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Publication numberUS20050223734 A1
Publication typeApplication
Application numberUS 10/513,289
Publication dateOct 13, 2005
Filing dateApr 30, 2003
Priority dateMay 1, 2002
Also published asEP1502007A1, WO2003093649A1
Publication number10513289, 513289, US 2005/0223734 A1, US 2005/223734 A1, US 20050223734 A1, US 20050223734A1, US 2005223734 A1, US 2005223734A1, US-A1-20050223734, US-A1-2005223734, US2005/0223734A1, US2005/223734A1, US20050223734 A1, US20050223734A1, US2005223734 A1, US2005223734A1
InventorsIan Smith, Nikola Stosic, Ahmed Kovacevic
Original AssigneeSmith Ian K, Stosic Nikola R, Ahmed Kovacevic
Export CitationBiBTeX, EndNote, RefMan
External Links: USPTO, USPTO Assignment, Espacenet
Screw compressor-expander machine
US 20050223734 A1
Abstract
A plural screw compressor-expander machine has a casing (10) in the interior of which are mounted intermeshing helical rotors (11 and 12). The rotors (11 and 12) are supported at each end in bearings in the end walls of the casing. The interior of the casing is divided by a transverse partition (14) into a relatively longer compressor portion (1) and a shorter expander portion (5). The higher pressure ports (16, 18) of the compressor and expander portions are adjacent the partition, on opposite sides of a plane through the rotor axes. Similarly, the lower pressure ports (15, 17) are on opposite sides of the plane through the rotor axes but adjacent the end walls. This arrangement reduces the loads on the bearings of the rotors.
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Claims(10)
1. a plural-screw expander-compressor machine comprising intermeshing first and second helically-profiled rotors mounted for rotation in opposite directions in a housing by means of bearings at the opposite ends of the rotors, the interior of the housing being divided by a transverse partition into a compressor portion and an expander portion, the rotors extending through the partition and having circularly profiled portions where they extend through the partition, the two housing portions each having a higher pressure port and a lower pressure port, the higher pressure ports being adjacent the partition and on opposite sides of the rotors.
2. A machine according to claim 1, wherein the compressor portions of the rotors have the same profile as the expander portions thereof.
3. A machine according to claim 1, wherein helical lobes of the first rotor make sealing contact on both sides with the second rotor.
4. A machine according to claim 1 wherein the partition extends into an annular groove of each rotor.
5. A refrigeration system comprising an evaporator, a cooler or condenser and a machine according to any preceding claim, the compressor portion being connected to deliver refrigerant from the evaporator under increased pressure to the cooler or condenser and the expander portion being connected to receive refrigerant from the cooler or condenser and to deliver it at reduced pressure to the evaporator.
6. A system according to claim 5 and including a mechanical power connection to drive the machine or take power therefrom.
7. (canceled)
8. A plural-screw compressor-expander machine comprising intermeshing first and second helically-profiled rotors mounted for rotation in opposite directions in a housing, the housing defining an expansion chamber and a compressor chamber, the rotors extending through both chambers, the chambers being separated and each having an inlet and an outlet, one of the rotors having a connection for supplying or taking off mechanical power.
9. A fuel-cell installation comprising at least one fuel cell, a supply of pressurised fuel for the fuel-cell, a plural-screw expander-compressor machine comprising intermeshing first and second helically-profiled rotors mounted for rotation in opposite directions in a housing by means of bearings at the opposite ends of the rotors, the interior of the housing being divided by a transverse partition into a compressor portion and an expander portion, the rotors extending through the partition and having circularly profiled portions where they extend through the partition, the two housing portions each having a higher pressure port and a lower pressure port, the compressor portion of the machine being connected to deliver compressed air to the fuel-cell or cells and the expander portion of the machine being connected to receive pressurised exhaust from the fuel-cell or cells.
10. A fuel-cell installation comprising at least one fuel cell, a plural-screw expander-compressor machine comprising intermeshing first and second helically-profiled rotors mounted for rotation in opposite directions in a housing by means of bearings at the opposite ends of the rotors, the interior of the housing being divided by a transverse partition into a compressor portion and an expander portion, the rotors extending through the partition and having circularly profiled portions where they extend through the partition, the two housing portions each having a higher pressure port and a lower pressure port, the higher pressure ports being adjacent the partition and on opposite sides of the rotors, the compressor portion of the machine for supplying air under pressure to the fuel-cell or cells for expanding exhaust gas from the fuel-cell installation.
Description

This invention relates to plural-screw compressor-expander machines. These are positive displacement rotary machines which consist, essentially, of a pair of meshing helically lobed rotors, contained in a casing.

Plural-screw machines are widely used as compressors. An important feature of such machines is that if the direction of gas flow is reversed, so that high-pressure gas is delivered to flow into the machine through the high pressure port and out through the low pressure port, it will act as an expander with the direction of rotation reversed. The machine will also work as an expander when rotating in the same direction as a compressor provided that the suction and discharge ports are positioned on the opposite sides of the casing to those for a compressor since this is effectively the same as reversing the direction of rotation relative to the ports. When operating as a compressor, mechanical power must be supplied to a main rotor to rotate the machine. When acting as an expander, the rotor will rotate automatically and generate power.

A major problem with the plural screw machines is that the pressure difference between entry and exit creates very large radial and axial forces on the rotors whose magnitude and direction is independent of the direction of rotation. It is normal practice to have bearings on each end of the rotors and these have to withstand both the radial and axial loads induced by the pressure difference. As a result, some of the power transmitted through the rotors is lost in bearing friction. More importantly, in these machines, the pressure difference by which it is possible to compress gases within one pair of rotors is limited to approximately 60 bar in normal designs. This is because the bearing sizes, needed to withstand large forces, are very big and the space between the rotor shafts, in which they must fit, is limited by the required distance between the axes of the rotors. Thus any means of reducing these bearing loads will extend the range of pressures and hence applications, for which such machines may be used.

In recent years, there have been proposals to replace the conventional throttle valve of a refrigeration plant by an expander, which is typically a plural-screw machine, and to use the power thereby generated to assist the main vapour compressor or to drive a separate vapour compressor. (See Smith, I. K. and Stosic, N. R. ,“The Expressor: An Efficiency Boost to Vapour Compression Systems by Power Recovery from the Throttling Process”. AES-Vol. 34, Heat Pump and Refrigeration Systems Design, Analysis and Applications ASME 1995. U.S. Pat. Nos. 5,192,199; 5,911 743; and 6,185,956 and International Patent Specification WO00/7558). This results in an expander-compressor machine of the kind for which the present invention relates.

According to the present invention, there is provided a plural-screw expander-compressor machine comprising intermeshing first and second helically-profiled rotors mounted for rotation in opposite directions in a housing by means of bearings at the opposite ends of the rotors, the interior of the housing being divided by a transverse, partition into a compressor portion and an expander portion, the rotors extending through the partition and having circularly profiled portions where they extend through the partition, the two housing portions each having a higher pressure port and a lower pressure port, the higher pressure ports being adjacent the partition and on opposite sides of the rotors.

By arranging the high pressure ports near each other but on opposite sides of the rotors, the net lateral and axial forces on the rotors are greatly reduced resulting in an important reduction in the rotor bearing sizes.

Also during recent years, there has been much interest in the use of what are described as “natural” working fluids in refrigeration and air conditioning systems in order to avoid the use of halogenated hydrocarbons and therefore minimise atmospheric pollution. One of the most favoured of these is carbon dioxide (CO2) . However, despite its environmental advantages, two major problems inhibit its widespread use. The first of these is that the pressure differences required across the compressor are of the order of ten times those needed using existing refrigerants. This implies that if screw compressors are used, the very heavy bearing loads associated with these pressures make the design very complex. The second limitation is that when used in normal air conditioning and refrigeration systems, the range of operating pressures and temperatures required are close to the critical point of CO2. Hence the losses associated with throttling are much larger than those associated with the use of conventional refrigerants. It follows that some recovery of power in the expansion process is therefore required in order to achieve an acceptable coefficient of performance from a CO2, cycle.

The use of an expander-compressor in accordance with the invention mitigates the problem of high bearing loads associated with twin screw machines, and at the same time enables some power to be recovered from the expansion of the fluid between the cooler and evaporator for example in a CO2vapour compression cycle system.

The invention will now be further described by way of example with reference to the drawings, in which:-

FIG. 1 is a schematic circuit diagram of a refrigeration system with carbon dioxide as refrigerant and incorporating a machine embodying the invention,

FIG. 2 is a longitudinal sectional view of the machine on the axes of the two rotors,

FIG. 3 is a longitudinal sectional view through the axis of the main rotor at right angles to FIG. 2,

FIG. 4 shows the forces acting on the compressor-forming portions of the rotors,

FIG. 5 shows the forces acting on the expander-forming portions of the rotors,

FIG. 6 shows the net forces acting on the rotors,

FIG. 7 is an enthalpy entropy diagram of the system shown in FIG. 1,

FIG. 8 is a schematic diagram of a fuel cell system incorporating a machine embodying the invention, and

FIGS. 9 and 10 are views similar to FIGS. 2 and 3 of an alternative machine suitable for use in the system of FIG. 8.

FIG. 1 shows the layout of a CO2 refrigeration system, operating between an evaporating temperature of 0° C. and a cooler exit temperature of 40° C. As in a conventional system, CO2at approximately 35 bar has its pressure raised to 100 bar in a compressor 1 driven by a motor 2. It then passes through a cooler 3 where it is cooled in the supercritical state at approximately constant pressure until it reaches a temperature of 40° C. The cooled dense fluid would then pass in conventional practice through a throttle valve in which the pressure is reduced back to 35 bar. As a result of the pressure drop, it liquefies and part flashes into vapour, causing the liquid-vapour mixture to fall in temperature to 0° C. The cooled liquid CO2, together with the vapour formed during flashing, then passes through an evaporator 4, where it receives heat from the cold surroundings at approximately 35 bar and 0° C. until all the refrigerant is evaporated. The dry, or slightly superheated vapour then enters the compressor 1 to complete the cycle. As can be seen, the required pressure rise across the compressor is 65.2 bar, which is beyond the limit of what is readily achievable in a single stage twin screw compressor; due to excessive loads on the rotor bearings. Further, the energy losses due to the throttle valve would be substantial.

To overcome these problems, the compressor rotors are extended to form expander rotor portions in an expander 5. The resulting machine is shown schematically in FIGS. 2 and 3 and includes a housing 10 defining a chamber containing a helically lobed main rotor 11 and a helically grooved gate rotor 12 which meshes with the main rotor 11. Each rotor has a cylindrical extension at each end by means of which it is rotatably supported in bearings (not shown) in the end walls of the housing 10, the extension at one end of the main rotor 11 being prolonged at 13 for a driving connection to the motor 2.

The interior of the chamber in the housing is divided by a transverse partition 14 into a longer compressor portion and a shorter expander portion. The partition 14 is divided along a plane through the axes of the rotors and extends into an annular groove in each rotor 11, 12. During assembly, the two halves of the partition are engaged in the rotors and the assembly thus formed is introduced into the chamber through one end thereof.

The compressor portion of the housing has a large diameter (and thus large area) inlet port 15 at one end of the housing (its position relative to the rotors being indicated in FIG. 2) and a smaller diameter (and thus small area) outlet port 16 adjacent the partition 14, on the opposite side of the rotors. The compressor inlet port 15 is connected by a line 21 (FIG. 1) to the outlets of the evaporator and the compressor outlet port 16 is connected to the inlet of the cooler 3 by a line 22.

The expander portion of the housing has a larger diameter (and thus large area) outlet port 17, at the opposite end of the housing 10 to the compressor inlet port 15, and a smaller area inlet port 18 adjacent the partition 14 on the opposite side of the rotors 11 and 12 to the outlet port 17. The expander outlet port 17 is connected by a line 24 to the inlet of the evaporator 4 and the expander inlet port 18 is connected by a line 23 to the outlet of the cooler 3.

The ports 16 and 18 are the high pressure ports of the compressor and expander. They are on opposite sides of the rotors (FIG. 3) but axially close to each other, adjacent the partition 14.

Referring to FIG. 3, high pressure dense fluid enters the expander port 18 at the top of the casing 10, near the centre, and leaves through the low pressure port 17 at the bottom of the casing at one end, as a mixture of liquid and vapour. The expansion process causes the temperature to drop, as in passing through a throttle valve. However, here the fall in pressure is used to recover power and causes or assists the rotors to turn. Vapour from the evaporator 4 enters the low pressure compressor inlet port 15, at the top of the opposite end of the casing, is compressed within it and expelled from the high pressure discharge port 16 at the bottom of the casing, near the centre, to be delivered to the cooler 3. Ideally, there is no internal transfer of fluid within the machine between the expansion and compression sections which each take place in separate chambers.

Because the high pressure ports are in the centre of the unit and arranged so that they are on opposite sides of the casing, the high pressure forces due to compression and expansion are opposed to each other and, more significantly, only displaced axially from each other by a relatively short distance. The radial forces on the bearings are thereby significantly reduced. In addition, since both ends of the rotors are at more or less equal pressure, the axial forces virtually balance out. The following example indicates the extent of the advantages, which are possible from this arrangement.

Consider a refrigerator in which CO2 leaves the evaporator at the rate of 2.75 m3/min as dry saturated vapour at a pressure of 35 bar to leave the compressor and enter the cooler at a discharge pressure of 100 bar.

To determine the reduction in mechanical forces, exact calculations were carried out on a large simulation program to aid the design of twin screw machines. The results of this showed that for the compressor, the main rotor required would be 102 mm in diameter with a length:diameter ratio of 1.5:1. The expander required to replace a throttle valve in this system would have a male rotor of the same diameter but with a length:diameter ratio of 1:1.

FIG. 4 shows the compressor rotors portions 11C, 12C and the bearing loads which must be resisted if the refrigeration system is designed with a conventional screw compressor drive. On the main rotor 11C alone, there is an axial force of 92 kN and radial bearing forces of 133 kN at the high pressure end and 45.5 kN at the suction end. These forces are at angles to the plane of the figure.

In FIG. 5 the expander rotor portions 11E, 12E and their corresponding bearing forces are similarly shown. Here, the axial bearing load on the main rotor is 92 kN while the corresponding radial loads are 86 kN at the high pressure end and 34 kN at the low pressure end.

FIG. 6 shows the bearing forces as a result of use of the invention if the compressor and expander rotors are machined on the same shafts with the high pressure ports in the middle and the low pressure ports at each end of the housing. By this means the main rotor axial load has been reduced to 0 kN. The radial bearing loads (added vectorially) are now 117 kN at the expander end and 101 kN at the compressor end.

Thus the total bearing load on the main or male rotor 11 alone has been reduced from 270.4 kN for the compressor alone to 218 kN for the combined compressor-expander. If both male and female rotors are included, then the total bearing load is reduced from 556 kN for the compressor alone to only 448 kN for the combined rotors. This amounts to a total decrease in bearing load of nearly 20%. Design problems associated with high bearing loads in screw compressors for CO2systems are thereby reduced.

With regard to the improvement in thermodynamic performance, an enthalpy entropy diagram of the idealised cycle with reversible compression and expansion of the CO2 is shown in FIG. 7.

Additionally in FIG. 7, the curve 31 is the saturation line for CO2 and the curve 32 is the saturation line for CO2 vapour.

As can be seen, point 21 corresponds to vapour being admitted to the compressor through the line 21 of FIG. 1, point 22 to discharge from the compressor 1 at 22 and entry to the cooler 3 and point 23 to exit from the cooler 3. If the fluid then passes through a throttle-valve, isenthalpic expansion will lead to it entering the evaporator at point 24 t. However, if it passed through the expander and work is extracted from it, then the expansion process will be adiabatic and the fluid will enter the evaporator at point 24 e. The difference between these two processes is that work extraction reduces the specific enthalpy of the fluid entering the evaporator by 14.9 kJ/kg. This causes the same mass of fluid to enter the evaporator with less vapour and hence has the effect of increasing the refrigerating capacity of the plant by 12.4%.

At the same time, this recovery of 14.9 kJ/kg in the form of shaft work is used to reduce the external work input to the compressor, shown by the difference between points 21 and 22, from 43.0 kJ/kg to only 28.1 kJ/kg. Thus there is a saving in power input of 34.6%.

The coefficient of performance will be improved by both these factors and thus be theoretically increased by 72%. However, these figures are based on idealised work input and output. In a practical system, allowance would have to be made for the compression and expansion efficiencies, which would reduce the expansion work and increase the compression work. Nonetheless, an overall gain in coefficient of performance of the order of 30% should still be achievable by this means.

A further preferred feature is the use of rotors which seal on both contacting surfaces so that the same profile may be used both for the expander and the compressor sections. In fact, since compression and expansion are carried out separately, the compressor and expander profiles could be different. However, this would make manufacture extremely difficult, due to the very small clearance space, which could be less than 10 mm, between the, two rotor portions. By using the same profile for both, the compressor and expander rotors can be machined or ground in a single cutting operation and then separated by machining a parting groove in them for the partition on completion of the lobe formation.

Additionally, the expansion section can contain a capacity control such as a slide or lifting valve to alter the volume passing through it at part load, in a manner identical to capacity controls normally used in screw compressors. This would be in addition to any capacity or volume ratio control used for the compression section. This would then replace the throttle valve control system normally required in conventional vapour compression systems.

Although the invention is especially suitable for operation on high pressure CO2 systems, it may equally be used with more conventional refrigerants, or indeed, wherever there is a need for combined expansion and compression processes or even if a combined expansion-compression process is established only to reduce the rotor loads.

The balanced rotor concept is also applicable for the “expressor” system of a motorless self-driven expander-compressor machine described in the paper ‘Expressor’ mentioned above.

Another application of such a machine is to the supply of air under superatmospheric pressure to a fuel-cell installation, the system being powered by exhaust gas from the fuel-cell installation as it expands through the expander portion of the machine. FIG. 8 is a block diagram of a fuel cell system using hydrogen as fuel and incorporating a machine as shown in FIGS. 2 and 3. Hydrogen is supplied from a source 41, such as a hydrogen generator or a pressurised tank, through a pressure regulator 42 to a fuel cell stack 43. Unused hydrogen from the stack is recirculated at 44. To provide the required oxygen, air is drawn in from an intake 45 and intake filter 46 via the compressor portion 1 of the machine shown in FIGS. 2 and 3. The air enters the compressor 1 through its port 15 and is delivered to the fuel cell stack 43 through the high pressure port 16. Combustion products from the fuel cell stack 43 under pressure are delivered to the inlet port 18 of the expander portion 5 and leave the latter through its outlet port 17, such exhaust consisting of water, and nitrogen.

Excess heat generated in the fuel cell stack 43 is removed by a cooling system including a radiator 47 and a coolant circulating pump 48 driven by electricity generated within the fuel cell stack.

The main electrical power output from the fuel cell stack is delivered to a power distribution unit 49 which distributes power to the driving motor 50 for the compressor expander machine, a DC converter 51 for charging a storage battery 52 and a traction motor assembly 53 for driving a vehicle axis 54 in the case of a vehicle.

FIGS. 9 and 10 correspond to FIGS. 2 and 3 and show an alternative form of machine which may in some cases be used. in the fuel cell system shown in FIG. 8. In FIGS. 9 and 10, parts corresponding to those of FIGS. 2 and 3 have the corresponding reference numerals increased by 100. It will be noted that the large area low pressure ports 115 and 117 are adjacent the partition 114 and the small area high pressure ports 116 and 118 are at opposite ends of the machine.

Referenced by
Citing PatentFiling datePublication dateApplicantTitle
US7530217Dec 16, 2005May 12, 2009General Electric CompanyAxial flow positive displacement gas generator with combustion extending into an expansion section
US7726115Feb 2, 2006Jun 1, 2010General Electric CompanyAxial flow positive displacement worm compressor
US7784303 *Aug 5, 2005Aug 31, 2010Daikin Industries, Ltd.Expander
US8096288Oct 7, 2008Jan 17, 2012Eaton CorporationHigh efficiency supercharger outlet
US20100071458 *Nov 30, 2009Mar 25, 2010General Electric CompanyPositive displacement flow measurement device
US20120090349 *Sep 21, 2011Apr 19, 2012Kabushiki Kaisha Kobe Seiko Sho (Kobe Steel, Ltd.)Refrigerator
CN102003214A *Dec 14, 2010Apr 6, 2011范年宝Novel screw expanding power machine
Classifications
U.S. Classification62/402, 417/423.1
International ClassificationF01C11/00, F01C21/00, F01C1/16, F04C18/08
Cooperative ClassificationF01C1/16, F01C21/003, F01C21/108, F04C18/084, F01C11/004
European ClassificationF01C1/16, F01C21/10D4, F01C11/00B2, F01C21/00C
Legal Events
DateCodeEventDescription
May 6, 2005ASAssignment
Owner name: CITY UNIVERSITY, UNITED KINGDOM
Free format text: ASSIGNMENT OF ASSIGNORS INTEREST;ASSIGNORS:SMITH, IAN KENNETH;STOSIC, NIKOLA RUDI;KOVACEVIC, AHMED;REEL/FRAME:015978/0803
Effective date: 20041111