communicating with the central part of an axial face of the rotor (20). Spaced from the duct (40) the circumferential surface of the rotor (20) is provided with an exit conduit (550) communicating with another portion of the axial face of the rotor (20). In large devices there may be provided more than one shaped duct (40).

In a spiral duct (500) embodiment the output pressure ratio may be controlled by providing a moveable sleeve (510) between the rotor (20) and cylinder (10). In operation, gas inlet (520) is through one axial end of the chamber (40) and outlet (550) through the other. Moving the sleeve (510) axially with respect to the rotor (20) changes the outlet area and so changes and controls the pressure ratio of the device.

Any of the above embodiments may be provided with means to adjust the offset of the rotor (20) from the central axis of the containing cylinder (10) and so adjust the clearance between the rotor (20) and cylinder (10) at the pinch point (50). This is advantageous for wear compensation, adjusting for different rates of thermal expansion, reducing leakage and to control capacity.

As shown in Fig. 9, an arrangement for driving the assembly, and additionally adjusting the offset of the rotor (20) from the central axis of the containing cylinder (10), and thereby adjusting the clearance between the rotor (20) and the cylinder (10) at the pinch point (50), is described herein. In overview, the rotor (20) has a rotor axis (670) each end of which is coupled to a drive rotor support and an idler rotor support (680, 690) respectively, each of the drive

rotor support and the idler rotor support (680,690) in turn are coupled to a drive shaft and an idler shaft (660, 650) respectively, which are arranged such that they are on the central axis of the cylinder (10) and are each supported by a bearing support (630).

In more detail, an end of the rotor axis (670) is joined by a coupling (600) to a drive rotor support (680), and an other end of the rotor axis (670) is joined by a coupling (600) to an idler rotor support (690). The idler rotor support (690) is joined by a coupling (600) to a fixed shaft (650). The drive rotor support (680) is joined by a coupling (600) to a drive shaft (660). Both drive shaft (660) and idler shaft (650) are arranged to be parallel to the rotor axis (670) and to lie on the central axis of the cylinder (10). Each rotor support (680, 690) is arranged to support the rotor axis (670) such that the rotor (20) surface is substantially positioned close to the inner circumference of the cylinder (10). Both idler shaft (650) and drive shaft (660) are supported by a bearing support (630) and are rotatable within, and axially constrained relative to said bearing support (630). Each bearing support (630) is arranged such that its axial distance from the centre of the rotor axis (670) is equal to that of the other bearing support (630) and is controllable. By controlling the distance of the bearing supports from the centre of the rotor axis (670) it is possible to vary the position and angle of each rotor support (680, 690) and resultantly it is possible to vary the running clearance between the rotor (20) and the housing (10).

Three classes of coupling (600) can be advantageously employed in the preceding arrangement. A first class of coupling (600) includes couplings which are suitable for forming a joint which is articulated in two axes between two shafts, but not capable of transmitting any axial torque. An example of a commonly known coupling (600) falling into the first class is a ball joint. A second class of coupling (600) includes couplings which are suitable for

forming a joint which is articulated in two axes between two shafts, and capable of transmitting axial torque. An example of a commonly known coupling (600) falling into the second class is a constant velocity joint, a Hardy-Spicer universal joint, certain types of rubber couplings or compliant rubber tubing. A third class of coupling includes couplings witch are suitable for forming a joint which is capable of articulating in one axis and capable of transmitting axial torque. An example of such a joint is a hinged joint.

The drive shaft (660) can transmit rotational torque via a drive coupling (640). The drive shaft (660) is coupled to the drive rotor support (680) by a coupling (600) of the third class. The end of the drive rotor support (680) which is coupled to the rotor axis (670) is thereby constrained to orbit in a circular motion around the draft shaft (660) axis.

In embodiments employing a rolling rotor, the drive rotor support (680) is coupled to the rotor axis (670) by a coupling (600) of the first class. The rotor axis (670) is coupled to the idler rotor support (690) by a coupling of the first or second class. In such embodiments, either at least one of the coupling (600) which couples the rotor axis (670) to the idler rotor support (690) and the coupling (600) which couples the idler rotor support (690) to the idler shaft (650) are of the first class, and/or the idler shaft (650) is free to rotate. The rotor (20) is thereby free to roll independently of the drive shaft (640) orientation and the idler shaft (650) orientation, but the rotor (20) is compelled to traverse the inner circumference of the cylinder (10) by the drive transmitted from the drive shaft (640) to the drive rotor support (680).

In embodiments employing an orbiting rotor, the drive rotor support (680) is coupled to the rotor axis (670) by a coupling (600) of the first class. In such embodiments, both of the coupling (600) which couples the rotor axis (670) to

the idler rotor support (690) and the coupling (600) which couples the idler rotor support (690) to the idler shaft (650) are of the second class, and the idler shaft (650) is fixed so that it cannot rotate. The rotor (20) is thereby constrained so as to maintain its orientation with respect to the cylinder (10) by virtue of its connection to the fixed idler shaft (650). The rotor (20) is compelled to traverse the inner circumference of the cylinder (10) by the drive transmitted from the drive shaft (640) to the drive rotor support (680).

In embodiments employing a rotating rotor, the drive rotor support (680) is coupled to the rotor axis (670) by a coupling (600) of the second or third class.

The rotor axis (670) is coupled to the idler rotor support (690) by a coupling of the first or second class. In such embodiments, either at least one of the coupling (600) which couples the rotor axis (670) to the idler rotor support

(690) and the coupling (600) which couples the idler rotor support (690) to the idler shaft (650) are of the first class, and/or the idler shaft (650) is free to rotate. The rotor (20) is thereby constrained to maintain its orientation with respect to the drive rotor support (680), and is unconstrained relative to the idler rotor support (690) orientation, and as a result, a fixed point on the rotor

(20) surface is maintained adjacent to the pinch point (50). The rotor (20) is compelled to traverse the inner circumference of the cylinder (10) by the drive transmitted from the drive shaft (640) to the drive rotor support (680).

Although the rolling, orbiting and rotating rotor constraint arrangements have been herein described with reference to the use of specific combinations of the aforementioned classes of coupling, it will be appreciated that the rotor characteristics described herein can be accomplished by other combinations not described. Accordingly, the descriptions of the orbiting, fixed, and rotating rotor constraint arrangements described herein are not intended to be limiting to the scope of the invention, the invention being set out in the claims.

As shown in Fig. 9, a means for counterbalancing the rotor (20) is provided. The drive rotor support (680) is extended past the coupling (600) which couples the drive rotor support (680) to the drive shaft (660), in a direction away from the rotor (20). A counterbalance weight (620) is provided either separately, or integrally with the drive rotor support (680) extension. Similarly, the idler rotor support (690) is extended past the coupling (600) which couples the idler rotor support (690) to the idler shaft (650), in a direction away from the rotor (20). A counterbalance weight (620) is provided either separately, or integrally with the idler rotor support (690) extension. Each counterbalance weight (620) is arranged to have a weight and a distance from the central axis of the cylinder (10) such that the weight of the rotating components on the opposite side of the central axis of the cylinder (10) is balanced. The mass or position of the counterbalance weights (620) can be adjusted during operation of the compressor, to compensate for thermal expansion or other effects which would otherwise upset the balance of the rotating components of the compressor. This can be achieved by the use of actuators to adjust the position of the counterbalance weights (620) on the rotor supports (680, 690), Alternatively, the mass of the counterbalance weights (620) can be altered, for example by pumping fluid or gas in or out of the counterbalance weights (620) which can incorporate a fluid or gas reservoir.

Fig. 10 shows an alternative arrangement for counterbalancing the rotor (20) where the drive shaft (660), rotor axis (670) and counterbalance weights (620) are housed within the cylinder (10), this being advantageous in that sealing of the chamber is facilitated.

Although the manner in which the various chambers are sealed and ducted are not described in all cases in detail it will be appreciated that in embodiments of

this invention the usual sliding seal means of the compressor art are provided to prevent leakage of gas from high pressure volumes to low pressure volumes. Ducting means to direct low pressure gas into devices and high pressure gas away from the device are also provided.

In any of the above embodiments conventional control means of the art, such as valves, may be used in combination to control and regulate flow.

Although embodiments have been described with a static cylinder (10) and a movable rotor (20), other embodiments may employ a moving cylinder (10) and static rotor (20) or both moving rotor (20) and cylinder (10).

A compressor constructed according to this invention may be reversed, with appropriate valving, to operate as an expander.

Advantages of the present invention are that high efficiency of compression and high stage pressure rise are achieved by compressing gas while imparting as little kinetic and friction energy to the gas. The invention also allows cooling of the gas while it is being compressed.

In axial and centrifugal compressors the necessity of multiple stages, caused by the low pressure rise per stage can be exploited to provide inter-cooling between stages. For high efficiency of compression all surfaces enclosing the gas may be cooled and the gas and/or surfaces continuously changed so that the gas is brought into contact with freshly cooled surfaces during compression. Preferably the gas should not flow relative to the walls as this causes frictional heating.

The invention has several advantages over prior art compressors. These include:

By employing supersonic rotation of the pinch point, the simple mechanical layout of the invention is made possible, since high pressure cannot propagate to the low pressure areas of the chamber and therefore no mechanical separation between low and high pressure regions of the chamber is required.

The continuous rotational compression means of the invention allows for smooth continuous compression. By employing smooth and continuous compression means, the invention advantageously reduces the energy imparted to the gas being compressed.

By employing adjustable running clearance means, and/or a rotor which rolls as it traverses the internal circumference of the cylinder, frictional losses are reduced, which reduces heating of the gas to be compressed and thereby increases efficiency.

The fixed chamber volume of the invention allows for enhanced heat transfer properties because the maximum chamber surface area is always in contact with the gas being compressed. This allows the gas being compressed to be more effectively cooled, which in turn aids compressor efficiency.

The amount of gas processed in each revolution is greater than the volume of the interior volume of the cylindrical chamber and the volume of the rotor. The swept volume is the cylinder volume less the volume of a rotor having a radius equal to: {radius of the rotor minus the radial offset of the rotor axis from the cylinder axis}. In other words, the sweep path of the rotor surface diametrically opposite the pinch point defines the swept volume.

A further advantage of the present invention is that it exhibits high flow properties compared to, for example, axial or centrifugal compressors of a similar physical size. As a result, the compressor of the present invention can be made physically smaller than known compressors.

Although the invention has been explained in relation to its preferred embodiments, these are not intended to limit the invention. It will be understood by those skilled in the art that many other modifications and variations are possible without departing from the scope of the invention as claimed. Embodiments and features of embodiments may be juxtaposed or interchanged as appropriate.