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Inventors
Kantor, Frederick W.
Application #
598366
Filed
Apr-9-1984
Published
Jun-25-1985
Current US Class
062/101
062/467
062/476
062/499
165/86
International Classes
F25B 015/00
Field of Search
62/499 62/467 62/476 62/101 165/86
US Patent References
| 4441337 |
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Rotary thermodyna... |
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Referenced by:
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Citation
Cite This Patent
More From Subclass 476
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Abstract
A rotary inertial thermodynamic absorptive system which can be used as a gas-driven heat pump, a heat-flow-driven gas pump, or, in combination, a heat splitter for moving low-grade heat energy from a lower temperature source to a higher temperature heat sink. In one embodiment, an absorptive type rotary inertial thermodynamic device employs overspill/underspill barriers in its absorption and desorption chambers to achieve counterflow heat exchange therebetween and to ensure effective control of thermodynamic impedance.
Claims
What is claimed is:
1. An absorptive rotary inertial thermodynamic device operable as a gas-flow driven heat pump in which heat is transferred from a lower temperature heat source to a higher temperature heat sink, comprising gas-flow driven heat pump means having a low pressure portion in which heat received directly or indirectly from said source evolves a volatile component as a gas from an absorbent liquor to concentrate said liquor and a high-pressure portion in which heat delivered to said sink is produced by absorption of the gas into the concentrated liquor; means supplying said gas at a relatively high pressure to said high-pressure portion of the gas-flow driven heat pump means; and means accepting said gas at a low pressure from said low-pressure portion of the gas-flow driven heat pump means.
2. An absorptive rotary inertial thermodynamic device as in claim 1, wherein said low-pressure portion includes a desorption chamber through which said liquor flows and in which heat from absorption in said high-pressure portion evolves said gas in said low-pressure portion to concentrate said liquor, and said high-pressure portion includes an absorption chamber through which the liquor concentrated in said desorption chamber flows absorbing the high pressure gas and from which heat of absorption is transferred at said higher temperature to said heat sink.
3. An absorptive rotary inertial thermodynamic device as in claim 2, wherein said absorption chamber is disposed radially outward of said separation chamber.
4. An absorptive rotary inertial thermodynamic device as in claim 3, further comprising pumping means for pumping said liquor from said absorption chamber to said separation chamber.
5. An absorptive rotary inertial thermodynamic device as in claim 1, said gas-flow driven heat pump means further including feedback means thermally coupling said high-pressure portion and said low-pressure portion of said gas-flow driven heat pump means for additionally concentrating the liquor in said low-pressure portion in accordance with the difference in temperature between said higher temperature heat sink and said source by transfer of at least a portion of the heat produced by absorption in said high-pressure portion to the liquor in said low-pressure portion.
6. An absorptive rotary inertial thermodynamic device as in claim 5, wherein said low-pressure portion includes a desorption chamber in which said liquor flows in one direction and heat from said source evolves said gas to concentrate said liquor; said high-pressure portion includes an absorption chamber in which the liquor concentrated in the desorption chamber flows in another direction countercurrent to said one direction; and said feedback means includes countercurrent heat transfer means transferring the heat of absorption of the liquor in said absorption chamber to the liquor in said desorption chamber.
7. An absorptive rotary inertial thermodynamic device as in claim 2 or 6, wherein said desorption chamber and said absorption chamber each include flow control means ensuring that the concentration of the liquor flowing therein varies in the direction of flow therein.
8. An absorptive inertial thermodynamic device as in claim 7, wherein said flow control means in said desorption chamber includes a plurality of underspill barrier means disposed between inlet and outlet ends of the desorption chamber and the flow control means in said absorption chamber includes a plurality of overspill barrier means disposed between inlet and outlet ends of the absorption chamber.
9. An absorptive rotary inertial thermodynamic device as in claim 8, wherein successive ones of said underspill barrier means define respective desorber stages within the desorption chamber such that the liquor at a radially more distant portion of each stage flows to a radially less distant surface of the liquor in the next successive desorber stage.
10. An absorptive rotary inertial thermodynamic device as in claim 9, wherein each said underspill barrier means comprises a first barrier open at a radially more distant edge thereof to permit flow therepast of said liquor from a radially more distant portion thereof, and a second barrier open at a radially closer edge to guide the liquor flowing past the first barrier to a radially closer portion of the next successive desorber stage.
11. An absorptive rotary inertial thermodynamic device as in claim 10, wherein the radially closer edges of the second barriers of successive ones of said underspill barrier means are stepped in radial distance.
12. An absorptive rotary inertial thermodynamic device as in claim 8, wherein successive ones of said overspill barrer means define respective absorber stages within the absorption chamber such that the liquor at a radially closer surface of the liquor in each such stage spills over to a radially more distant portion of the next respective stage.
13. An absorptive rotary inertial thermodynamic device as in claim 12, wherein each said overspill barrier means comprises a first barrier open only at a radially closer edge to permit spillage thereover of said liquor from a surface layer thereof, and a second barrier open at a radially more distant edge to guide the liquor spilling over said first barrier to a radially more distant portion of the next successive absorber stage.
14. An absorptive rotary inertial thermodynamic device as in claim 13, wherein the radially closer edges of the first barrier of successive ones of said overspill barrier means are stepped in radial distance.
15. An absorptive rotary inertial thermodynamic device as in claim 1, further comprising heat-flow driven gas-pumping means having a high-pressure side constituting said means supplying said gas at high pressure and a low-pressure side constituting said means accepting said gas at low pressure.
16. An absorptive rotary inertial thermodynamic device operable as a heat-flow driven gas pump in which gas supplied from a low-pressure inlet is transferred to an outlet at a higher pressure, comprising heat-flow driven gas pump means having for at least one concentration of liquor at least one relatively low temperature, low pressure portion in which gas received from said inlet is absorbed into said absorbent liquor to dilute the same, and for the same said or lesser concentration of said liquor, at least one relatively high pressure portion in which the gas delivered to said outlet is evolved from said liquor; heat source means supplying input heat at a higher temperature, directly or indirectly, to said high-pressure portion of said heat-flow driven gas pump means; and low-temperature heat sink means accepting said heat at lower temperature than said input heat from at least one portion of said low-pressure portion of said heat-flow driven gas pump means.
17. An absorptive rotary inertial thermodynamic device as in claim 16, wherein said low-pressure portion includes an absorption chamber through which said liquor flows and in which the low-pressure gas from said inlet is absorbed to dilute said liquor and liberate heat to be transferred to said low-temperature heat sink means; and said high-pressure portion includes a desportion chamber through which the liquor diluted in said absorption chamber flows receiving heat directly or indirectly from said heat source means and evolving the gas at high pressure to be delivered to said outlet.
18. An absorptive rotary inertial thermodynamic device as in claim 17, wherein said desorption chamber is disposed radially outward of said absorption chamber.
19. An absorptive rotary inertial thermodynamic device as in claim 18, further comprising pumping means for pumping said liquor from said desorption chamber to said absorption chamber.
20. An absorptive rotary inertial thermodynamic device as in claim 16, said heat-flow driven gas pump means further including feedback means thermally coupling said high-pressure portion and said low-pressure portion of said heat-flow driven gas pump means for additionally diluting the liquor in the low-pressure portion in accordance with the difference in gas pressure between said inlet and said outlet by transfer of heat produced in said low-pressure portion to the liquor in said high-pressure portion.
21. An absorptive rotary inertial thermodynamic device as in claim 20, wherein said low-pressure portion includes an absorption chamber in which said liquor flows in one direction and gas supplied to said inlet is absorbed to dilute said liquor; said high-pressure portion includes a desorption chamber in which the liquor diluted in said absorption chamber flows in another direction countercurrent to said one direction; and said feedback means includes countercurrent heat transfer means transferring heat of absorption of the liquor in said absorption chamber to the liquor in said desorption chamber.
22. An absorptive rotary inertial thermodynamic device an in claim 17 or 21, wherein said desorption chamber and said absorption chamber each include flow control means ensuring that the concentration of the liquor flowing therein varies in the direction of flow therein.
23. An absorptive rotary inertial thermodynamic device as in claim 22, wherein said flow control means in said desorption chamber includes a plurality of underspill barrier means disposed between inlet and outlet ends of the desorption chamber and the flow control means in said absorption chamber includes a plurality of overspill barrier means disposed between inlet and outlet ends of the absorption chamber.
24. An absorptive rotary inertial thermodynamic device as in claim 23, wherein successive ones of said underspill barrier means define respective desorber stages within the desorption chamber such that the liquor at a radially more distant portion of each stage flows to a radially less distant surface of the liquor in the next successive desorber stage.
25. An absorptive rotary inertial thermodynamic device as in claim 24, wherein each said underspill barrier means comprises a first barrier open at a radially more distant edge thereof to permit flow therepast of said liquor from a radially more distant portion thereof, and a second barrier open at a radially closer edge to guide the liquor flowing past the first barrier to a radially closer portion of the next successive desorber stage.
26. An absorptive rotary inertial thermodynamic device an in claim 25, wherein the radially closer edges of the second barriers of successive ones of said underspill barrier means are stepped in radial distance.
27. An absorptive rotary inertial thermodynamic device as in claim 23, wherein successive ones of said overspill barrier means define respective absorber stages within the absorption chamber such that the liquor at a radially closer surface of the liquor in each such stage spill over to a radially more distant portion of the next successive stage.
28. An absorptive rotary inertial thermodynamic device as in claim 27, wherein each said overspill barrier means comprises a first barrier open only at a radially closer edge thereof to permit spillage thereover of said liquor from an exposed surface layer thereof, and a second barrier open at a radially more distant edge to guide the liquor spilling over said first barrier to a radially more distant portion of the next successive stage.
29. An absorptive rotary inertial thermodynamic device as in claim 28, wherein the radially closer edges of the first barriers of successive ones of said underspill barrier means are stepped in radial distance.
30. An absorptive rotary inertial thermodynamic device as in claim 16, further comprising gas-flow driven heat-pumping means having a high-pressure side coupled to said gas outlet and a low-pressure side coupled to said gas inlet.
31. An absorptive rotary inertial thermodynamic device in which heat is transferred from a lower-temperature heat source to a higher-temperature heat sink, comprising a low-pressure portion in which heat received directly or indirectly from said source evolves a volatile component as gas from an absorbent liquor to concentrate the same; a high-pressure portion in which heat delivered to said higher-temperature heat sink is produced by absorption of the gas into the concentrated liquor; and self-adjusting means thermally coupling said high-pressure portion with said low-pressure portion for additionally concentrating the liquor in the low-pressure portion in accordance with the difference in temperature between said higher-temperature sink and said lower-temperature heat source by transfer of heat produced in said high-pressure portion to the liquor in said low-pressure portion.
32. An absorptive rotary inertial thermodynamic device in which gas supplied at low pressure from a low-pressure inlet is transferred at high pressure to an outlet, comprising for at least one concentration of liquor at least one relatively low-pressure portion in which gas received from said inlet is absorbed into an absorbent liquor to dilute the same; and for the same said concentration or lesser concentration of said liquor at least one high-pressure portion in which the gas delivered to said outlet is evolved from the diluted liquor; and self-adjusting means thermally coupling said low-pressure portion and said high-pressure portion for additionally diluting the liquor in the low-pressure portion in accordance with the difference in pressure between said oulet and said inlet by transfer of heat produced in said low-pressure portion to the liquor in said high-pressure portion with relative liquor concentrations, at such thermally coupled locations, such that absorption at the lower gas pressure produces heat at a temperature greater than or equal to the temperature at which heat can be absorbed by the liquor in the higher pressure portion evolving gas from said liquor at said higher pressure.
33. Absorptive rotary inertial thermodynamic method of pumping heat from a low temperature heat source to a higher temperature heat sink and driven by flow of a gas, comprising the steps of receiving heat from said source in a low-pressure chamber in which an absorbent liquor flows; evolving a volatile component as said gas from said liquor to concentrate the liquor; outletting the evolved gas at low pressure from said low-pressure chamber; supplying said gas at high pressure to a high-pressure chamber; and producing heat to be delivered to said sink by absorbing the gas into the concentrated liquor in said high-pressure chamber.
34. Absorptive rotary inertial thermodynamic method of pumping gas from an inlet at a lower pressure to an outlet at a higher pressure, and driven by flow of heat, comprising the steps of receiving the gas at low pressure in a low-pressure chamber in which an absorbent liquor flows; for at least one concentration of liquor, absorbing the gas into said liquor in the low-pressure chamber to dilute the liquor; supplying heat to a high-pressure chamber in which the diluted liquor flows; and evolving in said high-pressure chamber from said liquor at the same said or lesser concentration said gas to be delivered to said outlet at high pressure by heating the diluted liquor.
35. Absorptive rotary inertial thermodynamic method of pumping heat from a low temperature heat source to a higher temperature heat sink and driven by flow of a gas, comprising the steps of receiving heat from said source in a low-pressure chamber in which an absorbent liquor flows; evolving a volatile component as a gas from said liquor to concentrate said liquor; producing heat to be delivered to said sink by absorbing the gas into the concentrated liquor in a high-pressure chamber; and thermally coupling said low-pressure chamber with said high-pressure chamber for additionally concentrating the liquor in the low-pressure chamber in accordance with the difference in temperature between said heat sink and heat source by transfer of heat produced in said high-pressure chamber to the liquor in said low-pressure chamber.
36. Absorptive rotary inertial thermodynamic method of pumping gas supplied from an inlet at low pressure to an outlet at a higher pressure, and driven by flow of heat, comprising the steps of receiving the gas from said inlet at a low pressure in a low-pressure chamber in which an absorbent liquor flows; absorbing the gas into the liquor in the low-pressure chamber to dilute the liquor; evolving said gas at the higher pressure in a high-pressure chamber, with the evolved gas to be delivered to said outlet; and thermally coupling the low-pressure chamber with the high-pressure chamber for additionally diluting the liquor in the low-pressure chamber in an amount that varies with the difference in gas pressure between said outlet and said inlet by transfer produced in said low-pressure chamber to the liquor in said high-pressure chamber.
37. In an absorptive rotary inertial thermodynamic device including an absorption chamber in which a gas is absorbed in an absorbent liquor, a desorption chamber having means for conducting heat from a source into a liquor in said separation chamber to drive said gas out of said liquor, means for pumping said liquor through a closed circuit path including said absorption and desorption chambers, means for conducting heat from the liquor in said desorption chamber to a sink, and means for conducting from said absorption chamber to said desorption chamber heat developed by the absorption of said gas into said liquor, the improvement wherein said desorption and absorption chambers include self-adjusting means thermally coupling said desorption chamber and said absorption chamber for absorbing additional gas into said liquor in said absorption chamber and separating additional gas from the liquor in said desorption chamber in accordance with the difference in temperature between said source and sink.
38. In an absorptive rotary inertial thermodynamic device including an absorption chamber in which gas is absorbed into an absorbent liquor, a desorption chamber having means for conducting heat into liquor in said desorption chamber to drive said gas out of said liquor, means for pumping said liquor through a closed circuit path including said absorption and desorption chambers, and means for conducting from said absorption chamber to said desorption chamber heat developed by the absorption of said gas into said liquor; the improvement wherein for at least one concentration of liquor, in at least one portion of said absorption chamber said gas is absorbed into said liquor to dilute the same and, for the same said or a lesser concentration of said liquor, in at least one portion of said desorption chamber said gas is evolved from said liquor.
Description
This invention relates to rotary thermodynamic apparatus and methods. More specifically, this invention relates to means and methods for stabilizing the operation of rotary inertial thermodynamic apparatus.
A theoretically highly efficient but impractical refrigerator device has been proposed in U.S. Pat. No. 2,393,338 to J. R. Roebuck, and "A Novel Form of Refrigerator" 16 Journal of Applied Physics 285-295, May, 1945 by J. R. Roebuck. The basic form of the device proposed by Roebuch is shown in FIG. 1 of the drawings. The tube 11, which is supported in bearings 10, is rotated at a very high speed about central axis 1 in the direction indicated by the arrow 12 by means of a drive motor (not shown).
Compressed air is introduced into tube 11 at its inlet 13. It travels through section 2, parallel to the axis, through section 4 towards the axis, through section 5 parallel to the axis, and exits at outlet 14. As the gas moves radially outward it is subjected to centrifugal compressive forces. While moving in the section 3, the gas is compressed and heated by the centrifugal force created by the rotation of the tube 11. At least part of the heat of compression is removed from the gas in section 3 by heat exchange means (not shown) such as water flowing in cooling coils.
While moving in section 4, the gas expands, due to the reduction of the distance of the gas from the axis 1 and the concomitant reduction of the centrifugal force acting on the remaining mass of gas between it and the axis, and the gas becomes substantially cooler due to its expansion. The cooled gas then flows out of the outlet opening 14 for use in refrigeration.
The system described above is one form of a "rotary inertial thermodynamic system", as the latter expression is used herein. In such a system is performed a "rotary inertial thermodynamic method", as that expression is used herein. Other, greatly improved forms of such a system and method are disclosed in my U.S. Pat. No. 3,470,704, issued Oct. 7, 1969, and my U.S. Pat. No. 3,808,828. The disclosures of that patent and application hereby are incorporated in this patent application by reference.
It is an object of the present invention to provide a rotary inertial thermodynamic system and method in which the fluid flow is stable. It is another object of the present invention to provide such a system which is relatively compact, lightweight, uncomplicated and inexpensive, and which is capable of operating under a wide variety of conditions and in a wide variety of environments.
In accordance with the present invention, the foregoing objects are met by the provision of rotary inertial thermodynamic apparatus and methods in which the flow is stabilized by controlling the impedances to fluid flow in the system so that the overall pressure drop of the fluid flow in the system is made to increase with increasing fluid flow rate.
It is still another object of the present invention to provide a rotary inertial thermodynamic system and method in which the fluid flow, chemical concentrations, and thermal flows are stable. It is another object of the present invention to provide such a system which accepts heat from a heat source at an intermediate temperature and rejects a portion of that heat to a heat sink at a lower temperature while delivering another portion of that heat to a heat sink at a temperature higher than the input temperature.
It is yet another object of the present invention to provide a counterflow absorptive/desorptive heat exchange apparatus and method permitting more efficient use of high-grade heat to operate an absorption cycle heat pump and/or refrigerator.
It is a further object of the present invention to provide an absorption cycle system in which there is a rapid removal of liquor with changed concentration so as to reduce advantageous by the thickness of the liquor surface layer through which molecules must diffuse during absorption and/or desorption.
In accordance with the present invention, the foregoing objects are met by the provision of rotary inertial thermodynamic apparatus and methods in which control of the rotary inertial thermodynamic impedances for chemical concentration changes, temperature distributions, pressure distributions, and flows provide for self-adjustment of the system so as to use the energy provided by the flow of heat from the intermediate temperature heat source to the lower temperature heat sink to drive the pumping of heat from the intermediate input source to a higher temperature output.
The foregoing and other objects and advantages will be set forth in or apparent from the following description and drawings.
In the drawings are included several graphs. It should be understood that the graphs illustrate the relationships between the variables qualitatively, and not quantitatively.
BRIEF DESCRIPTION OF THE DRAWINGS
FIG. 1 is a schematic drawing of a prior art device;
FIGS. 2 and 3 are graphs used in explaining some of the principles of the present invention;
FIGS. 4 and 5 each show separate embodiments of the invention;
FIGS. 6 through 9 are graphs illustrating certain operational features of the invention;
FIG. 10 shows, schematically, another embodiment of the invention;
FIG. 11 is a graph illustrating operational principles of the embodiment shown in FIG. 10;
FIG. 12 is a schematic perspective view of a preferred form of the device shown in FIG. 10;
FIG. 13 is a cross-sectional view of the device shown in FIG. 12;
FIG. 14 is a schematic drawing of another embodiment of the present invention;
FIGS. 15 and 16 are graphs illustrating the operation of the device shown in FIG. 14;
FIG. 17 is a schematic drawing of another embodiment of the invention;
FIGS. 18 through 21 are graphs depicting the operation of various embodiments of the invention;
FIGS. 22, 23 and 24 each show a further embodiment of the invention;
FIG. 25 is a graph depicting the operation of the device shown in FIG. 24;
FIG. 26 is a schematic drawing of another embodiment of the invention;
FIG. 27 is a graph depicting the operation of the device shown in FIG. 26;
FIGS. 28 through 36 each show a separate embodiment of the invention;
FIGS. 37 through 40 show various impedance control devices for the invention;
FIGS. 41 through 52 and 54 each show another embodiment of the invention;
FIGS. 53 and 55 are graphs illustrating operational features of the devices shown in FIGS. 52 and 54, respectively;
FIGS. 56 through 58 show another embodiment of the invention;
FIGS. 59 through 64 each show another embodiment of the invention;
FIGS. 65 and 66 show another embodiment;
FIGS. 67 and 68 show another embodiment;
FIG. 69 is an elevation, partly schematic view of another embodiment of the invention;
FIG. 70 is a cross-sectional view taken along line 70--70 of FIG. 69;
FIG. 71 is a cross-sectional view taken along lines 71--71 of FIG. 69;
FIG. 72 is a cross-sectional view taken along line 72--72 of FIG. 69;
FIGS. 73 and 74 each show another embodiment of the invention;
FIG. 75 is a schematic drawing illustrating certain operational features of the invention;
FIG. 76, formed by joining FIGS. 76A and 76B, is a schematic drawing illustrating yet another embodiment of the invention; and
FIGS. 77 and 78 are enlarged schematic drawings illustrating the rotary inertial thermodynamic impedance control means of the embodiment of FIG. 76.
It will aid in the understanding of this invention to divide the pressure drop in the fluid, as it flows through the rotating conduit of a rotary inertial thermodynamic system, into a "thermodynamic" component and "mechanical" component.
The "mechanical" pressure drop is caused by friction between the fluid and the walls of the conduit it flows in. The mechanical pressure drop increases with F, the fluid flow rate in the system. FIG. 2 shows a curve .DELTA.P.sub.M which describes the typical variation of mechanical pressure drop with flow rate in the device of FIG. 1.
The mechanical pressure drop can be thought of as the pressure drop due to the "mechanical impedance" to flow through the system.
The second component of total pressure drop is the "thermodynamic pressure drop", .DELTA.P.sub.T. This component of the total pressure drop is that caused by thermodynamic conditions of the flow. One example of such a pressure drop is that caused by the difference in the temperatures of the gas in the tube sections 3 and 4 in the device shown in FIG. 1. The gas in section 4 is cooler and, therefore, denser than the gas in section 3. Therefore, the back pressure created by centrifugal action on the gas in section 4 is greater than the forward pressure (i.e., pressure tending to encourage flow in the direction indicated by the arrows in FIG. 1) created by the same centrifugal action on the gas in section 3, with the result that there is a net pressure drop due to this temperature difference. Other examples of thermodynamic pressure drops will be given below.
In FIG. 2, curve .DELTA.P.sub.T describes the typical variation of the above-described type of thermodynamic pressure drop with F, flow rate, for the device in FIG. 1. It can be seen that P.sub.T decreases with increasing flow rate, and the curve P.sub.T thus has a negative slope. The reason for this is that as the flow rate increases, the amount of heat removed from the gas flowing in tube section 3 decreases due to the fact that there is a decreasing amount of time for the heat exchange means to extract heat from each portion of gas passing through it. Thus, the temperature difference between the gas in sections 3 and 4 decreases and the thermodynamic pressure drop decreases.
The thermodynamic pressure drop can be thought of as the pressure drop caused by the "thermodynamic impedance" to fluid flow through the system.
In order to calculate the mechanical and thermodynamic impedances of the system, an operating point such as point 21 in FIG. 2 is selected. Then tangents 22 and 23 to the curves are drawn. For small fluctuations of flow rate or pressure drop near the operating point 21, the thermodynamic flow impedance is proportional to the slope of line 22, and the mechanical impedance is proportional to the slope of line 23.
In accordance with one basic principle of the present invention, it has been found that stability of the overall rotary inertial thermodynamic system can be maintained if the total mechanical and thermodynamic impedance of the system is positive (for small fluctuations about an operating point); that is, if the total pressure drop in the system increases with increasing flow rate. This relationship is illustrated in FIG. 3. Curve 24 represents a typical system with a total impedance characteristic which is positive; i.e. the pressure drop across the system increases with increasing flow. Curve 25 represents a system having a total pressure drop which is independent of flow rate. Curve 26 represents a system having a total pressure drop which decreases with increasing flow. For a given flow rate, e.g., the rate at line 27, the following conditions exist.
The system represented by curve 24 can operate stably. Any disturbance which tends to increase the flow gives rise to an increasing pressure drop through the system, which tends to reduce the flow. However, a system of the type described by curve 25 is not stable. A small purturbation of its operation, which would tend to increase the flow, does not give rise to a corresponding increase in pressure drop to restore the system to its initial condition. All decreasing curves, of which 26 is representative, describe systems whose operation is unstable, i.e. any purturbation of the system which tends to increase the flow gives rise to a reduced pressure drop which, in turn, allows the flow to increase further. There is no restorative mechanism to preserve the operation of the system at a stable level of flow.
There are several ways in which the operation of a rotary inertial thermodynamic system can be stabilized in accordance with the above-stated principles. One of these is to increase the mechanical impedance to flow, so that there is a composite impedance which has a rising characteristic at the operating point, 27. Another alternative, and one which can be more desirable from the standpoint of efficiency and flexibility of operation, is to reduce the thermodynamic impedance and its associated characteristic time lag for the exchange of heat so that, again, the system has a composite impedance displaying a rising characteristic curve, but achieves this without sacrifice of flow capability or mechanical efficiency.
FIG. 4 illustrates two different mechanisms by means of which stability of the system illustrated in FIG. 1 can be assured. Both use the principle of increasing the relative mechanical impedance so that it dominates the thermodynamic impedance, thus assuring a positive total impedance.
The first way in which a high mechanical impedance might be obtained is by using a stationary positive displacement pump 28 as the gas compressor. Such a pump is, for example, a reciprocating piston pump, a sliding vane pump, or any other pump offering high impedance to the flow of gas to the system from the pump. Cold air from the rotary portion of the system flows through refrigeration apparatus (not shown) including heat exchange means using the cold gas for cooling. Appropriate rotary seals are used to provide a gastight coupling between the rotary and stationary portions of the system. Relatively dense gases, such as "Freon 12" can be used to reduce the required rotational speed of the device. The very high impedance of the pump provides the positive impedance which stabilizes the system. It should be pointed out that this means of stabilization is not the preferred means, as will be more fully explained below.
Another stabilizing means shown in FIG. 4 is the use of a heat exchanger 29 of a form which restricts the gas flow in conduit 11 and thus adds a considerable amount of mechanical impedance to the system, and, therefore, reduces the thermodynamic impedance relative to the mechanical impedance, and gives a positive total impedance characteristic to the system. For example, the heat exchanger 29 can be a porous, thermally conductive plug such as a sintered metal plug; or one with a multiplicity of fine conduits. Metal bodies with many tiny conduits can be formed by use of the technology utilized in the production of grids for vacuum tubes by extrusion of a composite body and subsequent etching of material from this body to leave fine passages. Other exchanger structures include very finely-formed fin structures produced by extrusion, hobbing, or any of a wide variety of other techniques well known in the art of constructing heat exchange devices, and in the art of constructing mechanical damping devices to provide impedance to the flow of a gas. Of course, the device 29 should function well as a heat exchanger in order to efficiently remove the heat of compression from the gas in section 3 of the tube 11.
A useful modification of the latter embodiment also is shown in FIG. 4. In this modification, one or more additional conduit sections 33 is connected in parallel with the first such section. The sections preferably are added in opposed pairs in order to maintain rotational balance of the rotary portion of the system. A porous plug 29 or similar restrictor is located in the outwardly-extending section of each such parallel section. Thus, each separate section contains its own stabilizing means. This has the advantage of preventing unequal sharing of the fluid flow or even reversal of flow in some sections, which well might occur if the only stabilizing means were a single high-impedance external pump 28. Additional plugs 30 can be inserted in the inwardly-extending portions of the conduits to add further stabilization.
FIG. 5 illustrates the basic features of a compressor constructed in accordance with the present invention. Thermodynamic compressors of this type are described in my U.S. Pat. No. 3,470,704 and in my U.S. Pat. No. 3,808,828, as utilized in a sealed, closed-circuit rotating thermodynamic system to provide the actuating pressure to operate refrigeration apparatus.
FIG. 6 is a graph showing the qualitative relationships between the pressure P and temperature T of the gas in the compressor of FIG. 5. Referring to both FIGS. 5 and 6, a tube 52 like the tube 11 in FIG. 1 is provided, and is rotated as is the FIG. 1 device. Gas enters tube 52 at an inlet 53 at a pressure P.sub.17 and a temperature T.sub.17. During the motion of this gas radially outwardly through tube section 50, the pressure and temperature increase essentially adiabatically to a new pressure and temperature, P.sub.18, T.sub.18, corresponding to the pressure and temperature at location 48 in FIG. 5. Then, the working fluid returns towards the axis through tube section 51. In tube section 51 is located heat exchange means 46, such as a porous plug, which is provided to conduct heat from a source (not shown) in the direction of arrow 61 into the working fluid to maintain the working fluid at the temperature which it had reached during the adiabatic compression and which it possessed at location 48. Due to the heat exchanger 46, the expansion of the working fluid in part of tube section 51 is essentially isothermal and is represented by the isothermal decrease in pressure to P.sub.20, T.sub.18, shown in FIG. 6. In the remainder of section 51, the pressure and temperature drop to P.sub.19, T.sub.19. At each radial distance from the axis of rotation 1 in FIG. 5, a volume element in tube section 51 is at a higher temperature than a corresponding volume element at a corresponding radial distance from axis 1 in tube section 50. Therefore, the density in such a volume element in tube 51 is lower than that of its corresponding element in tube 50. Centrifugal forces acting upon the column of gas extending radially in tube section 51 exert a smaller total backward pressure than the forward pressure exerted by centrifugal forces acting on the denser column of gas in tube section 50. This gives rise to a net forward driving pressure which drives gas through the system in the direction indicated by arrow 54, i.e. P.sub.19 is greater than P.sub.17. Also, T.sub.18 is greater than T.sub.17. Thus, this device acts as a thermodynamic compressor which can be expected to have high thermodynamic efficiency.
The same physical effects that give rise to a relationship between thermodynamic pressure drop and flow rate for working fluid in a refrigeration system of the type shown in FIG. 1, also gives rise to a pressure and flow relationship in the case of a compressor of the type shown in FIG. 5. The compression arises from a difference in temperature between gases in tube section 51 and tube section 50 in FIG. 5. This temperature difference is maintained by a flow of heat into the working fluid as indicated at 61. This flow of heat is not instantaneous, that is, there is a characteristic time lag associated with the process of heating the fluid. The faster the flow of working fluid through the system, the less exposure time the working fluid has in heat exchange means 46. This results in less effective heat exchange and a smaller temperature difference between the working fluid in tube section 51 and that in tube section 50. Because of this, the thermodynamic pumping action of this pump decreases with increasing flow rate. FIG. 7 shows the relationship between the flow rate F and the pumping pressure .DELTA.P.sub.T produced by the compressor shown in FIG. 5. The abovedescribed behavior is indicated generally in FIG. 7 by curve portion 101, which represents a decrease in thermodynamic pumping pressure with increase in flow.
For the purpose of explanation, suppose that we were to force working fluid backwards through this pump against the pressure gradiant maintained by the thermodynamic pumping action. In that case, the working fluid would be flowing in the direction opposite to that indicated by arrow 54 in FIG. 5. The first consequence of this reversal in direction of flow is that the portion of tube section 51 which is located between heat exchanger 46 and the axis 1 no longer contains heated working fluid. The working fluid heated by heat exchanger 46 is carried radially outwardly and returns towards the axis 1 in tube section 50. Tube section 50 thus contains heated working fluid. The region 48 of the tube 52, which is located further from the axis 1 than the radius 63 of the radially outermost point of heat exchanger 46, can be regarded as an adiabatic region containing heated working fluid whose temperature and pressure depend only upon the radial distance of the point at which temperature and pressure are measured from the axis of rotation 1. The working fluid returns to axis 1 in tube section 50, and its expansion therein can be regarded as essentially adiabatic. In the tube section 65, between radial distance 63 and the axis 1, the temperature of the working fluid in tube section 50 is greater than the temperature of the working fluid in tube section 51 at corresponding radial location. For this reason, the working fluid in tube section 65, when acted upon by centrifugal forces, exerts a smaller pressure in the direction indicated by arrow 54 than does the corresponding body of working fluid in tube section 51, in the direction opposite to the arrow 54. For this reason, there is a net pumping action in the reverse direction. This is depicted graphically in FIG. 7 by curve segment 102. As the flow rate in the reverse direction is increased, the working fluid passing through heat exchanger 46 interacts less efficiently with it. One consequence of this is that the working fluid reaches its maximum temperature just at the point where it leaves the heat exchanger. For this reason, the working fluid in tube section 65, at those radii corresponding to locations within tube section 51 occupied by heat exchanger 46, has a higher temperature than does the working fluid within the heat exchanger itself. For this reason, it is possible for the pumping action to increase for working fluid flowing more quickly in the reverse direction. If the flow of the working fluid backwards is made very great, the effect of heat exchanger 46, in changing the temperature of the working fluid passing through the system, becomes essentially negligible. In that case, there is no physical effect of temperature difference, and the thermodynamic pumping pressure differential goes asymptotically to zero for large flows.
The most pronounced feature illustrated by FIG. 7 is that the thermodynamic pumping mechanism can work to pump working fluid in either the positive flow direction or the negative flow direction. There is a narrow region within which the pressure of the thermodynamic pumping action in the two different directions is joined, here designated by dashed curve segment 103. The details of the form of curve segment 103 depend upon the details of heat exchange and convection and are profoundly affected by the geometry of the device. However, essentially none of the possible variations of the geometry can stabilize the system at the zero flow point. Note that the pumping effect is an increase in pressure at the output compared to the pressure at the intake for the pumping system. For this reason, it has a negative sign, compared to a mechanical resistance to flow, and is designated in FIG. 7 by-.DELTA.P.sub.T.
FIG. 8 illustrates the relationship between the flow rate F and the mechanical pressure drop .DELTA.P.sub.M in the device of FIG. 5. Mechanical resistance to flow always opposes the flow, i.e. for a positive flow there is a positive drop in pressure, taking the inlet pressure minus the outlet pressure. Using that convention, for a positive flow there is a positive .DELTA.P.sub.M, and for a negative flow a negative .DELTA.P.sub.M. This is indicated schematically by curve 104 in FIG. 8. Note that in the region near zero flow rate where there is the greatest problem in stabilizing the thermodynamic compression mechanism, the effect of the mechanical impedance to flow is the least.
The overall performance of a single pumping loop of the type illustrated in FIG. 5 is represented graphically in FIG. 9. .DELTA.P in FIG. 9 is the total pressure difference, defined as pressure at inlet 53 minus pressure at outlet 49, for compressor 52 in FIG. 5. .DELTA.P is defined as the algebreic sum of the mechanical pressure and the thermodynamic pressure differences, observing the sign convention defined above. Curve portion 105 in FIG. 9 represents flow in the forward direction, and portion 106 represents flow in the reverse direction. The curve portion in the fourth quadrant represents operation of the system as a compressor, driving its own flow in the forward direction. The region in the second quadrant represents the compressor driving its own flow in the reverse direction. .DELTA.P, for small reverse flow, is represented very approximately by dashed curve portion 108. The intersection of a line 107 and line 105 is a selected operating point Q, with flow F.sub.Q and compression C.sub.Q. The following equation expresses the relationship between C and F: C=-.DELTA.P(F), when .DELTA.P(F) is a function of the flow F. Line 109 is tangent to curve 105 at the operating point Q, and the slope of line 109, for small fluctuations of flow near point Q, gives a measure of the rate of change of .DELTA.P with F. It can be seen that the tangent line 109 is positive. Thus, the compression C which serves to drive the flow F in the forward direction decreases with increasing flow. For this reason, operation of the system at the selected operating point Q, is stable with respect to small fluctuations in flow. A fluctuation tending to increase the flow decreases the driving compression available to continue the flow. This allows the flow to return to its original value. Similarly, a fluctuation which tends to decrease the flow increases the amount of compression available to drive the flow, which in turn restores the flow to its original value. It is important to select the operating point Q far enough from the zero flow point so that fluctuations in the temperatures and pressures of operation of the system will not produce an excursion in the flow rate sufficient to take the flow from its selected operating point through the zero value and force the system into operation in a reverse flow mode.
One of the most important single aspects of the operation of a rotary inertial thermodynamic compressor of the kind shown in FIG. 5 is that stable operation requires that the forward flow be substantially different from zero. Thus, stable operation of the compressor requires that the external impedance not be so large in relation to the internal impedances of the compressor as to reduce the flow to such a small level that the fluctuations in the pressure generated by the compressor or in the pressure reflected by the external load would force the flow of working fluid to reverse, even briefly.
FIG. 10 shows, schematically, a compressor 199 in which several single-loop compressors of the type shown in FIG. 5 are connected together in series ("cascaded") in order to provide an increase in compression over that available from a single loop. Such a cascaded compressor is particularly desirable in uses in which it is not possible, because of limitations on input and output temperature, rotational speed, size of the device, or because of the nature of the working fluid, to achieve the desired compression to allow stable operation in a single loop. The compressor 199 includes a conduit 200 which has three U loops. At appropriate places within the conduit 200 are located heat exchangers, e.g., porous plugs, 133, 135, 137, 139, and 141. Working fluid (gas) enters the system at inlet 131 and travels radially outward through conduit section 132, experiencing essentially adiabatic compression. It returns towards axis 1 in conduit section 134, passing first through heat exchanger 133 within which it experiences expansion, which can be regarded as essentially isothermal. The working fluid then continues towards axis 1 through region 144 of conduit section 134, within which its continued expansion as it approaches axis 1 is essentially adiabatic. In moving from inlet 131 to station 201, located on the rotational axis 1, the working fluid has passed through a rotary thermodynamic compressor stage essentially like that described in FIG. 5. In FIG. 10 this first stage of compression is given reference numeral 202. Second and third stages 203 and 204 follow stage 202.
The second stage 203 includes conduit sections 136 and 138, and heat exchange means 135 and 137. Working fluid continues from station 201 through heat exchange means 135, moving radialy outwardly. Within heat exchange means 135 the compression of the working fluid can be regarded as essentially isothermal, i.e. the heat exchanger 135 allows heat of compression to leave the working fluid during compression. After passing through heat exchange means 135, the working fluid continues through the region 145 beyond heat exchanger 135, within which its compression, as it moves radially outward, is essentially adiabatic. The working fluid then returns to axis 1 through heat exchanger 137, within which its expansion on returning towards the axis is essentially isothermal, and then through conduit region 146, within which its further expansion is essentially adiabatic.
The working fluid next flows through the third stage 204, in which it is acted upon in substantially the same manner as in the second stage 203. The compressed gas emerges from an outlet opening 143.
It should be understood that heat is added to the working fluid from one or more heat sources through each of the heat exchangers 133, 137 and 141 so as to maintain the flow through those exchangers essentially isothermal.
FIG. 11 shows the relationships which are believed to exist between the pressure and temperature of the working fluid as it passes through compressor 199 shown in FIG. 10. Point 149 represents the pressure and temperature of the working fluid at the inlet 131. Line segment 150 describes the adiabatic compression of the working fluid in conduit section 132. Line segment 152 describes the isothermal expansion of working fluid returning towards axis 1 in heat exchanger 133. Line segment 154 describes the adiabatic expansion of working fluid within region 144 of conduit section 134. The point 155 gives the pressure and temperature of the working fluid at station 201 in FIG. 19. This is its temperature and pressure after having completed passage through the first compressor stage 202. The arrows next to the various portions of the curve in FIG. 11 indicate the progress of a volume element of working fluid through the compressor 199.
The difference in pressure between points 155 and 149, designated C.sub.1, represents the compression provided by the first stage 202 of compression.
In the second stage of compression, line segment 156 designates the isothermal compression of working fluid within heat exchanger 135. Line segment 157 represents the adiabatic compression of working fluid within region 145 of conduit segment 136. Line segment 159 represents the isothermal expansion of working fluid within heat exchange means 137 in conduit segment 138. Line segment 160 represents the essentially adiabatic further expansion of working fluid in region 146 of conduit section 138. Point 170 represents the pressure and temperature at station 205. This completes the second stage of compression 203, and is designated C.sub.2 in FIG. 11.
Compression stage 204 is represented by line segments 171, 172, 173 and 174, corresponding respectively to isothermal compression within heat exchange means 139 of conduit segment 140, adiabatic compression within region 147 of conduit 140, isothermal expansion within heat exchange means 141 of conduit segment 152 and adiabatic expansion in region 148 of conduit segment 152. Point 161 in FIG. 11 represents the pressure and temperature of the working fluid at outlet 143 of the compressor 199. The difference in pressure between point 161 and point 170, designated C.sub.3 in FIG. 11, is the compression occurring within the third stage 204 of the compressor. The total compression provided by the entire compressor is represented by the difference in pressure between point 161 and point 149 (the sum of C.sub.1, C.sub.2 and C.sub.3), and is designated in FIG. 11 by C.
The provision of adiabatic compression in the first section 132 of the compressor 199 is optional. If desired or necessary, a heat exchanger positioned like heat exchangers 135 and 139 can be used to make the compression isothermal. Ordinarily, however, the gas entering the compressor will be cool and isothermal compression in the first stage will be unnecessary.
Except for the above-described optional feature of the first stage of the compressor 199, all of the stages preferably are essentially identical to one another.
There is another, perhaps simpler, way to analyze the behavior of the cascaded compressor 199. The compression produced by a rotary thermodynamic compressor of the form illustrated in FIG. 5, depends upon the input pressure for the device, assuming that all other operating parameters are held constant. This approximation applies to the case where the flow of working fluid through system is not so great as to render heat exchange within the heat exchange means relatively ineffective. This proportionality between the compression in a single stage and the input pressure to that stage is a consequence of the production of pressure by the action of centrifugal forces on the columns of gas within tube sections 65 and 62 in FIG. 5. The total pressure in the forward direction, as designated by arrow 54, produced by the column of gas in tube section 65 depends upon the centrifugal force acting upon the mass of gas in that tube section. The reverse pressure produced by working fluid in tube section 62, trying to force working fluid against the direction indicated by arrow 54, also depends upon the action of centrifugal forces on the mass of the working fluid present within column 62. The difference in density between the working fluid in the two columns, 65 and 62, in a consequence of a difference in temperature within those two columns. For an ideal gas the ratio of the density in one columm 65 to the density in the other column 62 depends upon the ratio of the temperatures. For a fixed relationship in temperatures the absolute difference in density between the working fluid in the two columns is proportional the the absolute density of the working fluid. This density, in turn, depends upon the overall pressure of the working fluid within the system, and, for relatively small flow rates, is explicitly a single-valued function of the pressure at the inlet to the compression stage. Thus, for an ideal gas, the absolute difference in pressure produced by the operation of a single-stage of a rotary thermodynamic compressor is proportional to the pressure at its inlet. This behavior for a cascaded compressor is illustrated in FIG. 11. The pressure difference C.sub.2 produced in the second stage of compression, is not as large as the pressure difference C.sub.3 produced in the third stage. This is because the inlet pressure at the intake to the second stage of compression is not as high as the inlet pressure at the third state, In FIG. 11, operation is assumed to be with a working fluid which is an ideal gas, and the pressure increment for each stage of compression after the first stage is roughly proportional to the inlet pressure for that stage. One consequence of this physical effect is that operation of a stage of a rotary thermodynamic compressor with many stages can be characterized as a multiplication of the inlet pressure by a ratio, which, for flows not so large as to render the operation of the heat exchange means within the system relatively ineffective, nor so large as to cause appreciable friction, is independent of both inlet pressure and flow. This leads to an exponential dependence of the form shown in the following equation:
P.sub.out .congruent.P.sub.in R.sub.p.sup.N
in which P.sub.out is the outlet pressure, P.sub.in is the inlet pressure, R.sub.p is the compression ratio for each stage and N is the number of stages in cascade.
Cascading rotary thermodynamic compressors of this form is a way to achieve capability of delivering working fluid at a higher pressure than would otherwise be possible. This increases the resistance to reverse flow through the compressor, and thus increases the impedance of a load to which such a compressor system can stably deliver working fluid. Moreover, because the output pressure increases exponentially with the number of cascaded stages, cascading the stages results in a greater total impedance than the sum of the individual impedances of each stage operating along, and thus stabilizes the compressor substantially more effectively than might be considered to be possible.
One modification of the compressor 199 can be formed by using several parallel branches, each of which contains several stages in cascade, in order to deliver a larger volume of working fluid, and in order to provide flexibility in the geometric arrangement of the various pumping stages within the device. For instance, such parallel branches can be used to provide for dynamic balance of the system when it is working into various gas pressure loads.
FIGS. 12 and 13 show a cascaded multi-stage thermodynamic compressor 500 with stages like those shown in FIG. 10, but arranged in a particularly advantageous formation.
As is shown schematically in FIG. 12, the compressor 500 includes two groups of loops 510 and 512 of tubing. Each group of loops is formed by winding a single length of tubing in a pattern tending to form a toroid. Each loop 510 is opposite to a loop 512 in the opposite group, and the loops are arranged symmetrically with respect to the central axis 517 of the toroid.
The starting end of the upper group of loops 510 is connected to the starting end of the opposite group 512. This connection is indicated by reference numeral 516. Similarly, the trailing ends of the groups are connected together as indicated at 518. Thus, the two groups are connected together in parallel. A refrigeration unit or other load 519 is connected to the conduits 516 and 518. The refrigeration unit 519 contains, for example, means of the type described above for centrifugally compressing, expanding and returning a working fluid to the compressor 500 through the conduit 516. The compressor 500 and the refrigeration unit 519 are connected together to be rotated as a rotary heat pump unit by a motor 504.
As is shown in FIG. 13, the loops 510 and 512 are secured between a pair of heat-conducting metal plates 506 and 508 by means of welding or soldering. The plates 506 and 508 are secured to a hollow shaft 502 through the center of which pass tubes 516 and 518. Insulation 514 fills the toroidal hole formed by the loops 510 and 512. The plates 506 and 508 may have suitable heat transfer fins on their outer surfaces.
The various compression stages are arranged so that all of the heat exchangers 135, 139, etc., through which heat is rejected contact the plate 508. Heat is conducted into plate 506 from the working fluid, and is dissipated from plate 508 into the environment. Similarly, heat exchange means 133, 137 and 141, through which heat is absorbed into the working fluid during expansion, make thermal contact with the plate 506 through which heat flows into the working fluid.
The compressor 500 operates as follows: Heat is added to the portions of the loops in which the working fluid flows towards the axis 517 by heating the plate 506, and the portions of the loops in which the fluid flows away from the axis 517 are cooled by cooling the plate 508. Rotation of the loops augments the pressure difference between the outwardly and inwardly flowing fluid columns in each loop in the manner discussed above. Since the loops in each group are connected togehter in series, the compression produced by each loop multiplies that produced by the preceeding loops in the group, with the result that relatively high total fluid pressures can be produced with working fluids of relatively low density, or with the use of relatively low rotational speeds, or with rotary devices having relatively small diameters. Alternatively, rather than using this embodiment of the invention to reduce the foregoing parameters, it can be used simply to produce very high total fluid pressures.
The arrangement of the loops into two parallel-connected groups is made in order to ensure that opposite portions of the rotary structure will have the same amounts of fluid in them at the same time and the rotational balance of the structure will be maintained. Additional parallel-connected groups can be added as desired.
All of the rotary thermodynamic devices discussed so far have in common the same physical principle of operation. This is true independent of whether the system is used for cooling or for compression, whether the system has a single branch through which fluid can pass or multiple branches in parallel, whether the system has a single stage or a series of cascaded stages, whether the system is part of a closed rotating loop, or is open in the sense that working fluid enters and leaves the rotating assembly. The principle of operation which all of these devices have in common is the interaction within a rotating system of inertial forces which arise within the rotating system and the thermodynamic properties of a working fluid. These inertial forces are known as centrifugal forces and coriolis forces. The centrifugal forces are the familiar forces which tend to throw material out toward the rim of a spinning chamber. The coriolis forces are those which act upon material moving outwardly in a duct to bring it up to speed so that its tangential velocity about the axis matches that of the channel within which it is moving. Similarly, when material is moving from near the periphery to near the axis, coriolis forces act to slow down the material so that when it reaches the axis its tangential velocity has been reduced from that which it had near the periphery. It is the interaction of these rotary inertial forces with differences in density of the working fluid, associated with differences in temperature, which link thermodynamic work in the form of the flow of heat to thermodynamic work in the form of flow of a pressurized working fluid within these systems. It is this relationship between thermodynamic flows of heat and mechanical flows of working fluid which gives rise to the characteristic dynamic properties discussed above. The dynamic instabilities which have been discussed are, therefore, a characteristic property of rotary inertial thermodynamic devices. These instabilities arise when there is an improper relationship between the thermodynamic impedances and mechanical impedances for the various parts of the thermodynamic device and the other parts of the system, of which it is a component. The mechanical impedance to the flow of working fluid within a rotary inertial thermodynamic device may be regarded as a property of the device itself and the external flow impedance to which it is coupled. The thermodynamic impedances presented to working fluid within the system include both the thermodynamic impedances for exchange of heat within the device itself and also the thermodynamic impedances external to the flow of the working fluid proper. All of these thermodynamic impedances should be considered in determining the stability of flow of working fluid within the rotary inertial thermodynamic device.
For example, suppose that the cooling device diagrammed in FIG. 4 had in tube section 3 a very efficient heat exchange means for allowing heat to flow from the working fluid during its compression into the heat exchange means itself. Suppose, however, that this heat exchange means was only relatively ineffectually linked to an external sink to which this heat could be dissipated. The ability of the heat exchange means to remove heat from the working fluid during compression would then depend, not only upon the effectiveness with which the heat in the working fluid could be exchanged with the heat exchange means itself, but also upon the effectiveness with which this heat exchange means could dissipate the heat to some other part of the system. This total thermodynamic impedance is what characterizes the thermodynamic impedance presented to the working fluid. If this total thermodynamic impedance is very high, even if there is only a relatively small flow, the system will not be capable of dissipating the heat of compression from the working fluid and will have a rapid drop in the back-pressure by which it acts to use the pressure of the working fluid entering the system to produce cooling. If, on the other hand, the total thermodynamic impedance with respect to the working fluid in tube section 3 is very small, even when there is a relatively large flow of working fluid a cooling effect can be expected.
The characteristic thermodynamic impedance of the heat source is another impedance of the system which should be taken into consideration in stabilizing a rotary inertial thermodynamic system. The impedance of a heat source is analogous, in some respects, to the internal impedance of a source of electrical energy.
In this analogy, heat flow corresponds to electrical current, and temperature corresponds to voltage. Thus, a high-impedance heat source is one in which the heat flow is relatively constant regardless of the temperature of the medium into which it delivers heat. Conversely, a low-impedance heat source is one in which the temperature is relatively constant regardless of the amount of heat flow in to the medium. Hence, a high-impedance heat source is analogous to a constant-current electrical source, and a low-impedance heat source is analogous to a constant-voltage electrical source.
Examples of high-impedance heat source are flames and hot air. An example of a low-impedance heat source is a relatively large body of hot water. Other examples of both types of heat sources will be given below.
As an example of the influence of the impedance of the heat source on the stability of a rotary inertial thermodynamic device, consider the device shown in FIG. 5. The heat exchange means 46 in the tube section 51 is coupled to an external source of heat which has, of course, a characteristic thermodynamic impedance. If the impedance of the heat source is very high, as the flow rate of working fluid through the compressor 52 decreases because of an increasing backpressure against which the system must deliver working fluid, the amount of flow of working fluid past heat exchange means 46 available to take heat away from it decreases, and, therefore, the temperature of heat exchange means 46 increases. The result of this increase in temperature of heat exchange means 46 is that the density of working fluid in tube section 51 decreases, increasing the effective compression available from the rotary inertial thermodynamic system. Thus, the use of a heat source with high thermodynamic impedance tends to stabilize the operation of the compressor in the presence of large back pressures from external loads in that the higher compression enables the device to better resist reversal of flow from the load back through the compressor.
The principles discussed in the preceeding section can be utilized to ensure the stable operation of a wide variety of rotary inertial thermodynamic systems, including a closed-loop rotary inertial thermodynamic system 300 of the form described in my U.S. Pat. No. 3,470,704 and shown schematically in FIG. 14. The device 300 has a compressor section 311, and a cooling section 312. The compressor section includes conduit sections 301 and 302, heat exchange means 303, and an expansion region 304 within conduit section 302. Compressor 311 is a rotary inertial thermodynamic compressor. The cooler section 312 includes conduit segments 305 and 308, and heat exchangers 306 and 309. Heat exchangers 306 and 309 extend for a substantially the full lengths of conduit sections 306 and 309. Cooler 312 is a rotary inertial thermodynamic cooler. Conduit section 310 completes the closed loop conduit.
The operation of rotary inertial thermodynamic compressor 311 is characterized by FIG. 15 which is a graph relating the change in pressure in the compressor and the cooler to the rate of flow of working fluid through the system. Note the sign convention for pressure change, which leads to a representation of P of the compressor being negative, i.e., the drop in pressure of working fluid flowing through it is negative; the pressure produced is positive.
The properties of the cooler 312 are represented by curve 313 in FIG. 15. It can be seen that P for the cooler decreases at first with increasing flow, as the thermodynamic back-pressure decreases. At relatively high flow rates, mechanical impedances dominate and the pressure drop through the system again rises. Note that the rate of drop of curve 313 in FIG. 15, representing the back-pressure generated in the cooler is a consequence of a loss of heat from the working fluid during compression in tube section 305 by means of heat exchange means 306, and the gaining of heat from the environment during expansion of the working fluid in tube section 308 by means of heat exchange means 309.
If the thermodynamic coupling of heat exchange means 306 and 309 to their environments is very poor, then the rate of drop of the back-pressure generated in cooler 312 would be much steeper than that shown in curve 313. The relationship of back-pressure to flow for this condition to their respective environments is represented by dashed curve 315 in FIG. 15.
FIG. 16 shows a curve 317 relating the total pressure drop through both the compressor 311 and the cooler 312, acting in series, to flow rate F. Stable operation of the system as a closed loop occurs when the total pressure drop in going aroung the loop is zero, and the slope dP/dF is positive. Curve 317 passes through zero total pressure drop at an operating point 216. Flow in the system at point 316 is stable and efficient. Limitations on the flow of the working fluid arise primarily from thermodynamic effects, rather than from mechanical, frictional constraints.
Curve 318 in FIG. 16 represents the total pressure drop in compressor 311 and cooler 312 in the case in which cooler 312 has only very little thermodynamic coupling to its environment. That is, curve 318 represents total pressure drop for the same conditions represented by curve 315 of FIG. 15. Curve 318 represents the algebraic sum of curves 315 and 314. Curve 318 crosses the zero axis at an operating point 319. This point represents a condition in which the working fluid is circulating very rapidly through the system and the amount of work done by the working fluid against the thermodynamic pressure drop within cooler 312 is very small. The principal limitations on the flow are caused by friction.
The efficient operation of the rotary inertial thermodynamic device 300 as a heat-actuated cooling system requires that the flow of working fluid within the device be limited principally by thermodynamic effects rather than by mechanical friction of the working fluid within the conduits and heat exchangers through which it passes. The reason for this is that the mechanical friction on the working fluid is an irreversible thermodynamic loss. Thus, to achieve stable, efficient operation of the device 300, it is desirable that the cooler 312, regarded as a total system (including those parts of its environment with which it exchanges heat) present a lower thermodynamic impedance than is presented by the compressor 311, regarded as a total system (including those parts of its environment with which it exchanges heat). If the foregoing constraints on thermodynamic impedance cannot readily be met, the operation of the system can be stabilized by inclusion, anywhere within the conduit, of a flow restricting means 320 (FIG. 14), for instance a constriction or a porous plug in the conduit. This flow restrictor can advantageously be combined with one of the various heat exchange means present within the conduit, although it is not necessary to make such a combination.
As a second example, consider the problem of the stability of an "open loop" rotary inertial thermodynamic system; e.g., a system of the type shown in FIG. 4 in which a stationary source of working fluid is used. A parallel-branch embodiment 349 of such a system is shown schematically in FIG. 17, in which 360 is the inlet, and 361 is the outlet. The device has four branches 350, 351, 352 and 353, containing, respectively, compression sections 348, 357, 358 and 359, which contain, respectively, heat exchangers 347, 354, 355 and 356. These heat exchange means are coupled to an external environment, into which they can reject the heat of compression which they receive from the working fluid as it is compressed.
FIG. 18 shows a curve 362 relating the frictional (mechanical) pressure drop of working fluid in passing through one of the branches (which are assumed to be identical) to the flow F of working fluid through that branch, and a curve 363 relating the thermodynamic pressure drop to the same flow. In FIG. 19 the curve 364 relates the total pressure drop within each branch to flow. Curve 364 is the algebraic sum of curves 362 and 363 in FIG. 18. An operating point 365 is selected at the point of tangency of a tangent line 366 which has a positive slope.
By making the mechanical impedance in each one of the branches sufficiently large, the pressure drop within that branch can be dominated by the mechanical impedance, rather than by the thermodynamic impedance. In this way, it is possible to select an operating point where the curve 364 has a positive slope, i.e., a small increase in the flow through that branch would be accompanied by an increase in the pressure drop within the branch. This means that the flow through the branch would decrease. Similarly, any decrease in the flow through that branch would lead to a decrease in the pressure drop, thus allowing the flow to increase. Therefore, the flow at the operating point 365 is stable.
The mechanical impedance can be simply the impedance of the heat exchanger in that branch. Separate flow restrictors also can be used. The high mechanical impedance necessary for stability causes irreversible thermodynamic losses. For this reason, although the multiple-branch cooler 349 shown in FIG. 17 might appear attractive, a detailed analysis, including an analysis of possible dynamic instabilities, shows that its thermodynamic efficiency is not nearly as high as in alternative embodiments disclosed herein.
The length and positions of the heat exchangers in the conduits is a factor to be considered in the construction of rotary inertial thermodynamic devices. Consider the case where the rotary inertial thermodynamic device in FIG. 4 is operated as a single-branch compressor of the type in FIG. 5 with heat supplied to heat exchange means 30. Consider first the case where heat exchangers 29 and 30 are relatively short, and exchanger 29 is placed near the axis 1, and exchanger 30 is remote from the axis. For this discussion, it is assumed that fluid in the rest of the system is relatively cool Assume also that flow is opposite to the arrow 31. This causes region 4 to be filled with cool working fluid, and region 32 to be filled with the working fluid which has been heated by its passage through heat exchanger 30. The result is that, even for small reverse flows (in a direction opposite to arrow 31), the system ceases to operate as a compressor driving working fluid in the forward direction and begins to drive working fluid in the reverse direction.
FIG. 20 is a graph illustrating the operation as a compressor of the device shown in FIG. 4 with varying lengths of the heat exchangers 29 and 30. Frictional effects are included. Curve 409 represents flow in the forward direction indicated by arrow 31. Curve 412 represents reverse flow with short heat exchangers positioned as described above. Curve 411 represents conditions identical to those of curve 412, except that the heat exchangers are longer, and curve 410 represents the case in which the exchangers are so long that they substantially fill the tube sections in which they are located.
The change from forward to reverse flow results in a sudden change G.sub.1 or G.sub.2 in the compression available. As the heat exchangers are elongated to fill progressively larger portions of the conduit segments, the effect produced by changing the temperature of the working fluid in sections 32 and 4 becomes smaller. For small flows, it is assumed that the working fluid within heat exchangers is essentially the temperature of the heat exchanger. Therefore, the change in pressure appearing for small flows in the reverse direction is reduced by extending the length of these heat exchangers. Thus, with the longest heat exchangers, the change in pressure appearing upon reversal of flow through the system is essentially zero. However, extending the length of the heat exchangers beyond the length required to produce the isothermal compression and expansion required for operation in the Carnot cycle reduces the efficiency of the system. This is because the region 32 for adiabatic compression and the region 4 for adiabatic expansion become very small. This does not allow adequate compression to occur in section 32 to allow the working fluid to achieve at station 15 a temperature equal to that of heat exchanger 30. In the case of reduction of length of adiabatic region 4, the result is that working fluid leaving the system through outlet 34 is at a higher temperature because it has had less adiabatic expansion to reduce its temperature from that which it possessed upon leaving heat exchanger 30. The result is that the thermodynamic efficiency of the system is decreased at the same time that the gap in pressure upon small reversal of flow of working fluid through the system is decreased.
Dashed curve 417 in FIG. 20 represents pressure variations with small flows in the forward direction in a device as shown in FIG. 4 used as a compressor in which the heat exchange means has a relatively high thermodynamic impedance to the flow of heat from the external heat source into the working fluid. This high thermodynamic impedance can be caused either by the high internal impedance of the source itself, or a high thermodynamic impedance to the flow of heat into the working fluid, or by both. The result of this high impedance is that, for small flows of working fluid forward through the compressing system, the temperature of heat exchanger 30 increases so that the amount of compression available, due to the difference in density of the fluid in the outwardly and inwardly directed segments of the conduit 11, increases.
Systems represented by the graph 409 in FIG. 20 can be said to be conditionally stable, i.e., if they are operated at an operating point 420 for which the flow in the forward direction is sufficiently different from zero so that fluctuations in flow are very unlikely to drive the flow into the reverse mode, then the system will operate properly as a compressor, driving working fluid in the forward direction. Such compression systems can be made unconditionally stable, i.e., stable independent of whether they are forced in a reverse mode or not, by coupling them with mechanical impedance means (e.g., fluid traps disclosed in my above-identified pending patent application and herein below) providing a relationship between flow and pressure as indicated graphically in FIG. 21. In FIG. 21, curve 415 represents flow in the forward direction through this impedance, curve 416 represents flow in the reverse direction through the impedance, and the gap G.sub.3 in the pressure in the region of zero flow represents the change in pressure required to force this impedance into the reverse flow mode. As long as gap G.sub.3 is larger than the gap, G.sub.1 or G.sub.2 (FIG. 20), appearing when the compressor is forced into reverse flow at small flow levels, the system will be unconditionally stable in the vicinity of zero flow. Operation in this mode is illustrated graphically in FIG. 21 by curve 413 (forward mode) and curve 414 (reverse mode). Gap G.sub.4 represents Gap G.sub.3 algebraically summed with gap G.sub.2, and is the amount of pressure required, beyond the back-pressure at which the compressor has zero flow, the force flow backwards through the compressor. Note that at no point does curve 414 enter the second quadrant of the graph, which would represent pumping of working fluid in the reverse direction.
Also, there is a gap G.sub.5 between the greatest forward compression and the least reverse flow pressure. This ensures that a set of parallel branches of a compressor can operate together without some branches forcing working fluid back through others.
Although the basic physical nature is the same, there is a useful difference between the form of instability which occurs in a gaseous rotary inertial thermodynamic cooler containing a set of branches connected in parallel and that which can occur in a rotary inertial thermodynamic compressor containing a set of parallel branches. In the case of the cooler, the instability is believed to consist of an excessively rapid flow of working fluid through one of the branches of the system. In this type of instability, working fluid flows through the branch in the desired direction. In the case of the compressor, the form of the instability is believed to be that the flow through one branch of the compressor is reversed. Therefore, there are a number of techniques which can be used in the compressor to avoid the reversal of flow and thereby make the system unconditionally stable which are not available for use in stabilizing the cooler. This is made especially clear by FIG. 21.
One of the simplest devices which will produce the assymmetric behavior with respect to flow, represented by FIG. 21, is a check valve. For example, this could be a flap valve or ball valve which opens to allow flow in the forward direction and closes to prevent flow in the reverse direction. Alternatively, it could be a liquid trap as disclosed and as is shown in FIG. 22 of the drawings herein.
FIG. 22 shows an impedance control means 420 with an outwardly-extending conduit section 421, an inwardly-extending section 424, and a broad chamber 422 between sections 421 and 424. The chamber 422 has an inlet port 427. The liquid 423 is held against the outer wall 425 of the chamber 422 by centrifugal force caused by rotation of the device about the axis 1. For a gaseous working fluid to flow in the direction indicated by arrow 428, it need have only enough pressure to bubble up through the shallow liquid 423 in chamber 422 and then out through exit conduit section 424. For the gas to pass through the trap in the opposite (reverse) direction, the gas must push liquid 423 back up into conduit section 421 a substantially greater radial distance than it must push the liquid in order to flow in the dorward direction. This introduces a pressure gap corresponding to G.sub.3 in FIG. 21, and, in effect, forms a type of check valve. Branches identical to the one shown in FIG. 22 can be connected in parallel if desired.
FIG. 23 shows a compressor 430 like that shown in FIG. 5, except that it has plural parallel branches instead of one, a plenum 431, and a flap valve 432 near the outlet of each branch. This compression system is unconditionally stable against excessive back-pressure at outlet 49. Additionally branches, each with its own check valve 432, can be added as desired.
It is possible to use several forms of stabilization simultaneously in the same device. For instance, the heat exchangers might be coupled to high-impedance heat sources so as to enhance the protection against reversal of flow. Examples of high-impedance heat sources which might be utilized in this way are, in addition to a flame burning a fixed amount of fuel per unit time, the decay of a radioistope heat source, radiant heating, electromagnetic inductive heating, etc. In another embodiment, the same electromagnetic induction field which transfers thermal energy to the heat exchangers might also provide a rotating electromagnetic field which by its electromagnetic drag on the rotating system, rotates the device about the axis 1. For practical reasons, generally it is preferable that this high impedance be achieved by having a high thermodynamic impedance between the environment and the heat exchange means, rather than by a high impedance between the heat exchange means and the working fluid itself. This insures that the heat exchange means within the compressor will not have a temperature very much higher than that of the working fluid and tends to protect the working fluid from thermal degradation.
Use of an unconditionally stabilized rotary inertial thermodynamic compressor with multiple branches facilitates the construction of a system in which the same compressor is capable of providing a small amount of flow into a very high back pressure and/or a large amount of flow into a low back pressure. For instance, one of the branches in a multiple branch system can have a large number of rotary inertial thermodynamic compression stages cascaded. This allows it to produce a very high pressure at its delivery outlet. This would be delivered through an appropriate check valve into the output plenum. In the face of such back pressure, the other branches in the system would have zero flow. Flow would not go backwards through them because of their check valves. For this reason, they would not |