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Publication numberUS3698182 A
Publication typeGrant
Publication dateOct 17, 1972
Filing dateFeb 16, 1971
Priority dateSep 16, 1970
Also published asDE2109891A1, DE2109891B2, DE2109891C3
Publication numberUS 3698182 A, US 3698182A, US-A-3698182, US3698182 A, US3698182A
InventorsKnoos Stellan
Original AssigneeKnoeoes Stellan
Export CitationBiBTeX, EndNote, RefMan
External Links: USPTO, USPTO Assignment, Espacenet
Method and device for hot gas engine or gas refrigeration machine
US 3698182 A
Abstract
A closed-cycle gas expansion engine and a closed-cycle or open-cycle gas expansion refrigerator, including a thermal regenerator, devices for transferring heat into the gas or subtracting cold gas, and a piston assembly for developing approximately 180 DEG phase difference in the varying volumes of the warmer and colder gas chambers. A general valve mechanism is employed for properly timed flow of working gas into the primary gas chamber from a plenum chamber containing cooling devices for closed-cycle operation, or for flow of working gas into the primary gas chamber from an ambient atmosphere for open-cycle operation. A valve mechanism, separate or coupled to the first mechanism, is used for proper initiation and termination of gas flow between the primary and secondary chambers in one particular phase of the reciprocating piston motion. Still another valve, which may be integrated with any of the other mentioned valve mechanisms, is used for release of suitably recompressed working gas in the secondary chamber into the plenum chamber with a closed cycle, or for properly timed exhaust of such gas back to the ambient atmosphere with an open cycle.
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Description  (OCR text may contain errors)

United States Patent Kniiiis [451 Oct. 17, 1972 [54] METHOD AND DEVICE FOR HOT GAS [57] ABSTRACT ENGINE 0R GAS REFRIGERATION A closed-cycle gas expansion engine and a closed- MACHINE cycle or open-cycle gas expansion refrigerator, includ- [72] Inventor: Stellan Kniiiis, Fregattvagen 2, ing a thermal regenerator, devices for transferring T b S d heat into the gas or subtracting cold gas, and a piston assembly for deve oping approximately 180 phase dif- [22] led: 1971 ference in the varying volumes of the warmer and 21 A N 115,547 colder gas chambers. A general valve mechanism is employed for properly timed flow of working gas into the primary gas chamber from a plenum chamber con- [30] Forelgn Apphcatlo P'mmy Data taining cooling devices for closed-cycle operation, or March 2, 1970 Sweden ..l2640/70 for flow of Working gas into the P y gas Chamber from an ambient atmosphere for open-cycle opera- 52 us. Cl ..60/24, 62/6 tion- A valve mechanism, separate of coupled to the [51] Int. Cl ..F0lk 7/06 first mechanism is used for Proper initiation and [58] Field of Search ..60/24; 62/6 minati gas between the Primary and dary chambers in one particular phase of the l 56] Reierences Cited reciprocating piston motion. Still. another valve, which may be integrated with any of the other mentioned UNITED STATES PATENTS valve mechanisms, is used for release of suitably recompressed working gas in the secondary chamber 33 into the plenum chamber with a closed cycle, or for properly timed exhaust of such gas back to the am- 2,7s4,54s 3/1957 Fiala ..60/24 his, atmosphere with an open cycle 3,400,281 9/1968 Malik ..60/24 X a 3,460,344 8/1969 Johnson ..60/24 3,552,l20 l/l97l Beale ..60/24 Primary Examiner-Martin P. Schwadron Assistant Examiner-A. M. Ostrager Attorney-Sokolski & Wohlgemuth 35 Cla m-S, Draw Eire-" 48 so 36 H E AT H/ 46 1T 38 EXCHANGER BURNER 44 PLE NUM 42 C OOLE R 40 HEAT THERMAL v 54 EXCHANGER REGENERATOR 30 2a 5s v 58 2o Q 26 PATENTEUUBHTIQTZ 3.698.182

SHEET 2 OF 9 TEMPERATURE E NTROPY (,0)

FIG. 2

COOLER HEAT v 'V EXCHANGER 30b HEAT H EXCHANGER" THERMAL J H REGENERATOR THERMAL 28b REGENE RATOR 28a FIG.5

INMENTOR STELLAN KNOOS somsxl a WOHLGEMUTH ATTORNEYS PAIENTEDnm 17 I972 3. 6 98. 18 2 sum 3 or 9 PLENUM COOLER HEAT THERMAL I 56 EXCHANGER REGENERATOR t j as |2 If. .1 l8 I60. I an l6b FIG.4 J 22 FROM AMBIENT TO AIMBIENT REGENERATOR TO EN VOLUME lGb FIG. I I INVENTOR STELLAN Kubs SOKOLSKI a WOHLGEMUTH ATTORNEYS PATENTED BT 17 1 3,698,182

saw 1; [1F 9 HEATER FIG. 6

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20s O v M REGENERATOR REGENERATOR l [L 2I4 FIG. I4 I I I TO AMBIENT 1 I74 ATMOSPHERE REGEN- /2 FROM ANBIENT ATMOSPHERE I 9 28 mg I86 I REGEN.

L II I,

INVENTOR 224 STELLAN KNbs SOKOLSKI GI WOHLGEMJTH ATTORNEYS PATENTEDHCI 1 I912 3.698. 182

SHEET 9 [1F 9 FRQM A I T ATMOSPHERE 28\ T PRIMARY V 56 A THERMAL CHAMBER REGENERATOR FIG. |5b CHAMBER .HEAI 22s EXCHANGER 23o FIG. I56

FIG. 15 I 228 m \THERMAL o REGENERATOR 22a m INVENTOR STELLAN KN66S SJKOLSKI 8| WCHL'GEMUTH ATTORNEYS METHOD AND DEVICE FOR HOT GAS ENGINE OR GAS REFRIGERATION MACHINE BACKGROUND OF THE INVENTION of performing the desired thermodynamic cycles for the working gas in such devices.

Several arrangements are available for performing the Ericsson and the more well-known Stirling cycle for hot gas engines with external-combustion heat addition. The most recent of the Stirling engines employ two pistons reciprocating in a single cylinder approximately 90 out of phase with each other. Stirling engines of this kind are receiving much attention, particularly due to their potentially low-pollution exhaust gas emission, but also due to high thermal efficiency and low noise. They further have the advantage of capability of use with a great number of different liquid and gaseous fuels. The Stirling engines are necessarily equipped with thermal regenerators in order to obtain a high thermodynamic efficiency. Characteristic of present Stirling-cycle engines is a requirement of liquid cooling. The working gas, which could be helium, hydrogen or less preferred air, rejects heat in a cooler placed in immediate connection with the thermal regenerator inthe gas line between the hot and colder work chambers. For optimum efficiency, the dead space of the cooler, as well as regenerator and heater dead-spaces should be minimum. A compromise generally has to be made between the cooler deadspace and the efficiency of the cooler, by keeping the dead-space moderately small but allowing the temperature difference between the liquid coolant and the working gas to be relatively large, and hence a lower overall temperature ratio and lower thermodynamic efficiency than could otherwise be attained. Contributing reasons why hot gas engines are not in widespread use today are large weight and high mechanical complexity of existing designs, and their estimated high production cost, e.g. as compared to Otto-cycle and Diesel-cycle engines with the same power output. Materials and sealing problems of the past have to a great extent been solved by modern technology, and no large technological barriers exits today to the practical use of hot gas engines, e.g. for automobile propulsion. However, the use of Stirling engines for automobiles has so far not been favored, primarily due to the projected high manufacturing cost of the engine, the relatively poor power/weight ratio, and practical difficulties to be experienced in handling high pressure helium and hydrogen working gases, e.g. in service and overhaul. With the more severe air-pollution problem on hand, the demand of a low-pollution engine is stronger than ever, and the Stirling engine or hot gas engines in general could provide the sought for solution if they could be made practicable.

Modern Stirling engine concepts and associated devices are described, for example, in the following U.S. Pats: No. 3,166,911 to Meijer; No. 3,442,079 to Meijer; No. 3,011,306 to Meijer; No. 3,036,427 to Meijer; No. 3,015,475 to Meijer; No. 3,472,037 to Kohler, and No. 3,458,994 to Heffner. Other hot gas engines are described in U.S. Pats. No. 3,183,662 to Korsgren, No. 3,460,344 to Johnson, No. 3,407,593 to Kelly, No. 3,138,918 to Baker, No. 3,174,276 to Baker, No. 3,385,051 to Kelly, No. 3,080,706 to Flynn and No. 2,564,363 to Horwitz. An interesting invention for a hot gas engine is described in U.S. Pat. No. 3,457,722 to Bush, where the cycle is interrupted during idle periods by the use of valves, and heating and cooling coils are connected into the system. Unfortunately, the device appears to be mechanically complex and uses valves operating in high-temperature environments.

Several gas-expansion refrigerator systems have been devised and built with the Stirling thermodynamic cycle. Such gas-expansion refrigerators operate with a closed thermodynamic cycle resembling the engine cycle, with the difference that mechanical work is added to the gas, heat transferred to the gas at a low temperature (the refrigeration load), and heat rejected at a higher temperature. Stirling-cycle refrigerators have come into widespread practical use, starting a few decades ago, while the corresponding engine still awaits exploitation. Stirling-cycle refrigerators are today capable of providing temperatures considerably below 100K, and a few systems have demonstrated refrigeration-load capabilities much above 10 kW at these temperatures. Measured overall efficiencies for such systems have been as high as 40 percent of the Carnot efficiency (the theoretical limit). Stirling-cycle refrigerators have been miniaturized and used for cooling, e.g., electro-optical devices. Temperatures below 20K in multiple-step operation have been demonstrated with helium as the working gas. The refrigeration-load capabilities for these devices have been typically of the order of 1 watt, with the total weight of the refrigerator system of the order of 10 kilograms. As for the Stirling hot-gas engines, the Stirling-cycle refrigerators of today are mechanically relatively complex, primarily due to their use of two pistons with phase difference therebetween. This makes for an expensive design, mitigating against widespread use of this type of refrigerator.

Among other refrigerators using gas-expansion principles with a reciprocating piston can be mentioned the Gifford-McMahon-cycle refrigerators. These devices are related to the Solvay refrigeration cycle, and employ a displacer piston with pneumatically controlled motion, and a separate high-pressure source of working gas, e.g. from a compressor. Irreversible expansion are characteristic for these refrigerators, and the associated entropy production lowers the thermodynamic efficiency to values considerably below the ideal Carnot efficiency. Therefore, the Gifford-McMahoncycle refrigerators are mostly being used in systems where thermodynamic efficiency and low power consumption are not of primary concern, but rather where reliability and maintenance-free operation are essential, e.g. for airborne and military applications. Refrigerators of this kind are described in the following U.S. Patents: No. 2,966,035 to Gifford; No. 2,906,101 to McMahon;No. 3,119,237 to Gifford; No. 3,188,819 to Hogan.

Other interesting gas-expansion refrigerators are the pulse tube refrigerator by Gifford, No. 3,237,421, and the heat-powered refrigerator of the Vuilleumier type, which has received attention, eg for space and airborne cryogenic applications for small refrigeration loads. A recent embodiment of the Vuillemier refrigerator is described in US. Pat. No. 3,423,948 to Cowan. A great number of gas-expansion refrigerators with counter-current heat exchangers exist as well (reverse Brayton cycle, Claude cycle, etc.). These devices are not of immediate interest with reference to the present invention, since they do not contain a thermal regenerator of the kind normally used in this invention. Only with reference to air dryers and heat pumps would counter-flow heat exchangers be of great interest relative to this invention. Of no immediate interest are Joule-Thomson refrigerators or common two-phase refrigerator systems using freons and other refrigerants, since the present invention is primarily a single-phase (gas) device.

SUMMARY OF THE INVENTION The present invention provides an improved and simple device for a hot gas engine or a gas-expansion refrigerator, using a new and simple method of performing the expansion step, and if such is desired, recompressing the working gas with a simple reciprocating piston assembly. Basically, simple versions of the invention employ two cylindrical work chambers of volumes which vary harmonically exactly or approximately 180 out of phase by virtue of a single reciprocating piston assembly, or several pistons. Working gas is stored in a plenum chamber for closedcycle operation. The gas is allowed to flow from the plenum chamber, or from the ambient atmosphere for open-cycle operation, into the so-called primary chamber, first passing through a valve arrangement of suitable kind, though a thermal regenerator, and finally through a heat exchanger (heat for the engine; refrigeration-load heat exchanger for the refrigerator). With cold-gas bleed for the open-cycle refrigerator the heat exchanger may be eliminated, and the heat load presented to the thermal regenerator by the discrepancy in mass of inflow and outflow through the regenerator. The gas is permitted to flow into the primary chamber during a controllable portion of the forward piston motion in which the volume of the primary chamber increases. In the reverse" stroke, gas from the primary chamber is permitted to flow through the mentioned heat exchanger and thermal regenerator, and through the mentioned valve mechanism, or through another separate valve, into the secondary chamber, where the volume is now increasing. This transfer of gas from the primary to secondary chambers is performed in a nearly reversible fashion, in contrast to e.g. Gifford-McMahon devices, by the very important fact that the flow into the secondary chamber is initiated when the volume of this chamber is zero (ideal case) or small, and hence no or little irreversible free expansions will take place. In the subsequent forward stroke the work gas trapped in the secondary chamber is recompressed to the pressure level of gas in the plenum chamber (closed-cycle operation), or to a suitable level (open-cycle operation) for exhaust. The suitably recompressed work gas is permitted to leave the secondary chamber through still another valve mechanism, which may be integrated with one or both of the first mentioned valves, and flows into the plenum chamber, or into the ambient atmosphere for opencycle operation. For closed-cycle operation, working gas is cooled preferably in the plenum chamber to the original temperature, and can thereafter be used to perform a new work cycle.

The mentioned valve mechanisms could be of various types, e.g., disk valves, sliding linear valves, or rotary valves. Proper synchronization between the valve action and the piston position (or shaft angle) could be achieved in many ways, e.g., with help of cams followers on the rotating shaft, or simply by placing a rotating valve or several rotating valves on the main rotating power shaft, or on a shaft that rotates with the same angular speed as the main shaft for the reciprocating piston motion.

In the transfer flow from the primary to the secondary chambers, the gas pressure should be lower than during injection into the primary chamber in order to obtain the correct thermodynamic overall cycle. This could be accomplished by a proper selection of crosssection area ratios of the primary and secondary chambers, and/or'proper termination of the flow into the primary chamber during the forward piston stroke. With injection during the full forward stroke for an engine, as a special case it is possible to employ an effective area ratio 1:1, which case could be of significant practical interest. For the refrigerator, an area ratio of 1:1 must always be accompanied by terminated injection during the forward stroke, at least for cases with no cold-gas bleed from the primary chamber. Secondary chambers with considerably smaller effective cross sectional area than that of the primary chamber can be used in the device of the invention is conjunction with injection only during a small fraction of the forward stroke, but this would be of practical interest primarily only to the engine.

The thermodynamic processes involved with devices of the invention are difficult to describe accurately, in particular since different portions of the work gas undergo different thermodynamic paths, e.g. when described in a temperature-entropy diagram. However, using a simple model for an average gas element it is easy to verify that the thermodynamic path is a loop of the desired kind for the closed cycle, with a calculable net amount of mechanical energy that can be extracted for the engine, and a refrigeration-load for the refrigerator. For moderate overall pressure ratios for the work cycle, e.g. ranging from 2 to 5, the computed thermodynamic efficiencies according to ideal models are very high, indicating a performance of devices of the invention that potentially could be brought close to that of an equivalent Carnot cycle. Realistic calculations show efficiencies which in fact are better than for common Stirling-cycle devices, primarily due to superior cooling characteristics and lack of cooler deadspace for devices of this invention. a

The mechanical output for the engine of the invention, and the cooling and load characteristics of the refrigerator can be controlled and regulated in a multitude of ways. The invention includes means for providing this control. A simple principle of regulation and control is to connect the plenum chamber (or another special chamber) with any or both of the work chambers during particular phases of the piston motion, and thereby controlling the pressure ratio of the cycle. The plenum chamber is then acting as deadspace during part of the cycle. The mentioned valve mechanisms can be used for this purpose, e.g. by using three-dimensional cams (with axial) movement) for a disc valve assembly. Alternative schemes could use a separate dead-space chamber or multitude of chambers that fully or stepwise could be connected to the work chambers continuously during the work cycle. As a special example of the first mentioned control technique, we mention termination of the gas flow from the plenum chamber into the primary chamber at an appropriate early position of the piston in its forward stroke. Hereby the amount of work gas entering the device each cycle can be controlled, and as a result thereof also the power output for the engine, or heatload characteristics of the refrigerator. Maximum power for the engine may be achieved if the injection is terminated, e.g. after the piston has performed threequarters of the forward stroke. Still another control technique is direct control of the pressure level in the plenum chamber. This could be done by help of hydraulic and pneumatic means, but could be power consuming and therefore is not a favored control technique.

The advantages of the invention will be readily understood from the description of the preferred embodiments thereof. However, we summarize obvious advantages of the closed-cycle engine of the invention in the following list.

1. Possibility of using only one piston or piston assembly, as compared with two pistons in common Stirling-cycle engines. This permits simple mechanical construction, low weight, and low production cost.

2. More efficient cooling system as compared to other hot gas engines. With cooling of the working gas taking place in a separate loop (in the plenum chamber), the residence time for a working gas element in the cooler can be made suitably large as can the heat-transfer area, and the heat-rejection fluxes conveniently low, in certain cases permitting the use of a gaseous coolant (e.g. ambient air). The cooling could therefore be more complete, and hence also the thermodynamic efficiency larger than for conventional Stirling engines.

3. The decoupled cooler does not introduce any dead-space effect, a fact which further increases the efficiency of such an engine.

4. No balancing buffer gas is needed, in contrast to recent Stirling engines employing rhombic-drive mechanisms and high pressures for the working gas. With this invention the power output means could be implemented with a relatively thin piston rod, or simply with the crank mechanism located between the two work chambers.

5. With air as coolant, heated air from the cooler could be used in the combustion process in the burner and heater. Reductions in weight and manufacturing cost can be made with such a scheme, in part from the fact that a single device (e.g. impeller) can be used for propelling this air both through the cooler and through the heater.

6. Possibility of locating the cooler and plenum chamber remote from the work chambers. This could be a most important feature for engines used in automobiles. Possibility of using a filter in the plenum chamber.

7. Simple and efficient control and regulating possibilities.

8. Multiple-cylinder arrangements can use a common cooler and plenum chamber.

9. Multiple-cylinder arrangements can be built mechanically simple with a high power/weight ratio.

10. For certain applications air could be used as working gas. Sealing problems would then be less severe, since gas lost by leakage could simply be replaced from the ambient atmosphere. In common Stirling engines air is a less favored working gas, in part due to difficulties in cooling the gas in the cooler (air has much smaller thermal conductivity than helium or hydrogen).

Advantages of the refrigerators of the invention (closed or open cycles) are summarized as follows 1 1. Highly efficient cooling system. Superior cooling performance is of primary importance in obtaining low temperatures and refrigeration-load/power ratios.

12. Highly reversible flow. Refrigerators of the invention can be built to give a minimum of irreversible entropy production, resulting in high overall efficiency.

13. Multiple-step staging possible with simple means. Two or more stages can be used for generation of very low temperatures, without addition of valves, and with a simple extension of the piston assembly to include additional work portions.

14. Open-cycle refrigerators can be built with extreme simplicity, without the use of heat exchangers other than the thermal regenerator. Such devices could be used for heat-pump applications and for air-conditioning systems. With modifications they could be used as air dryers, or devices for producing cryogenic liquids. The heat-pump application would be most interesting, since the coefficient of performance and hence the economy would be much better thanexisting systems using two-phase refrigerator systems with heat exchangers.

BRIEF DESCRIPTION OF THE DRAWINGS FIG. 1 is a schematic view of a first embodiment of a hot gas engine according to the invention, with a combustion heat source;

FIG. 2 is an idealized temperature-entropy diagram for a closed-cycle hot gas engine as shown in FIG. 1;

FIG. 3 is a diagrammatic view of another embodiment of the invention which is a two-step recompression hot gas engine;

FIG. 4 is a schematic view of a further embodiment of the invention which is a hot gas engine with its crank mechanism nearly perfectly balanced and located between the work chambers;

FIG. 5 is a schematic view of still another embodiment of the invention which is a double-acting hot gas engine with an oscillating shaft and vanes;

FIG. 5a is a schematic view of the output drive of the engine of FIG. 5;

FIG. 6 is a schematic view of another embodiment of the invention which is a four-cylinder hot gas engine;

FIGS. 7a -7d are diagrammatic views illustrating some ways of regulating the power output of the hot gas engine of the invention;

FIG. 8 is an idealized temperature-entropy diagram for a closed-cycle refrigerator of the invention;

FIG. 9 is a schematic view of still another embodiment of the invention which is an open-cycle refrigerator with cold-gas bleedoff;

FIG. 10 is a schematic view of a further embodiment of the invention which is a simple open-cycle refrigerator with cold-gas bleedoff;

FIG. 11 is a schematic view of another embodiment of the invention which is a refrigerator with phase difference between the two pistons operating in the primary and secondary chambers, respectively, and particularly adapted for refrigerators with large cold-gas bleedoff;

FIG. 12 is a schematic view of a further embodiment of the invention which is an open-cycle refrigerator with a primary counter-flow heat exchanger and two-step expansion, without recompression of the working gas;

FIG. 13 is a schematic view of another embodiment of the invention which is a two-step cooling device;

FIG. 14 is a schematic view of still another embodiment of the invention which is a three-step cooling device;

FIG. 15a is a schematic view of a further embodiment of the invention which is an air dryer based upon the open-cycle refrigerator principle, and employing centrifugal separation of the condensed liquid in the thermal regenerator; and

FIGS. 15b-l5e are schematic views of alternate (stationary) regenerator embodiments using counterflow heat-exchangers and valve additions, of particular use for air dryers and heat pumps, but also in particular cases for any of the other closed or open-cycle refrigerators or engines of the invention.

DESCRIPTION OF THE PREFERRED EMBODIMENTS (ENGINE) Referring now to the drawings, FIG. 1 discloses a hot gas engine generally indicated by numeral 10. The primary cylinder chamber is denoted 12 and the secondary chamber 14. The piston 16 has two cylindrical portions of generally different diameters, one cylindrical portion 16a moving in the primary chamber 12 formed by cylinder 18, and the other portion 16b moving in the secondary chamber 14 formed by cylinder 20. The piston is connected to rotatable shaft 22 for extraction of mechanical energy by means of connecting rod 24, and crank 26. The portion of cylinder between piston portions 16a and 16b is maintained at an arbitrary pressure level, e.g., the ambient pressure by means of port 20a formed in the cylinder wall. The shaft could be connected to a flywheel arrangement, if desired, that could store mechanical energy and maintain correct reciprocating piston motion even when energy from the gas is not added to the piston assembly.

A thermal regenerator 28 is connected in series with a heat exchanger 30 for addition of heat to the working gas. The regenerator could be a metal matrix, sintered material, a bed of pebbles, packed metal wire-mesh, etc., or a counterflow heat-exchanger with one-way channels (with additional valving to achieve one-way flow in each leg). The regenerator is used for temporary storage of heat of the oscillating working gas. For the engine the regenerator matrix receives thermal energy from the working gas leaving the primary chamber, and

adds the same amount (ideally) to gas flowing into the primary chamber; for the refrigerator to be described later, the regenerator receives thermal energy from gas entering the primary chamber, and returns the same amount of energy when the working gas leaves the same chamber. Hence, ideally the regenerator is in no thermal contact with any other material than the working gas.

Ambient air is fed into the heater assembly 31 (which includes burner 36, e.g. for a liquid petroleum fuel, and heat exchangers 30 and 34), from line 32, through heat exchanger 34, preferably of the counterflow type, and into the burner 36. The fuel is fed into the burner through line 38. The products of combustion in burner 36 are fed through line 40 into the,

heat exchanger 30, where heat is transferred to the working gas, which could be helium, hydrogen, or air, etc. FIG. 1 is only schematic, and in practice the burner and heat exchanger 30 could be integrated into a single unit. From the heat exchanger the combustion products are fed through line 42 back into the counterflow heat exchanger 34, where further heat is transferred to the incoming air, and finally exhausted through line 44. The counterflow heat-exchanger 34 could be replaced by a suitable thermal regenerator, e.g. by a rotating regenerator of metal or ceramic material.

The hot gas engine shown in FIG. 1 has a combined cooler and plenum chamber 46, in which working gas is cooled with help of a coolant circuit, and working gas can be stored at a convenient working pressure for the engine. The coolant (liquid or gaseous) flows into the cooler through line 48 and out of the cooler through line 50. The working gas enters the cooler 46 through line 52, and leaves the cooler through line 54. The volume of the Plenum-cooler 46 is preferably larger in comparison to the swept volume of the piston in the primary chamber. The advantage of a large plenumcooler is that it affords a large residence time for the gas in the plenum-cooler; this residence time could be made many times larger than the period of one work cycle. The heat fluxes through the cooler could then be made relatively small, and the efficiency of the cooler high, and in certain cases direct gaseous cooling (e.g., ambient air) could be used. With ambient air as the coolant, the coolant flowing from the cooler in line 50 may be fed directly into line 32 of the heater, with or without a certain amount of bleedoff, and used in the combustion process. The flow of coolant air can be sustained with the help of an impeller arrangement (not shown), that could be coupled to the rotating shaft 22 (or preferably geared to a higher rotational speed than provided by shaft 22). Pumping air coolant could alternatively be achieved by using the harmonically changing volume between pistons 16a and 16b in cylinder 20, and an added check valve.

The hot gas engine is shown with a general 3-port 2- way valve 56, and a 2-port l'way valve 58, for control of the flow of work gas through the engine. These valves could be combined into a single valve assembly, or further divided into e. g., a total of three 2-port l-w ay valves of a suitable type (disc valves, sliding valves, rotating valves, etc.). The valve action could be controlled in many different ways, but generally most simply by coupling to the rotating shaft 22. This coupling is indicated schematically in FIG. 1. The coupling could be eliminated if valve 58 is a simple check valve, or could include cams and cam followers if valve 56 is two disc valves, etc. The arrangement shown in FIG. 1 will suffice for thedescription of the operating principle of this type of engine which now.

follows.

Let us assume that quasi-steady conditions have been established, particularly with regard to the temperature profile in the regenerator. The right hand side of the regenerator 28 then has a temperature close to the temperature of gas in the plenum-cooler 46, and the left hand side of the regenerator has a temperature close to that of the heat exchanger 30, with a continuously changing temperature profile therebetween. Thermal conduction inside the regenerator (from left to right) should preferably be small compared to the heat flux through the heat exchanger 30. The work cycle can be described thermodynamically with the aid of FIG. 2 which is a strongly idealized and schematic temperatore-entropy diagram for an average gas element flowing through the closed-cycle engine. The volume of the plenum-cooler is assumed to be much larger than the swept volume of the piston, so that the flow of gas from the plenum-cooler 46 into the primary chamber 12 takes place at approximately constant pressure, as indicated by the straight line 60 (isobaric line) in the diagram which employs a logarithmic temperature scale and a linear entropy scale. The plenum-cooler exit condition is denoted 62.

The flow into chamber 12 is preferably initiated when the volume of that chamber is near its minimum, and is achieved by actuating valve 56 (FIG. 1), to connect the plenum-cooler to chamber 12 through regenerator 28 and heat exchanger 30. When the work ing gas flows from the plenum chamber through the regenerator, the gas temperature is raised by transfer of heat (with a small temperature difference between the gas and the regenerator working elements) from the regenerator. Working gas leaves the thermal regenerator 28 with a temperature and thermodynamic condition indicated by point 64 in FIG. 2. The flow of gas into the primary chamber 12 can be terminated well before the shaft 22 has rotated to a position at which the primary chamber has maximum volume. Depending upon the geometry and desired characteristics of the engine, the termination can take place, e.g., between 25 percent and 100 percent of the forward stroke, with an optimum termination point e.g. near three-quarters of the forward stroke. Valve 56 is then actuated to break the path from the plenum chamber to the primary chamber 12.

The path between the primary and secondary chamhers is preferably connected by means of valve 56 when the secondary chamber 14 has minimum volume, in order to avoid possible free expansions and flow irreversibilities. In the following reverse piston motion, the gas in chamber 12 is forced to flow back through the heat exchanger 30 and thermal regenerator 28, into the secondary chamber 14. During this transfer flow process the gas pressure is lower than the injection pressure represented by line 60. Depending upon the area ratio of the primary and secondary chambers and the temperature ratio across the thermal regenerator the pressure during the transfer process may decrease,

l0 stay constant or even increase. FIG. 2 relates to a preferred case with decreasing pressure. The isentropic expansion along line 66 to point 68 for the average gas element is then caused by the transfer process, with the element still located in the primary chamber, but could in part also be caused by an expansion following a terminated injection in the forward stroke.

When the average gas element enters the heater during the reverse stroke an isobaric heating process takes place, provided that the dead-space of the heater can be considered small and negligible (strong idealiza tion). The gas element leaves the heater at point 72 in the temperature-entropy diagram. In the regenerator, the gas transfers heat to the regenerator, as indicated by the isobaric line 74, the cooling process ending in point 76 at the exit of the regenerator.

In the present case with decreasing pressure during the transfer flow, the gas flowing into the secondary chamber mixes with slightly colder gas. This mixing process is generally accompanied by an overall increase in entropy, which is generally small and could be ignored in a first-order analysis. In FIG. 2 we neglect such a mixing process and show an isentropic line 78 for the further expansion of the average gas element in the chamber 14. The minimum pressure of the cycle is reached at point 80, when the volume of the secondary chamber is maximum. It should be noted that if the area ratio of the work chambers were chosen such that the gas pressure was increasing during the transfer flow (this process is not shown in FIG. 2), the minimum pressure of the cycle would have to occur at the end of the forward stroke for gas located in the primary chamber, having experienced an expansion following the terminated injection process.

In the following forward piston motion, the average gas element in the secondary chamber is compressed as indicated along line 82 to point 84, at which point the pressure is the same as in the plenum-cooler 46. Valve 58 is then opened (guided valve or check valve) from its closed position, in order to let gas flow through line 52 into the plenum-cooler 46 during the final portion of that piston stroke.

As shown in FIG. 2, working gas that enters the plenum-cooler has a higher temperature (point 84) than when leaving the plenum-cooler (point 62). The average gas element is cooled along the isobaric line 86 to the original condition 62, and the thermodynamic path is closed. A necessary condition for the averagegas model is that the regenerator has transferred zero net amount of heat to the gas, per cycle. Therefore, in a normal temperature-entropy diagram employing a linear temperature scale, the area a line corresponding to line 60 to the entropy axis must be identical with the area under a line corresponding to line 74 to the entropy axis, according to fundamental thermodynamic theories. In such a diagram (linear temperature axis), the net heat addition per unit mass. of the average gas in the heat exchanger 30 would be represented by the area under the corresponding line 70, and is generally smaller than the heat which is temporarily transferred to the gas from the regenerator, represented by the area under a line 60 in the modified temperatureentropy diagram. From this we can see that the regenerator is a significant element in the engine of the invention.

The thermodynamic process described in FIG. 2 is only approximate but can nevertheless be used for evaluating performance characteristics of the hot gas engine. Calculations with such a model, and also with much more accurate models, show that the anticipated thermodynamic performance of the device could in deed be high, and generally close to the Carnot-cycle performance. This is particularly true when the overall pressure ratio of the work cycle is kept moderate or small. For a practical engine the net power output per cycle is of importance, as well as efficiency. Therefore the overall pressure ratio cannot be kept close to unity as efficiency considerations would dictate, but rather between two and eight for a practical design.

It should be obvious with the simplified model of FIG. 2 that the thermodynamic processes involved in the hot gas engine of the invention are practically reversible in all separate steps. Entropy-producing free expansions are avoided, as well as throttled flows, since these have strong adverse effect upon the efficiency. With finite volumes of the regenerator, heater, and the channels from the primary to secondary chambers, and hence non-negligible dead-space, the valve actions could be modified to completely avoid free expansions. For example, valve 56 could be actuated to close the path between the regenerator 28 and the secondary chamber 14 before the piston has reached the left turning point in the reverse stroke; the closing could be such that gas in the regenerator, heater and primary chamber is recompressed to the plenum-cooler level in the very last portion of the reverse stroke, after which valve 56 is actuated to open the path between the plenum-cooler and the primary chamber.

FIG. 3 shows a hot gas engine employing recompression in two steps with the help of a third work chamber 88, formed by the piston and the cylinder 20. Besides the first plenum-cooler 46, a second plenum-cooler 90 could be used, operating at a higher pressure level. Additional valves are used, here shown as check valves 92 and 94. These could be valves of any suitable type and controlled not by the gas pressure as indicated, but by the shaft angle as for valve 56 (e.g. cam, cam follower and disc valve). In the embodiment of FIG. 3, gas in the secondary chamber 14 is forced into the plenum-cooler 46 through the valve 58 at minimum pressure loss, when the pressure in the secondary chamber is raised to a desired level (equal to the gas pressure in the plenum-cooler 46) in the forward stroke. During the same stroke, another portion of g as flows from the plenum-cooler 46 through valve 92 into the compression chamber 88. Final compression takes place in the reverse stroke in chamber 88, to the pressure level of the plenum-cooler 90. In the final part of that stroke, the valve 94 is opened and gas transferred from chamber 88 into the plenum-cooler 90, where the gas is cooled to its original temperature. The steady-state pressure level of the first plenum-cooler 46 will depend upon the relative sizes of chambers 14 and 88, as well as the cooling in plenum-cooler 46. Among advantages of two-step or multiple-step recompressions can be mentioned the possible reduction in the total amount of heat that has to be rejected in the coolers, for a given overall pressure ratio of the cycle, as well as a possible increase in the net power output and thermal efficiency due to less compression work necessary. Another ad vantage could be better balancing and distribution of angular moment on the output shaft 22. In practice, these features should have to be weighed against the added mechanical complexity with two-step recompression.

FIG. 4 shows another embodiment of my hot gas engine. The rotating output shaft 22 is here placed between the primary chamber 12 and the secondary chamber 14. For simplicity, we have chosen the same diameter for these chambers and piston portions 16a and 16b, making the design attractively simple. A bevelled gear 96 on the shaft 22 drives a pinion 98 and another bevelled gear 100, which is coaxial with the gear 96, and counter-rotating. Two connecting rods 24 are used to rotate the piston assembly 16. The rods are each fastened to a respective one of the counter-rotating gear wheels. Counter-weights 102 and 104 are shown schematically attached to each of the gears. With this arrangement a practically perfect balanced engine can be obtained, and with no significant side forces on the piston assembly. Such an absence of side forces is desirable in solving the sealing problem, and would permit an evenly distributed wear on piston rings, etc.

The valves 56 and 58 are here shown as rotating valves, located directly on the axis of gear 100. The valve action is the same as described for the engine in FIG. 1. The output power is delivered during the first half of the forward stroke with this type of engine. Therefore, the engine could preferably be equipped with a flywheel arrangement for smooth constant angular speed of shaft 22. The flywheel could preferably be located on one of the gear shafts, or to another highspeed shaft, geared to any of the original shafts. The plenum-cooler 46 is shown schematically and not in proper size relationship to the primary and secondary chambers. The plenum-cooler is preferably large with respect to the chambers as for prior embodiments; the principles discussed for FIG. 1 as to the size of the cooler also holding here.

FIG. 5 and FIG. 5a show a hot gas engine having an oscillating main shaft in a double-acting mode. The rotation of shaft 22 is developed from the oscillating motion of shaft 106 by means of crank 26 and rod 24. Several other schemes of generating the oscillating motion and limitation of the rotation angle are possible, but they shall not be further discussed here since such configurations are widely known. Two vanes 108 and 110 are fastened to the oscillating shaft 106 and constitute moving walls in the chambers 112, 114, 116 and 118. Chambers 112 and 114 constitute one basic hot gas engine, and chambers 116 and 118 another, with a common plenum-cooler 46. The A-engine (112 and 114 chambers) and the B-engine (116 and 118 chambers) have individual heaters 30a and 30b and regenerators 28a and 28b. If desired, the heaters could be partially integrated. The vane 108, oscillating in the primary chambers 112 and 116, is shorter than the vane 110. Therefore, the maximum volumes of the primary chambers are smaller than those for the secondary chambers, as for the engine of the embodiment of FIG. 1. The total angular motion of the oscillating shaft 106 is limited to less than The four valves 56a, 56b, 58a and 58b could be of any suitable type and, for example, integrated and located on a rotating shaft in phase with shaft 22, and operated in the same basic manner as described for the embodiment of FIG. 1. This double-acting hot gas engine could be designed to deliver a positive amount of mechanical power to the rotating shaft 22 for typically half the period of a 360 rotation of the shaft. A flywheel could preferably be connected to the output shaft, or to a shaft geared to shaft 22.

FIG. 6 shows a hot gas engine of the invention with a multiple cylinder arrangement. Four separate work cylinders 20a-20d are shown, each of constant crosssectional area, similar to the engine of FIG. 4. Each cylinder has an individual regenerator 28a-28d, but the heater 30 is partially common for the work cylinders, and the plenum-cooler is completely common for all cylinders. As indicated in the figure, the plenum-cooler could be remotely located without dead-space penalty, an advantage, e.g., for automobile-engine applications. The valves 56 and 58 (of the embodiment of FIG. 1) are in the embodiment of FIG. 6 integrated into valves l24a-124d of the 4-port 3 way type. The valve units 124a-l24d could again be e.g., cam-guided disc valves or linear valves, or valve of rotating types mounted on a shaft rotating in phase with the main shaft 22. Bevelled gears 96 and pinions 98 are utilized as in the embodiment of FIG. 4, in order to achieve good balancing and small side forces on the pistons. Naturally, other known crank mechanisms can be used as well. A single flywheel is attached to the main shaft 22 (two counterrotating flywheels could preferably be used on two counter-rotating gear wheels in order to eliminate possible strong gyroscopic effects). The work pistons in FIG. 6 are arranged with 90 phase difference, in an individual order that could be changed.

FIGS. 7a-7d show schematically three ways of regulating the power output from a hot gas engine by changing the valve opening and closing characteristics. FIG. 7a shows a basic hot gas engine of the invention with a schematic 4-port 3-way valve 124, replacing the two valves 56 and 58 in FIG. 1. The possible paths of gas through this valve are denoted I for injection into the primary chamber, II for the transfer flow to the secondary chamber, and III for the exit flow from the secondary chamber 14 into the plenum-cooler. The crank angle of the main shaft is denoted 41, with d for the piston at the left turning point. FIG. 7b shows schematically a preferred way of regulating the power output from the engine, namely by regulating the amount of injection of gas into the primary chamber through control of path 1. Maximum power is here obtained with curve 128, for shutoff of path I for a certain crank angle less than 180. Delaying shutoff as indicated by curve 130, the overall pressure ratio of the cycle is decreased and the mechanical power output per cycle is decreased. The delay may continue to crank angles larger than 180 if desired.

This regulating technique can be said to introduce the plenum-cooler as dead-space when the injection is not terminated according to the ideal curve 128 in FIG. 7b. The dead-space is only coupled to the work chamber 12 for a fraction of the cycle. The control function can be controlled mechanically, e.g. with an axially moveable three-dimensional cam for the case when cam, cam followers and disc valves are used, or by rotating the outer valve housing for a rotating valve,

or in any other known way generate the desired change in phase (with respect to the crank shaft) of the closing of path I.

FIG. 7b shows the alternative regulating procedure, whereby path I is closed earlier than normal, represented by line 132. This method decreases the mass flow through the engine compared to the first scheme, and also increases the overall pressure ratio of the work cycle (and therefore slightly decreases the efficiency compared to the optimum value).

FIG. shows a second basic way of regulating the power output, which will generally be less preferred. The normal opening of path III (indicated by curve 134) takes place at a crank angle for which the recompression in the secondary chamber has proceeded to a stage at which the gas pressure equals the plenum-pressure. By delaying the opening of path III, e.g., as indicated by curve 136, additional pressure build-up in the secondary chamber will take place, before the pressurized gas will flow irreversibly into the plenum-cooler 46. The result of such a delay is a lowering in the net mechanical power output per work cycle, but also a lowering in the thermodynamic efficiency due to the flow irreversibility.

The technique of delaying opening of path Ill can be used even when the path is regulated with help of a check valve. A special type of check valve could be employed using a variable counterpressure to a hydraulic or pneumatic circuit for the control. Such a regulating technique could be made relatively simple, and is to be preferred before a direct mechanical control technique.

The normal closing of path III] takes place for as shown by curve 138 in FIG. 70. If this closing instead is delayed, as represented by curve 140, and path II is opened in a normal fashion, the lowering of the gas pressure in the work chambers is prevented during the reverse stroke, and gas from the plenum-cooler redrawn into the work chambers through line 52. This regulating scheme is essentially identical with that described in FIG. 7b by curve 1130, and could be a favored scheme for certain applications.

Finally, FIG. 7d shows power regulation by changing the opening of path II from the ideal, as represented by the curve 142, to a delayed opening, as represented by curve 144. This scheme gives entropy production from irreversible flow into the secondary chamber, but could be effective in rapidly stopping or slowing the engine.

Other power regulating techniques than the mentioned ones could be used. A simple method could make use of a separate dead-space chamber of variable volume, which could be connected to any of the work chambers continuously during the work cycle (previous methods used interrupted connection with the plenum-chamber dead-space). With the dead-space being minimum the power output would be maximum. The dead-space could be continuously variable, e.g. with help of a hydraulically actuated piston in a cylinder deadspace. Alternatively, step-wise changing of the dead-space volume could be used with help of a series of smaller dead-space chambers that could be connected one by one with the work chambers, by help of a valve arrangement (e.g., sliding valve). Still other power control schemes are possible whereby the pressure level of the plenum-cooler (and possible the introducing plenum-cooler volume) are directly changed. Many variations are possible here, e.g. using separate gasstorage chambers of different pressure than the main plenum chamber, using hydraulically actuated pistons to compress gas in the plenum-cooler, and special devices for limiting the amount of gas in the plenumcooler that can flow into the primary chamber (without unnecessary irreversibilities). These schemes shall not be further discussed here, since they are known within the art.

While the engine of the invention has been described by reference to preferred embodiments, it should be apparent that numerous changes could be made within the spirit and scope of the invention disclosed herein. Some of the possible features are discussed in the following sections of preferred embodiments for the refrigerator, and many of the discussed engine features can be used for the refrigerator of the invention.

DESCRIPTION OF THE PREFERRED EMBODIMENTS (REFRIGERATOR) For the refrigerator we shall discuss both closedcycle and open-cycle devices. The closed-cycle refrigerators have much in common with the previously described closed-cycle hot gas engines. Open-cycle engines have not been discussed in detail or in the drawings of preferred embodiments, mainly due to the relatively low power output per unit swept volume, when operated with ambient air as the working medium.

We first discuss the closed-cycle refrigerator in thermodynamic detail using the average gas model, as for the engine. FIG. 8 shows the closed thermodynamic path with the average gas model (strongly idealized). Basically, the refrigerator is operated similarly to the engine, but with the difference that the heat exchanger 30 is now a simple refrigeration-load heat exchanger 30, with a temperature lower than the gas in the plenum-cooler 46, and that mechanical energy has to be delivered to the shaft 22 in FIG. 1, instead of being extracted therefrom. FIG. 8 relates to a refrigerator with decreasing pressure during the transfer flow process, i.e. to a case when the effective cross section area of the secondary chamber is at least as large as that of the primary chamber. As mentioned earlier, the refrigerator of the invention could be operated in a mode with increasing (or constant) pressure during the transfer flow, provided that an initial expansion in the primary chamber takes place during the later portion of the forward stroke. Such a refrigerator could utilize a secondary chamber with even equal of smaller crosssection area than that of the primary chamber, but will not considered in FIG. 8.

Referring to FIG. 8, the average gas element leaves the plenum-cooler 46 (in a configuration as FIG. 1) with the thermodynamic condition indicated by point 146. Thereafter, the gas flows through the thermal regenerator (ordinary thermal regenerator or counterflow heat-exchanger with two one-way legs) along path 148 of constant pressure to point 150, which ideally has the temperature of the refrigeration-load heat exchanger 30. In the primary chamber 12 the gas element is expanded adiabatically in the reverse stroke, as indicated by line 152, to point 154. The average gas element then returns through the refrigeration-load heat exchanger 30, receiving the refrigeration load as shown by path 156, to point 158 (having the same temperature as point The refrigeration load accepted by the gas element per unit mass is represented by the area under line 156 in a temperature-entropy diagram with linear temperature scale (FIG. 8 employs a logarithmic scale). The gas element returns further through the regenerator 28 and is then reheated by the regenerator matrix to point 160, which has the temperature of gas leaving the plenum-cooler, in this idealized treatment.

Before completion of the reverse stroke, the average gas element undergoes additional expansion, along an isentropic line 162 to point 164, which has the lowest pressure of this ideal cycle. The recompression takes place ideally along path 166 to point 168 (with pressure identical to point 146). Finally, the gas element is cooled in the plenum-cooler along the isobaric line to point 146, and the thermodynamic path is closed.

For an open-cycle, the gas element under study is simply exhausted at point 168, provided that employed inlet and exhaust pressures are equal, and the refrigerator not driven fully or partially with pressurized gas. Such a device will be discussed later.

With the idealized refrigerator model of FIG. 8, the refrigeration-load was applied in a separate heat exchanger 30. Naturally, in a practical situation no such separate exchanger may be required, and the load applied e.g. directly to the walls of the primary chamber, or to other heat-exchanging surfaces that the working gas can be in contact with. The same argument could hold for the heat rejection in the plenum-cooler. Part or all of the heat rejection could take place through a heat exchanger (not shown in any of the figures) located in the secondary chamber or in the gas line to the right of the thermal regenerator, or simply by cooling the walls of the secondary chamber. However, for already discussed reasons, the major portion of the cooler could preferably be located in the plenum chamber for closed-cycle operation.

FIG. 9 shows an open-cycle refrigerator with cold gas bleedoff, without a refrigeration-load heat exchanger. This refrigerator is specially adapted for cooling an enclosed volume of gas (air), and transfers the absorbed heat from this volume to gas located outside the enclosed volume (the ambient air). The enclosed volume is assumed to contain gas which is substantially cooler than the ambient gas. For optimum efficiency, the refrigerator is equipped with two thermal regenerators 28 and 170, one cold gas inlet 172 from the enclosed volume, one warm gas inlet 174 from the ambient atmosphere, a still colder gas outlet 176 (into the enclosed volume), and a still warmer gas outlet 178 (to the ambient atmosphere). The necessary valves are here shown as five 2-port l-way valves 180, 182, 184, 196 and 190. These valves could be combined into fewer units, as discussed previously for the engines.

The open-cycle refrigerator in FIG. 9 can be operated in several ways, e.g. as follows: At the first portion of the forward stroke, valve 182 is kept open to permit cold gas to flow through the regenerator 28 into the primary chamber 12. The right side of the regenerator 28 has a temperature which ideally is close to the inlet temperature in channel 172. Thereafter, in a later portion of the forward stroke warm gas from line 174 is fed through valve 184 and through both re-generators into the primary chamber. The right side of the regenerator 170 has a temperature at steady-state condition which ideally is close to the temperature of gas in the warm inlet line 174. The injection of gas through line 174 could be terminated at an arbitrary point of the forward stroke, e.g. at the end of the stroke, and for which case the effective area of the secondary chamber must be larger than that of the primary chamber (as shown in FIG. 9). With the primary and secondary chambers made up from a single constant-diameter tube, the injection through line 174 must be terminated before the end piston has completed the forward stroke. In the reverse stroke still colder gas is forced through valve 180 to the exit line 176, followed by the transfer process of the remainder of the gas in chamber 12 to chamber 14 through valve 190. In an alternative scheme the transfer process could take place before ejection of gas through line 176, in the reverse stroke, whereby entropy increase would be less and the overall thermodynamic efficiency ideally better.

The cold gas that is bled through valve 180 has attained the lower temperature from the falling temperature profile in regenerator 28 (going from to left). The amount of bleed is controlled by valve 180 and should be identical with the amount of gas taken in through line 172.

If the cold gas bleedoff takes place before the transfer flow in the reverse stroke, the valve 190 should not be opened until valve 180 is closed. If no backfiow has been permitted through valve 186 into the secondary chamber during the cold-gas bleedoff, a free expansion will take place when valve 190 is opened that have an adverse effect on the overall efficiency of this refrigerator. For this reason such backflow could preferably be used.

In the forward stroke, valve 186 is opened when the gas pressure in the secondary chamber has been raised to the ambient level, and the still warmer gas exhausted through line 178. Valve 186 could be a simple check valve if no backflow of the mentioned kind is desired.

With this type of refrigerator, bleeding an intermittent flow of cold gas without the use of refrigerationload heat exchangers, the load is obviously placed upon the regenerators. With the refrigerator of FIG. 9 the load is on regenerator 28, and gas from the ambient atmosphere used in the work cycle to maintain the desired falling temperature profile in regenerator 28 for cooling of the cold gas, and for transport of the heat (and mechanical power from the shaft converted to heat) to the ambient atmosphere. The thermodynamic characteristics of such systems have been computed using gas models as shown in FIG. 8, as well as with correct infinite element models. With cold-gas bleedoff a fundamental difficulty exists in the treatment of the energy balance of the regenerators, due to the fact that the regenerator experiences a net mass flow and a small amount of entropy has to be generated from heat transfer in the matrix.

In the device of FIG. 9, the cold-gas bleed could be replaced with a refrigeration-load heat exchanger, with elimination of valves 180 and 182. Naturally, the same open-cycle refrigerator could be modified to employ closed-cycle operation for gas that flows through the secondary chamber and establishes the desired temperature profiles in the regenerator matrixes. In this case a cooler and a plenum chamber have to be added in the normal fashion. Cold-gas bleedoff is maintained as previously described, with the cold-gas bleedoff rate equal to the rate of inlet flow for steady-state operation.

FIG. 10 shows a simpler open-cycle refrigerator than FIG. 9, with only one regenerator 28, one inlet 174, and a cold gas outlet 176, and a hot gas outlet 178. The valves 184 and 190 of FIG. 9 have been replaced in FIG. 10 with a single 3-port 2-way valve 192, with an operating function similar to that of valve 56 in FIG. 1.

According to one operating scheme, permitting an arbitraty area ratio for the work chambers, injection through line 174 is terminated before completion of the forward stroke. In the reverse stroke the transfer process is accomplished through valve 192, after which recompression of gas in chamber 12 follows, and subsequently cold-gas bleedoff occurs through line 176 when the pressure level in the primary chamber has been raised to the desired level. A device of this kind could be built with the primary and secondary chambers made from a single cylinder and a simple piston, and could be used for simple heat-pump or refrigeration applications.

With the device shown in FIG. 10 recompression takes place in the secondary chamber, with subsequent exhaust through valve 186 (check valve or mechanically controlled valve). The refrigeration load is applied on the thermal regenerator as for the previously discussed open-cycle device.

An alternative operating scheme with a device as shown in FIG. 10 could employ cold-gas bleed before the transfer flow process. Such a device should not use an initial expansion following injection during the forward stroke, and must therefore necessarily have a larger cross section for the secondary chamber. to accomplish a decreasing pressure during the transfer process.

In order to avoid splitup in transfer flow and coldgas bleedoff during the reverse stroke, e.g. with the configuration of FIG. 10, a separate cold-gas storage chamber can be employed. Such a chamber would be filled during the forward stroke, and emptied during the full reverse stroke. Overflow from this additional chamber to the primary and secondary chambers during the reverse stroke (with decreasing pressure in the last chambers) can be prevented with help of a checkvalve arrangement. The additional chamber could be formed by help of an extension of the piston and an additional cylinder. Such devices will be briefly discussed with regard to FIG. 13, and are within the spirit and scope of the invention.

FIG. 11 shows an open-cycle refrigerator similar to that of FIG. 10, but with the basic difference that the piston assembly is divided in two parts, with a phase difference in motion. This is accomplished with the two cranks 26 that have individual connecting rods 24. The part of the piston assembly that reciprocates in the primary chamber 12 is in an earlier phase than the piston in the secondary chamber 14. With this device, coldgas bleedoff through valve 180 can take place during the first portion of the reverse stroke of piston 16a (in the primary chamber), and the transfer flow can be initiated when the volume of chamber 14 is minimum, i.e.

the flow into chamber 14 can be made in a reversible fashion. Proper initiation of the transfer flow is assured by valve 192, which in a conventional way could be guided by the angular position of the rotating shaft 22.

The types of refrigerators shown in FIG. and FIG. 11 could be driven by pressurized working gas, or by a combination of pressurized gas and mechanical power into the rotatable shaft 22. With no mechanical energy supplied to the rotatable shaft 22, the working gas is then expanded when flowing through the refrigerator (the recompression could be eliminated), and the exit pressure is smaller than the inlet pressure. The crank mechanism of the rotating shaft is then used to control the length of the piston stroke, to guide the valves (proper opening and closing characteristics), and possibly also to provide a connection with a flywheel device for smooth reciprocating motion of the piston assembly.

FIG. 12 is an example of an open-cycle refrigerator, driven by pressurized working gas, with the refrigeration load applied to a heat exchanger, and employing a two-step expansion process. The inlet pressure in line l94 is higher than the exit pressure inline 196, here fed to the ambient atmosphere. No mechanical energy is supplied to shaft 22, and mechanical energy may instead even be extracted through the shaft, or dissipated in an arbitrary fashion. An optional counterflow heatexchanger 198 is used for an initial cooling process, in which the incoming gas (line 194) transfers heat to the outgoing gas (line 196). The refrigeration load is applied to exchanger 30, thus implementing the refrigera tion. A third working chamber 200 is employed for the second expansion step in the transfer from the secondary chamber 14 to this chamber. Therefore, the effective cross-sectional area of the face of the piston 16b in chamber 200 is larger than in the secondary chamber 14. An additional valve 202 (e.g. controlled by the angular position of shaft 22) is used for the control of the second transfer flow into chamber 200. The geometry of the refrigerator is preferably adjusted such that the gas pressure in chamber 200 is identical with the ambient pressure (the pressure in the exhaust line 196) after the final transfer flow to chamber 200.

Similar to the previous embodiments, the matrix of the regenerator has a decreasing temperature profile, going from right to left in FIG. 12. This temperature profile is established and maintained so that the gas pressure in the transfer flow process is lower than during gas injection into the primary chamber. Basically, the cooling in this refrigerator takes place in two steps, first in the counterflow heat-exchanger 198, and thereafter in the regenerator 28. The decreasing temperature profile in the heat exchanger 198 (following the inlet gas) is maintained with help of the gas cooling from adiabatic expansion in the second transfer flow from the secondary chamber 14 to chamber 200. The function of valve 202 is to provide an open path between chambers 14 and 200 during the forward stroke of the piston. The function of valve 58 is to provide an open path between chamber 200 and the exit channel 196 during the reverse stroke.

The average residence time of gas in the refrigerator of FIG. 12 corresponds to nearly two complete piston cycles: injection into the primary chamber 12 in the first forward stroke, transfer to the secondary chamber 14 in the first reverse stroke, transfer to the additional expansion chamber 200 in the second forward stroke, and finally exhaust from chamber 200 to line 96 in the second reverse stroke.

The refrigerator of FIG. 12 could be modified in a large number of ways. One modification could be elimination of both the heat exchanger 30 and the regenerator 28, but with addition of one valve. This valve would provide cold-gas bleedoff from the primary chamber. The cooling in such a refrigerator would only be in one step, due to the transfer flow from the secondary chamber 14 to the additional chamber 200, and the primary chamber 12 used for pumping cold gas. Briefly, the operation of the refrigerator would involve the following steps: filling of the primary chamber 12 during the first forward stroke, filling of the secondary chamber from line 194 with help of valve 56 (this valve now has a different function) during the first reverse stroke, and cold-gas bleedoff from chamber 12 through the mentioned valve addition during the same stroke, transfer flow under decreasing pressure from the secondary chamber to chamber 200 during the secondary forward stroke (by help of valve 202), new filling of the primary chamber 12 during the same stroke, and finally exhaust from chamber 200 to line 196 and coldhas bleedoff during the second reverse stroke. Since this refrigerator would operate with full gas pressure of the source in chamber 14 after completed reverse stroke, the pressure ratio for the device should be smaller than for the first described device, or alternatively the injection from the source (line 194) into chamber 14 terminated before completion of the reverse stroke, in analogy with schemes discussed for the engine of the invention.

FIG. 13 shows a closed-cycle refrigerator with twostep cooling and three work chambers. The plenumcooler is not shown, nor are the rotatable shaft and crank mechanisms, that could be used to provide the proper reciprocating piston motion. Two regenerators 28 and 204 are used, and an additional chamber 206 formed by the reciprocating piston and a cylinder 207 with a larger effective cross-sectional area than cylinder 209 (forming the cold chamber 12).

Chambers 14 and 206 could be made with the same constant-diameter cylinder and hence with the corresponding piston portions of the same diameter, making the mechanical construction considerably simpler. The operation of such a device necessitates that the injection into chamber 12 be terminated before the completion of the forward stroke, and is also characterized by an increase in pressure during the transfer flow to chamber 14.

In understanding the operation of this device, the regenerator could be considered to operate in conjunction with the larger primary chamber 206 and provide the first-step cooling, with the refrigeration load being second step (heat exchanger 30, etc.). Hence, the colder primary chamber 12 could be considered to be the primary chamber in a refrigerator with chamber 206 as the secondary chamber.

To adapt this refrigerator for use as an open system with cold-gas bleedoff with expenditure of gas from the gas source in line 54, the heat exchanger 30 would be eliminated and a bleed valve introduced near the chamber 12. If the bleed from chamber 12 takes place

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Classifications
U.S. Classification60/522, 60/526, 62/6, 60/525
International ClassificationF02G1/00, F02G1/044, F02G1/05
Cooperative ClassificationF02G1/044, F02G2244/50, F02G1/05, F02G2258/10
European ClassificationF02G1/044, F02G1/05