Patent Grant
10400708
The present invention relates generally to the field of Stirling-cycle machines and more specifically to Stirling engines, —coolers or—heat pumps. In particular, the present invention relates to pistonless Stirling-cycle machines utilising rotary expander- and compressor mechanisms.
It is commonly known that the Stirling cycle is a thermodynamic cycle that includes, inter alia, the cyclic compression and expansion of air or other gas (i.e. a working fluid) at different temperatures, such that there is a net conversion of thermal energy to mechanical work. It is also known that the cycle is reversible, which means that, if supplied with mechanical power, the apparatus can function as a heat pump or cooling machine for respective heating or cooling, and even for cryogenic cooling.
More specifically, the Stirling cycle is a closed regenerative cycle utilizing, in general, permanently gaseous working fluid. Here, “closed-cycle” means that the working fluid is permanently contained within the thermodynamic system, and the term “regenerative” refers to the use of an internal heat exchanger, also called a regenerator. The regenerator increases the device's thermal efficiency by recycling internal heat that would otherwise pass through the system irreversibly. The Stirling cycle, like many other thermodynamic cycles, comprises the four main processes of (i) compression, (ii) heat addition, (iii) expansion, and (iv) heat removal. However, in real engines these processes are not discrete, but rather such that they overlap.
An example of a typical Stirling engine 10 with a crank-drive mechanism is shown in
A regenerator 18 is introduced between the heater 16 and the cooler 20 to prevent heat losses that would otherwise occur if the heater 16 and cooler 20 were in direct contact. The regenerator 18 in this example comprises a porous medium that is enclosed in a metallic casing. This porous medium is made from a material with a high heat capacity and should ideally have infinite radial- and zero axial thermal conductance. The porous medium can be understood to act as a heat sponge, where heat is transferred to the material of the regenerator and stored when the working fluid flows from the “hot” zone to the “cold” zone. When the working fluid flows in the opposite direction, the stored heat is returned from the regenerator to the working fluid. Thermo-insulation is usually used to separate the porous medium from the walls of its casing in order to further reduce heat losses.
To provide for most of the working fluid to be in the hot zones (i.e. hot cylinder 12 and heater 16) during the heat input phase, and for most of the working fluid to be in the cold zone (i.e. cold cylinder 14 and the cooler 20) during the heat rejection phase, the piston 22 in the hot cylinder 12 is leading the piston 24 of the cold cylinder 14 by usually 90° to 110° (degrees of crankshaft angle) in the displacement, so the volume of the hot cylinder 12 leads the volume of the cold cylinder 14 in its variation by 90° to 120° degrees.
The two variable volumes (hot and cold) that are connected by a set of heat exchangers (heater 16, regenerator 18 and cooler 20), the variation of volume in the hot space which is leading the variation of volume in the cold space by 90° to 110° (degrees), and the reciprocating flow of the working gas between the variable hot space and cold space through channels of a set of heat exchangers 16, 18, 20, are characterising features of Stirling cycle machines. Typical PV-diagrams for the variable hot or expansion volume (dashed line) and the cold or compression variable volume (solid line) are shown in
Therefore, if the heater 16 is exposed to a relatively high temperature environment and the cooler 20 is exposed to a relatively low temperature environment, then the machine works as an engine that exerts power (i.e. the hot or expansion space area is greater than the cold or compression space area in the PV diagram, see
However, if the cooler 20 is exposed to a relatively low temperature environment and the pistons are driven using an electric motor (e.g. via a shaft) or any other actuation sources, then the temperature of the working fluid in the heat exchanger 16 and variable expansion space 12 will reduce significantly (e.g. down to cryogenic levels), so that the machine operates as a cooling device generating cold (i.e. the expansion space area is less than the compressions space area in the PV diagram).
Alternatively, if the heat exchanger 16 is exposed to the relatively low temperature environment and the pistons are driven using an electric motor (e.g. via a shaft) or using any other actuation sources, then the temperature of the heat rejection in the cooler 20 will be significantly higher than the temperature of the heat exchanger 16, and the machine is working as a heat pump (i.e. absorbing heat at low temperature and delivering it ah high temperature).
The cycle of conventional Stirling machines with reciprocating motion of pistons in cylinders is usually completed after 360 degrees of the shaft angle.
However, conventional Stirling machines with reciprocating piston motion in cylinders (kinematical drive engines or free piston reciprocating machines) come with considerable disadvantages, such as, for example:
In order to reduce the size and weight of these machines, designers may separate the crank-case from the gas circuit of the engine using a “sealing” of the vertical rod connecting pistons and drive mechanism (i.e. a so called unpressurised crank-case).
Such a sealing has been achieved only on a very limited number of Stirling machines, and even in those engines the working fluid in the internal gas circuit has to be replenished repeatedly, since it is not possible to fully eliminate working fluid leakages in a rod sealing.
Furthermore, in free piston machines there is no conventional drive mechanism and pistons are driven reciprocally utilising the gas forces provided in the internal gas circuit of machines and mechanical springs. The oscillating motion of the cold piston may be converted into electrical power by attaching rare-earth magnets to the piston and these magnets are surrounded by copper coils (i.e. the concept of linear generator). Such machines do not have a large crankcase and the engine is fully sealed by placing the linear alternator inside the engine casing. Its specific weight and dimensions are significantly improved than those of conventional kinematical machines, but so far the power output is limited to about 3 to 10 kW (Kilowatts), which is considerably lower than the output of conventional kinematical engines. The frequency of oscillation of the pistons corresponds to a rotational speed of the shafts between 2000 and 4000 RPM (revolutions per minute).
The solution to the problems of reciprocating piston machines is believed to be in rotatory machines. Thus considerable effort has gone into the development of rotary Stirling engines/machines.
For example, prior art document U.S. Ser. No. 13/795,632 describes a rotary Stirling cycle engine using “hot” and “cold” gerotor sets that are mounted on the same shaft and which are separated by an insulation barrier. The barrier provides a regenerative gas passage allowing gases to flow through, therefore, connecting the displacing chambers of the “hot” and “cold” gerotor sets. The gerotor Stirling-cycle engine may be used for generating electricity or mechanical power.
Prior art document U.S. Ser. No. 05/790,904 discloses another example of a Stirling-cycle machine having a rotary mechanism. In this particular design, a rotary vane expander and a rotary vane compressor are mounted on the same shaft, wherein each vane unit forms four working volumes. Corresponding working volumes of the expander and the compressor are connected via a set of heat exchangers that are provided in the casing of the expander and in the shaft.
All of these prior art examples have the same essential features of Stirling-cycle machines, i.e. a harmonic or near harmonic variation of continuously connected corresponding working spaces in the expander and the compressor units. Thus, once respective chambers are connected through a set of heat exchangers, the working gas flows between corresponding working spaces in a reciprocating motion. However, it can be understood by the person skilled in the art that the described rotary mechanisms are very complex and come with disadvantages of their own.
Consequently, rotary mechanisms such as twin-screw or scroll mechanisms were considered for use in Stirling-cycle machines. Twin-screw mechanisms in particular have been a very popular choice for compressors.
As shown in
Another example of a rotary mechanism used for compression or expansion of gases is a conical screw rotary compressor 50 (for example as manufactured by VERT Rotors Ltd), as shown in
However, the cyclic volume changes provided by the twin-screw, scroll or conical screw rotary mechanisms follow a linear or nonlinear saw-tooth function as shown in
Yet, the saw-tooth character of the working fluid volume variation, as provided by these rotary machines, has made twin-screw, scroll or conical screw mechanisms unsuitable for use in the Stirling cycle.
Currently available thermodynamic apparatuses that utilise twin-screw or scroll mechanism either are applied in the Rankine or the Joule/Bryton cycle, each of which requires an axial flow of the working fluid in one direction only. For example prior art documents DE10123 078 or AT412663 describe thermodynamic cycles utilising twin-screw expanders.
In particular, DE10123078 discloses a machine that operates on a closed thermodynamic cycle where the high-pressure gas is supplied into and expanded by a twin-screw mechanism. The work generated by the gas expansion is converted into useful mechanical work through the rotating twin-screw shafts, before the working fluid is then re-heated (and re-pressurised) and directed back to the twin-screw mechanism, where the cycle is repeated.
Another example of a rotary thermodynamic engine (now utilising a scroll mechanisms), is disclosed in a publication by Youngmin Kim, Dongkil Shin, Janghee Lee and Kwenha Park (“Noble Stirling engine employing scroll mechanism”, Proceedings of the 11th International Stirling Engine Conference, 19-21 Sep. 2004, pp. 67-75), but a simple analysis reveals that the so called Stirling engine actually operates on the closed Joule/Bryton cycle, because the gas flow is circled around in one direction and not in a reciprocating motion.
Accordingly, it is an object of the present invention to provide a Stirling-cycle apparatus that is adapted to utilise rotary expander and compressor mechanisms, such as twin-screw, scroll or conical screw mechanism, even if the provided working fluid volume changes are described by linear or non-linear saw-tooth waveforms. Furthermore, it is a particular object of the present invention to provide a rotary Stirling-cycle cooler that can be made smaller than currently available Stirling-cycle cooler, and which has an improved efficiency.
Preferred embodiment(s) of the invention seek to overcome one or more of the above disadvantages of the prior art.
According to a first embodiment of the invention there is provided a Stirling-cycle apparatus comprising:
a hermetically sealable housing;
a first rotary displacement unit in fluid communication with a second rotary fluid displacement unit, each operably mounted in a separate, fluidly sealed portion within said housing and adapted to provide a cyclic change of at least one thermodynamic state parameter of a working fluid during use, each said first and second rotary displacement unit comprising:
a compressor mechanism, having a first compressor working chamber that is adapted to receive a first portion of said working fluid, and at least a second compressor working chamber that is adapted to receive a second portion of said working fluid, said first compressor working chamber comprises a first outlet port and said second compressor working chamber comprises a second outlet port;
an expander mechanism, having a first expander working chamber that is adapted to receive said first portion of said working fluid, and at least a second expander working chamber that is adapted to receive said second portion of said working fluid, said first expander working chamber comprises a first inlet port and said second expander working chamber comprises a second inlet port;
a drive coupling assembly, adapted to operably and operatively couple said first expander mechanism to said first compressor mechanism, comprising:
a rotating valve mechanism, adapted to provide a predetermined sequence of a cyclic fluid exchange between said first compressor working chamber and said first expander working chamber, and between said second compressor working chamber and said second expander working chamber, at predetermined intervals of the angle of rotation of said first and second rotatory displacement unit;
an actuator, operably coupled to said first and second rotary displacement unit, and adapted to synchronously link the rotational movement of said first rotary displacement unit with said second rotary displacement unit, such that said first predetermined cyclic change of at least one thermodynamic state parameter of said working fluid is offset in relation to said second predetermined cyclic change of at least one thermodynamic state parameter of said working fluid by a predetermined phase angle, during use.
The apparatus of the present invention provides the advantage that linear or non-linear “saw-tooth like” cyclic changes of at least one thermodynamic state parameter (i.e. volume) of the corresponding rotary compressor and expander mechanisms of the two rotary displacement units are paired and combined in such a way to provide a total variation of working space volumes that follows a periodic near-harmonic function that is typical for conventional Stirling cycle machines (e.g. piston motion), therefore providing a genuine rotary Stirling-cycle apparatus that is simpler in construction and which has an improved efficiency and performance, especially when provided in miniaturised form. The apparatus of the present invention can be operated so as to provide mechanical work, but also in reverse as a cooler or heat pump.
Advantageously, said first drive coupling assembly may further comprise at least one first drive shaft and at least one first shaft casing having an inner wall and which is configured to operably enclose said at least one first drive shaft.
Advantageously, said at least one first shaft casing may comprise a plurality of axially-spaced and partially circumferential first fluid channels provided at respective predetermined first axial positions extending over a first circumferential segment of said inner wall, and a plurality of axially-spaced and partially circumferential second fluid channels, provided at respective predetermined second axial positions extending over a second circumferential segment of said inner wall, and wherein said first circumferential segment is provided radially opposite said second circumferential segment, and wherein each one of said first axial positions is axially offset from each one of said second axial positions.
Preferably, each one said plurality of axially-spaced and partially circumferential first and second fluid channels may subtend an angle greater than 180 degrees.
Advantageously, said at least one drive shaft may comprise a first set of two corresponding conduits, a first conduit having a first opening fluidly coupled to said first outlet port and a second conduit having a first opening fluidly coupled to said first inlet port, each one of said corresponding said first and second conduits has two conjoined axially adjacent second openings exiting radially out of said drive shaft at a first predetermined radial angle, wherein a first one of said two conjoined axially adjacent second openings is adapted to fluidly engage with one of said plurality of first fluid channels, and a second one of said two conjoined axially adjacent second openings is adapted to fluidly engage with one of said plurality of second fluid channels.
Even more advantageously, said at least one drive shaft may comprise at least a second set of two corresponding conduits, a first conduit having a first opening fluidly coupled to said second outlet port and a second conduit having a first opening fluidly coupled to said second inlet port, each one of said corresponding said first and second conduits has two conjoined axially adjacent second openings exiting radially out of said drive shaft at a second predetermined radial angle, wherein a first one of said two conjoined axially adjacent second openings is adapted to fluidly engage with one of said plurality of first fluid channels, and a second one of said two conjoined axially adjacent second openings is adapted to fluidly engage with one of said plurality of second fluid channels.
Even more advantageously, each one of said plurality of first fluid channels may be fluidly coupled to a corresponding one of said plurality of second fluid channels, so as to allow a predetermined sequence of fluid exchange between said first compressor working chamber and said first expander working chamber, and between said second compressor working chamber and said second expander working chamber, during use.
Advantageously, a first and a second working space may be formed for each one of fluidly coupled said first compressor working chamber and said first expander working chamber, and fluidly coupled said second compressor working chamber and said second expander working chamber, in said first rotary displacement unit.
Advantageously, a first and a second working space may be formed for each one of fluidly coupled said first compressor working chamber and said first expander working chamber, and fluidly coupled said second compressor working chamber and said second expander working chamber, in said second rotary displacement unit.
Advantageously, each one of said first and second working space of said first rotary displacement unit may be in fluid communication with a corresponding one of said first and second working space of said second rotary displacement unit.
Preferably, each one of said corresponding fluidly coupled first and second fluid channels of said first rotary displacement unit may be in fluid communication with a respective one of each one of said corresponding fluidly coupled first and second fluid channels of said second rotary displacement unit.
Advantageously, each fluid communication between each one of said corresponding fluidly coupled first and second fluid channels of said first rotary displacement unit and each one of said corresponding fluidly coupled first and second fluid channels of said second rotary displacement unit may comprise any one or any serial combination of a first heat exchanger, a regenerator and a second heat exchanger.
Preferably, said first heat exchanger may be adapted to provide heat to said working fluid, and wherein said second heat exchanger may be adapted to remove heat from said working fluid. This provides the advantage that the apparatus can be operated in different modes, for example, as a cooler or as a heat pump depending on where the first and second heat exchangers are located in combination with the regenerator.
Even more preferably, said regenerator may be fluidly coupled between said first and second heat exchanger.
Alternatively, said first heat exchanger is an integral part of said first rotary displacement unit and/or said second heat exchanger is an integral part of said second rotary displacement unit.
Preferably, each one of said first and second rotary displacement unit may comprise a twin-screw mechanism.
Alternatively, each one said first and second rotary displacement units may comprise a scroll mechanism or a rotary conical screw mechanism.
In another alternative embodiment, each one of said first and second displacement unit may comprise any one of a twin-screw mechanism, scroll mechanism or a rotary conical screw mechanism.
Advantageously, said actuator may comprise a motor and a transmission adapted to synchronously drive said first and second rotary displacement units.
Alternatively, said actuator may comprise a motor and a transmission adapted to be powered by any one of said first and second rotary displacement units.
Advantageously, each one of said compressor and expander mechanism of said first rotary displacement unit, and each one of said compressor and expander mechanism of said second rotary displacement unit, may be provided in a discrete and hermetically sealed portion of said housing.
Preferably, said first rotary displacement unit may be a compression unit, and wherein said second rotary displacement unit may be an expansion unit.
Alternatively, the first rotary displacement unit may be an expansion unit and the second rotary displacement unit may be a compression unit, depending on the application of the apparatus, i.e. heat pump, cooler or engine.
Preferred embodiments of the present invention will now be described, by way of example only and not in any limitative sense, with reference to the accompanying drawings, in which:
The exemplary embodiments of this invention will be described in relation to a rotary Stirling-cycle cooler. However, it should be appreciated that, in general, the rotary Stirling-cycle apparatus of this invention will work equally well in a Stirling engine mode (i.e. output of mechanical work) or heat-pump (output of heat).
In addition, meshing male and female screw rotors may be provided with different ratios for the number of lobes. Theoretically, the ratio may start at ‘1’ (i.e. ‘2/2’), but in practice other (e.g. greater) ratios may be used. Typical examples of ratios used in practice may be ‘3/4’, ‘3/5’, ‘4/6’, ‘5/7’, ‘6/8’ etc. Also, the screw lobes may have a symmetric or asymmetric profile. For the sole purpose of illustrating the basic principle of the invention, the example embodiment comprises the more simplistic symmetrically profiled screw rotors with ‘2/2’ ratio lobes (i.e. the ratio is equal to ‘1’). Also, it is understood by the person skilled in the art that optimal performance may only be achieved utilising any other (i.e. more suitable) ratio and/or lobe profile (i.e. asymmetric or symmetric). However, the basic principle of the invention is applicable for any suitable lobe number ratio and lobe profile.
Referring now to
Furthermore, each one of the two compression parts 124, 128 and the two expansion parts 126, 130 are arranged in their own hermetically sealed enclosure 136 (see
A motor (not shown) and transmission (not shown) are operatively coupled to respective the twin-screw mechanisms 116, 118, wherein the rotation of male 120 and female 122 rotors is synchronised using the transmission (e.g. meshed gears that are mounted as a drive coupling assembly, for example, in the box 138. Box 138 also comprises an actuator (i.e. an efficient and controllable electrical motor), which is adapted to drive the twin-screw mechanisms via the transmission. Alternatively, the transmission (i.e. bearings, gear mechanism) may also be arranged in a different part of the housing, e.g. casing 140 surrounding the shafts 132, 134 of the twin-screw mechanisms 116, 118.
Referring now to
Referring now to
As shown in
As shown in
Referring now to
Referring now to
The variation of volumes of one of the chambers (i.e. chamber 1) in the compression part 128 and one of the chambers (i.e. chamber 1) in the expansion part 130 of the “cold” unit 102 is shown in
The following is a description of the individual processes taking place in the apparatus 100 of the present invention. A first working space 170 is formed during reciprocating compression and expansion of the working fluid (i.e. gas) trapped in chamber 1 of the compression part 128 and the expansion part 130 of the twin-screw mechanism 118 of the “cold” unit 102, and a second working space 172 is formed during reciprocating compression and expansion of a fluid volume (i.e. gas) trapped in chamber 2 of the compression part 128 and the expansion part 130 of the twin-screw mechanism 118 of the “cold” unit 102. Equivalent first and second working spaces (not shown) are formed by the twin-screw mechanism 116 of the “warm” unit 104.
To simplify the description of the process, chamber 1 of the “cold” unit 102 is considered as representative example for this embodiment of a cooling machine. The whole cycle (i.e. 360 degrees rotation of the twin-screw rotors 116, 118) can be split into three distinctive phases:
Phase 1:
The duration is from 0 degrees rotation of the shafts 132, 134 to the start of the overlap of the offset partially circumferential fluid channels 158, 160. Here, respective first set of fluid channels 158 remain aligned with corresponding first outlets 154. The first set of fluid channels 158 are fluidly connected to corresponding second set of fluid channels 160 through external fluid connections 162 (see
Phase 2:
The duration is from the start of the overlap to the completion of the overlap of the offset and partially circumferential fluid channels 158, 160. Close to the middle of the cycle, a fluid connection takes place between the chamber 1 volume of the compression part 128 and the chamber 1 volume of the expansion part 130. The duration of this phase is predetermined by the predefined overlap between the two axially offset and partially circumferential first and second sets of fluid channels 158, 160. The exact overlap is optimised to “smoothen” the gas exchange between the chamber 1 volumes of the compression part 128 and the expansion part 130, i.e. so as to minimise or even avoid pressure shocks between the compression part 128 and the expansion part 130.
Phase 3:
The duration is from the completion of the overlap to the full 360 degrees of the cycle. During this phase, respective second set of fluid channels 160 remain aligned with corresponding second outlets 156. As mentioned in the description of phase 1, each one of the first set of fluid channels 158 is fluidly connected to a corresponding one of the second set of fluid channels 160 through external fluid connections-162 (see
As mentioned previously, after the overlap period is completed, the volume of gas that is close to being compressed during the first half of the cycle in the compression part 128 will be expanding in the expansion part 130 during the second half of the cycle. Simultaneously, the volume of gas that is close to being expanded in the expansion part 130 will go through the compression process in the compression part 128 during the second half of the cycle. Thus, the magnitude of volume variation in the two formed working spaces 170 and 172 is approximately the same (see
Furthermore, it is understood that the variation of volume in each working space in the “warm” unit 104 “follows” the variation of volume of its corresponding paired working space in the “cold” unit 102, but with a delay of 90 to 120 degrees of the shaft angle (phase angle). In this particular example of the embodiment of the present invention, the variation of volume in each working space in the “warm” unit 104 may follow the variation of volume of its corresponding paired working space in the “cold” unit 102 with a 90 degree delay. However, it is understood by a person skilled in the art that other phase angles delays may be used between the “cold” unit 102 and the “warm” unit 104 so as to control the output of the Stirling-cycle apparatus 100 (e.g. cooling output).
A typical diagram of the variations of the paired working volume 174 in the “cold” unit 102 and the paired working volume 176 in the “warm” unit 104 is shown in
Alternative designs of the screw mechanism are shown in
Furthermore, a range of different rotor lobe geometry configurations and profiles may be used for the Stirling-cycle apparatus of the present invention, for example, utilising screw rotors with more than two lobes, provided that the phase angle between compression and expansion working spaces is suitable to generate adequate cooling/heating performance or output of mechanical work. Also, rotors and lobes may be made of different diameters and/or lengths, e.g. the diameter of the twin-screw rotors either in the “cold” unit may be made greater than that in the “warm” unit, or vice-versa, in order to augment power, cold or heat generation at relatively low temperature differences between the heat source and the heat sink.
In another alternative embodiment of the present invention, the drive coupling assembly may comprise an alternative valve mechanism 502 as illustrated in
In another alternative embodiment 600 of the present invention is shown in
In yet another alternative embodiment (not shown), different compression/expansion mechanisms (e.g. scroll and twin-screw) may be combined. However, it is understood that the variation of volumes (following a linear or nonlinear saw-tooth like function) is synchronised, so as to form a closed regenerative Stirling cycle.
Furthermore, connections of volumes in the embodiment, when utilising rotary conical screw mechanisms, may be similar to that with twin-screw rotors.
In addition, a multi-stage arrangement of the present invention (in cooling mode) may be used to achieve even lower temperatures as would be possible with the embodiment as described above. Furthermore, the Stirling-cycle machines of the present invention may be provided as a flat, box-type, cylindrical and other form. As mentioned previously, the heat exchangers or at least a portion of the heat-exchangers may be integrated into at least part of the casing or shaft of rotors, so as to minimise the size of the Stirling-cycle apparatus of the present invention. Alternatively, parts of the casing or shafts may be utilised as one of the heat exchangers.
It will be appreciated by persons skilled in the art that the above embodiment(s) have been described by way of example only and not in any limitative sense, and that various alterations and modifications are possible without departing from the scope of the invention as defined by the appended claims.
| Number | Date | Country | Kind |
|---|---|---|---|
| 1521880.3 | Dec 2015 | GB | national |
| Filing Document | Filing Date | Country | Kind |
|---|---|---|---|
| PCT/GB2016/053405 | 11/3/2016 | WO | 00 |
| Publishing Document | Publishing Date | Country | Kind |
|---|---|---|---|
| WO2017/098197 | 6/15/2017 | WO | A |
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| 20080098751 | Terada | May 2008 | A1 |
| 20120267898 | Mazza | Oct 2012 | A1 |
| 20140271308 | Franklin | Sep 2014 | A1 |
| Number | Date | Country |
|---|---|---|
| 412663 | May 2005 | AT |
| 10123078 | May 2002 | DE |
| H03286170 | Dec 1991 | JP |
| 2014037777 | Feb 2014 | JP |
| 20110120092 | Nov 2011 | KR |
| WO-01036801 | May 2001 | WO |
| WO-2005078269 | Aug 2005 | WO |
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| Number | Date | Country | |
|---|---|---|---|
| 20180372022 A1 | Dec 2018 | US |
