Bland/Ewing Cycle Improvements

This application is a continuation-in-part application of U.S. patent application Ser. No. 17/746,848, filed May 17, 2022, which claims priority to, and the benefit of, U.S. Provisional Application No. 63/189,634, filed May 17, 2021, the entire content of each of which is hereby incorporated by reference. This application is also a continuation-in-part application of U.S. patent application Ser. No. 18/197,092, filed on May 14, 2023, which claims priority to, and the benefit of, U.S. Provisional Application No. 63/342,093, filed on May 14, 2022, the entire content of each of which is hereby incorporated by reference. This application is also a continuation-in-part application of U.S. patent application Ser. No. 18/362,951, filed on Jul. 31, 2023, which claims priority to, and the benefit of, U.S. Provisional Application No. 63/393,960, filed Jul. 31, 2022, and U.S. Provisional Application No. 63/439,781, filed Jan. 18, 2023, the entire content of each of which is hereby incorporated by reference. This application is also a continuation-in-part application of U.S. patent application Ser. No. 19/275,589, filed on Jul. 21, 2025, which claims priority to, and the benefit of, U.S. Provisional Application No. 63/674,261, filed on Jul. 22, 2024, the entire content of each of which is hereby incorporated by reference.

This application also incorporates by reference the entire content of U.S. patent application Ser. No. 18/095,463, filed on Jan. 10, 2023, and issued as U.S. Pat. No. 12,352,250, on Jul. 8, 2025.

This application is in part related to U.S. Pat. Nos. 4,817,388, 5,179,839, 3,067,594, 3,225,538, and 3,871,179.

U.S. Pat. No. 3,067,594 proposed an open-cycle Bland/Ewing chemo-thermodynamic process. U.S. Pat. No. 3,067,594 proposed a closed-cycle Bland/Ewing chemo-thermodynamic process. U.S. Pat. No. 3,871,179 proposed the application of the B/E cycle to the classic Stirling cycle.

A Stirling cycle involves a strictly ideal series of two isothermal processes separated by two isochoric (i.e., constant volume) processes. That is, for a vapor or gaseous working fluid, from an initial point of lowest pressure and temperature, a Stirling cycle engine posits the use of:

A classic Stirling cycle is clearly described in the specification and drawings of U.S. Pat. No. 3,871,179, FIGS. 1A through 1D. Per those figures, beginning at the position of the upper and lower pistons as shown in FIG. 1A of U.S. Pat. No. 3,871,17, an ideal Stirling cycle engine would seem to require:

Thus, a classic Stirling cycle would contain sharp “points” where each constant volume process ends and each constant temperature process begins. That is, the cycle would require an essentially impossible scenario to be enacted, where the various pistons are required to instantaneously stop, hold position, and then instantaneously restart in the opposite direction. It would also require both isothermal heating and isothermal cooling.

So-called “Stirling engines” do exist and do produce useful work relatively efficiently. Those practiced in the art of Stirling engine design are highly cognizant of the difference between an ideal “Stirling cycle” and a practical “Stirling engine”. In a Stirling engine, the Stirling cycle's “pointed” processes are essentially “blended” within a single contiguous volume that is essentially shuttled between two or more positive displacement, contiguous but non-synchronous cylinders with pistons or lobes acting upon the contiguous volume. As a result, in a practical Stirling engine, the sharp “points” of the classic Stirling cycle are rounded into curves that are blended into one another, containing no single segment with a true constant volume or constant temperature process. In other words, since halting a piston in mid-movement is extremely difficult if not impossible to accomplish efficiently, constantly reciprocating pistons are used to create an approximation of a Stirling cycle.

In essence, a Stirling engine can be considered a kind of blend of a Stirling cycle and an Ericsson cycle.

On Sep. 20, 2000 Joseph B. Bland was awarded a grant from the California Energy Innovation Small Grant (EISG) program for construction of a Closed Cycle Valved Cell (CCVC) testbed. See U.S. patent application Ser. No. 18/362,951,andfor a photograph andfor a cutaway solid model view of the EISG CCVC testbed. The “EISG CCVC testbed” was constructed and tested under the terms of that program and the results were submitted to the program. The data received indicated that the processes for which it was designed functioned largely as predicted.

The EISG grant was based on the concepts patented in U.S. Pat. No. 4,817,388, “Engine with Pressurized Valved Cel”1, and U.S. Pat. No. 5,179,839, “Alternative Charging Method for Engine with Pressurized Valved Cell”, both granted to Joseph B. Bland, which proposed the concept of an auxiliary “valved cell” for connecting a compressor to an expander via a “transfer valve”. Over time, it has come to be appreciated that valved cells can be seen as means to connect various kinds of heat engine processes to one another, including, for example, a constant volume displacement regenerator.

The valved cell process described in U.S. Pat. No. 4,817,388 is as follows: “There is, therefore, provided in practice of this invention according to a presently preferred embodiment a method of operating an engine comprising the steps of compressing a gas to a pressure approximately the same as a pressure in the engine, temporarily isolating a mass of the compressed gas, and opening communication between the isolated gas and the engine while the isolated gas is at approximately the same pressure as in the engine for intermittently releasing substantially all of the temporarily isolated mass of gas into the engine for expansion.” U.S. Pat. No. 5,179,839 discloses an alternate means of accomplishing the same result.

As stated in U.S. Pat. Nos. 4,817,388 and 5,179,839, a valved cell is used to present a compressed gas and/or vapor to an expander, and a special kind of intake valve or “transfer valve” is used to connect the valved cell to the expander. The transfer valve is designed to almost instantly connect the valved cell to the expander just following early closure of the expander exhaust valve and the naturally-resulting recompression of remnant gases thus trapped in the cylinder head, such as to at least match the pressure within the recompressed gas with the pressure of the gas within the valved cell. Finally, constant pressure recharging of the valved cell is made to occur just following or just prior to instant closure of the transfer valve following displacement of the contents within the valved cell into the expander, depending on the use of either the charging system of U.S. Pat. Nos. 4,817,388 or 5,179, 839.

In either a Stirling cycle or a Stirling engine, a cyclical positive displacement system with a single contiguous volume is assumed. However, as has been demonstrated in the EISG CCVC testbed, true constant volume (isochoric) displacement heat transfer processes are in fact possible using non-contiguous but synchronized volumes cyclically connected by valves. In the EISG CCVC testbed, the valved cell concept in those patents was used to cyclically connect and disconnect two separate constant volume or isochoric working fluid displacement processes.

Starting at BDC, in the first EISG CCVC testbed isochoric working fluid displacement process, ideally all the gases within the engine are at the same pressure, as will be shown, which is the pressure of the feed gas just taken into the small first displacer cylinder. At BDC, (a) the expander exhaust valve has just opened while simultaneously (b) closing the transfer valve. (Note that, in closing the transfer valve, the transfer valve spring bias, which is towards open, has been compressed. However, the transfer is held physically closed by the opened exhaust valve. Eventually, when the expander exhaust valve closes, the pressure holding the transfer valve will be much higher, keeping it closed until it is eventually triggered after pressure across the transfer valve equalizes, as will be shown.) Hot, low pressure expanded working fluid is about to be exhausted from the expander by the expander piston. If there is less pressure on the exhaust side of the expander cylinder than there is in the expander cylinder, pressure difference will (c) force open a spring-biased expander exhaust check valve. If there is more pressure on the exhaust side of the expander cylinder than there is in the expander cylinder, the exhaust check valve will not open. Instead, compression on the expander side of the expander exhaust check valve coupled with simultaneous expansion on the displacer/compressor side of the expander exhaust check valve will eventually force open the expander exhaust check valve. Once the expander exhaust check valve is open, exhausting working fluid from the expander (d) passes through one side (the expanded expansion gas side) of a recuperating heat exchanger or recuperator. (A heat exchanger may be used to capture and thus exchange heat, and typically has separate counter-flowing heating and cooling fluids, separated from one another by a common physical wall, much as how a radiator on a car separates a flow of cooling air from the heated water that was used to cool the engine. A recuperator is essentially a heat exchanger with the specific function of reusing otherwise-waste heat to reduce the overall thermal input requirement for a heat engine or similar heat consuming process. In short, a recuperator is not a heater or a cooler but a “re-user” of otherwise-waste heat.) The exhaust from the expander then (e) isochorically passes the recuperator, through a cooler, and into the combination displacer/compressor, which approximates the same volume as the expander (in the EISG CCVC testbed, the expander volume was slightly larger than the displacer/compressor volume due to the existence of a small diameter drive shaft connecting the compressor/displacer piston to the engine crankshaft). As a result, the “exhaust displacement” process gave up heat virtually isochorically to the recuperator. At the end of the expansion exhaust process, (f) the expander exhaust valve closes. Note that the expander exhaust valve is designed to close before it reaches TDC such that it can (g) recompress any remnant working fluid to at least the pressure of working fluid held on the other side of the transfer valve. Note that this remnant fluid recompression process is exactly the same is disclosed in U.S. Pat. Nos. 4,817,388 and 5,179,839.

In the second isochoric EISG CCVC testbed working fluid displacement process, exhausted heat was isochorically absorbed from the recuperator when a small first displacer exhausted a quantity of pressurized working fluid, (h) through the other side (the recompressed and cooled gas side) of the recuperator, (i) through a heater, and (j) into an equal-sized small second displacer, isochorically raising both the “captured” working fluid's temperature and pressure with very little work put in. When the stroke reversed, the second displacer then (k) passes the isochorically heated and thus higher pressure working fluid back through the heater to the expander's transfer valve, and, via the transfer valve, (l) into the expander which, at that point, is at TDC. As the expander piston begins moving away from TDC, several things happen; (m) the small first displacer exhaust check valve closes due to spring bias, which (n) allows pressure to begin dropping in the small first displacer. Simultaneously, (o) the small second displacement piston begins to expel working fluid through the heater, past the transfer valve, and into the much larger expander in what can be termed a displacement/expansion process. Simultaneously, (p) the exhaust displacer/compressor piston reverses direction and begins compressing the cooled working fluid captured between it and the expander exhaust check valve towards the input pressure of the small first displacer. That increase in pressure eventually (q) opens a check valve connected to the displacer/compressor, (r) exhausting the compressed and heated working fluid past the open check valve and(s) through an external constant pressure cooler/heat exchanger. The cooled and densified working fluid exiting the cooler/heat exchanger is directed to (t) the small first displacer's intake check valve, completing the working fluid circuit. (One possible improvement over the EISG CCVC testbed would be to exhaust into a pressurized working fluid storage tank, thus ensuring that the pressure exhausting from the displacer/expander is the same as the pressure entering the small first displacer but at a much lower temperature and therefore density.) Meanwhile, the remnant higher pressure working fluid captured within the small first displacer (u) drops in pressure as the small first displacer piston expands it until (v) the pressure on the outer side of the small first displacer intake check valve becomes higher than the pressure within the small first displacer, at which point (w) the spring bias of the small first displacer intake check valve is overcome and (x) the cooled working fluid is taken at constant pressure into the small first displacer. Ideally, as BDC is reached, the pressure in the small first displacer would be equal to the pressure in the expander, the small second displacer, and the connecting plenums (including the recuperator) between the small first displacer's exhaust check valve and the expander, which is also equal to the pressure in the displacer/compressor, the recuperators “other side”, and the various connecting plenums. That is, all volumes within the engine are momentarily at the same pressure. (Note that this moment of pressure equalization will ultimately depend on how much thermal energy is added to the working fluid during the two heat input processes described above.) At that moment, (y) the expander exhaust valve is opened, which is made to (z) simultaneously close the transfer valve. The process then begins again at (a) above.

The result of this process is a heat engine with true isochoric displacement heat transfer capability as compared to the non-isochoric displacement heat transfer capability found in typical Stirling engines. Careful measurements were made of the prototype engine, verifying that thermal energy was successfully passed isochorically from the exhausting gas to the inflowing gas during the two isochoric processes. Note that a small amount of recompression was used within the expander to “open” the transfer valve at Top Dead Center (TDC), in the manner described within U.S. Pat. Nos. 4,817,388 and 5,179,839. Since the same working fluid was constantly recirculated through the engine, the engine is thus classifiable as a “closed cycle” (CC) heat engine. And since it used a recompression process to aid in opening the transfer valve, the engine is thus classifiable as a “valved cell” (VC) engine. Hence the term CCVC engine.

One less obvious advantage of the EISG CCVC testbed over Stirling engines concerns the expansion and compression processes. In a Stirling engine, expansion takes place generally in the hotter expander and compression takes place generally in the cooler compressor, which is thermodynamically desirable. However, some expansion also occurs in the cooler compressor space, and some compression occurs in the hotter expansion space. Note that, in the EISG CCVC testbed, all expansion takes place in the hotter expander space, displacer space, the exhaust side of the recuperator space, and connecting manifolds. Similarly, all compression takes place in the cooler displacer/compressor space, the cooler space, the exhaust side of the recuperator space, and the connecting manifolds.

Note that, to the degree that “dead space”, which is the excess volume above and beyond that (constant) volume within the displacers, that is, the volume within the heater, the cooler, the recuperator, and the various connecting manifolds, can be reduced, the overall thermodynamic efficiency of the EISG CCVC testbed can be increased, as will be shown.

An Ericsson cycle, like a Stirling cycle, reuses gases following the expansion process to “preheat” gases prior to the addition of source heat, this theoretically greatly increasing thermal efficiency. In a Stirling cycle, that preheating and source heat addition occurs at constant volume or isochorically, while an Ericsson cycle, that preheating and source heat addition occurs at constant pressure or isobarically. In addition, both cycles propose constant-temperature or isothermal expansion and compression, although neither perfectly accomplishes that. However, because Ericsson cycles occur at constant pressure, it is much easier for them to utilize counter-flow heat exchangers such as those used in the EISG CCVC testbed, since the length of the heat exchangers and the time required for exchanging heat are not as important in constant pressure processes as they are in constant volume processes.

Thus, in passing through large-volume heat exchangers for heating, recuperating, and cooling, the EISG CCVC testbed can be said to resemble an Ericsson cycle. In fact, the main difference is that the heat exchange in the EISG CCVC prototype is done in a “pulsed” pressure manner, not at constant pressure. The EISG CCVC prototype also used a gas compression phase, just as does an Ericsson cycle.

In other words, since the EISG CCVC engine uses a gas compression phase and heat exchangers but also an isochoric heat input phase, it is classifiable as a hybrid Ericsson/Stirling closed cycle heat engine.

There is an alternative to recuperation called regeneration. In what can be called a “single-stream” regenerator, heat is exchanged to and from a working fluid into and out of what might be envisioned as a “thermal sponge”, which is often a mass of very thin wires, sometimes separated into layers by very thin sheets of perforated thermal insulation, that can very quickly absorb or release thermal energy at varying temperatures from one end to the other of the “sponge”.

Presently, thermal regeneration within Stirling engines uses a “single-stream” regenerator (SSR). Usually, an external heater is attached to the hotter end of an SSR and an external cooler is attached to the colder end. By (A) alternating the working fluid's flow (1) in a direction towards the “hot” end of the sponge during expansion, and (2) in a direction towards the “cold” end of the sponge during compression and (B) cyclically adjusting the volumes of the hot and cold spaces in conjunction with the requirement for expansion or compression, expansion takes place generally in the hotter expander and compression takes place generally in the cooler compressor. The sponge thus is seen as a “preheater” when the working fluid is moved towards the expander, and as a “pre-cooler” when the working fluid is moved towards the compressor. In a classic Stirling cycle or Stirling engine, a single contiguous quantity of working fluid is flowed through a combination cooler/regenerator/heater, first in one direction, as to be heated, and then in the opposite direction, as to be cooled.

A second kind of regenerator is possible that can be called a “dual-stream” regenerator (DSR). When ducts/ports are used, it may be termed a “ducted DSR” (DDSR) and when valves are used it may be termed a “valved DSR” (VDSR). A DDSR is sometimes composed of a thick rotating disc of regenerator material with an outlet duct/port on one side of the rotating disc and an inlet duct/port on the other side of the rotating disc. A first stream of gas or vapor is passed through one duct/port and a second stream is passed through the other duct/port, usually in opposite directions perpendicular to the plane of rotation. These can be used, for example, by a combustion heater exhausting hot combusted fuel and air through the ducted regenerator or a portion of the ducted regenerator, the thermal energy captured within the rotated or rotating regenerator “thermal sponge” material then being used to preheat non-combusted air (and sometimes fuel) entering the combustion heater.

One embodiment of a valved regenerator can create a kind of “valve-switched DSR” (VSDSR), where two or more regenerators cycle between (a) being charged with thermal energy and (b) giving up that energy. One technique for constructing such a VSDSR is to use regenerators that “switch places” using valves. One type of VSDSR was proposed within U.S. patent application Ser. No. 17/746,848. A more prosaic schematic of a VSDSR is shown in U.S. patent application Ser. No. 18/362,951,and, referred to therein as a Synchronized Thermal Regenerator Exchange Pump (STREP). In essence, two or more stationary regenerators “take turns” via switching valves either being charged with thermal energy or having thermal energy removed. Other techniques, such as a rotation-switched DSR (RSDSR), are possible as well, but may still require valves or ports to shut down the flow while the regenerator cores shift, especially in the case of high pressure differentials between heating and cooling working fluids.

A VSDSR operating in syncopation may be used to achieve a more constant flow than a simple VDSR. In U.S. patent application Ser. No. 17/746,848, it was proposed that, rather than a recuperator, the VSDSR shown in FIG. 6 of U.S. patent application Ser. No. 17/746,848 be used for a simple high temperature heat exchange. In that instance, a reactant mix flowing into an endothermic reactor flowed through one regenerator core while a product mix flowing out of the endothermic reactor flowed through a second regenerator core. At intervals, valves “switched” the regenerator core flowing into the reactor and the regenerator core flowing out, thus “using” for a time the store of heat in the core recently charged with product thermal energy to heat up the reactant while the other core was thermally recharged, and thus allowing an essentially constant pressure flow of reactant/reactant mix into the reactor and product/product mix out of the reactor, while simultaneously achieving a highly efficient heat exchange at the minor cost of a small amount of mixing between the reactant and the products.

Sometimes an isobaric flow may be slow enough and/or the regenerator may be long enough to permit an intermittent pressure change small enough to permit use of a VDSR or DDSR with a single shared reactor core. An example of a VDSR can be seen asthroughin U.S. patent application Ser. No. 18/362,951. A more prosaic schematic of a simple VDSR is shown in U.S. patent application Ser. No. 18/362,951,and

Note that, in the present EISG CCVC testbed design, the expander exhausts at the same time the small first displacer exhausts, apparently making use of a VDSR or DDSR impossible. In one improved EISG CCVC testbed, the existing recuperator would be replaced by a VSDSR or possibly a RSDSR to permit regeneration rather than recuperation. However, various issues will arise due to the requirement for such a VSDSR or RSDSR to “pressure match”, increasing engine complication.

An interesting EISG CCVC testbed embodiment has been discovered that easily converts the existing EISG CCVC testbed recuperator into what can be termed a combination exhaust regenerator/recuperator. Rather than exhausting directly from the displacer/compressor, it is herein proposed that, following the lower pressure working fluid displacement from the expansion chamber and into the displacement/compressor chamber, ending at approximately BDC for the engine, the flow be reversed in direction back through the cooler and then back through the recuperator, where the working fluid is finally exhausted external to the engine via a second exhaust valve which can be termed a main engine exhaust valve. Note that this second exhaust process via the main engine exhaust valve can either (a) follow a drop in exhaust pressure to, for example, ambient pressure, (b) occur at the achieved lower pressure at the end of the expander exhaust stroke, (c) include a recompression of the low pressure working fluid prior to a constant pressure exhaust out of the engine, or (d) include an early closure of the final exhaust valve near TDC to allow “pressure matching”.

In this improved EISG CCVC testbed embodiment, the engine would not exhaust through an exhaust check valve connected directly to the displacer/compressor. Instead, a main engine exhaust valve, which may be an exhaust check valve, an actuated exhaust valve, or some combination of the two, would be connected to the engine where the expander exhaust port connects the expander exhaust to the hot end of the existing recuperator. As a result of this change, at the end of the compression process the working fluid is exhausted at constant pressure out of the displacer/compressor cylinder, back through the cooler, back through what becomes in essence an exhaust regenerator, and out of the displacer/compressor exhaust valve. This accomplishes several things:

Note that the exhaust recuperator/regenerator concept proposed above changes the syncopation between the two displacement processes, such that the final exhaust from the displacement/compressor now occurs out of phase with the displacement process between the small first displacer and the small second displacer.

In the EISG CCVC testbed, the working fluid exhausting from the expander undergoes two distinct cooling processes; the working fluid is first cooled by the recuperator, and is then cooled by a separate cooler. In a modified EISG CCVC testbed regenerator/recuperator, the cooler may be removed, connecting the regenerator/recuperator directly to the displacer/compressor. All cooling may then be undertaken by the working fluid being displaced from the small first displacer, through the recuperator side of the recuperator/regenerator, through the heater, and into the small second displacer. That has the benefit of likely increasing the waste thermal input to the EISG CCVC testbed small displacer displacement process.

Alternatively, the EISG CCVC testbed regenerator/recuperator may be completely converted to a kind of combination regenerator. First, the expander exhaust would pass through what would amount to an SSR with the gas exhausting from the expander, through the SSR, through the existing cooler, and into the displacer/expander. An intake check valve is added either on the hot side of the expansion side cooler or on the intake to the displacer/compressor. The existing displacer/compressor exhaust check valve is attached directly to the exhaust SSR cold side, Because of the two check valves, the reverse flow out of the displacer compressor will bypass the cooler and connect directly to the cold end of the SSR, increasing SSR hot end exhaust temperatures.

Second, to tap into the waste heat exiting the SSR, the recompressed and cooled gas side recuperator may be replaced with a recuperator VDSR between the small first displacer and the source heater. An EISG CCVC testbed SSR exhaust heat-powered recuperator VDSR would be composed of a single regenerator core, housing and manifolding, and four valves; a recuperator VDSR hot side intake check valve connected via a manifold to the SSR hot side, a small first displacer actuated exhaust valve connecting to the recuperator VDSR cold side, a small second displacer intake check valve connecting the recuperator VDSR hot side to the heater and the small second displacer, and a main engine exhaust valve connecting to the recuperator VDSR cold side.

U.S. patent application Ser. No. 17/746,848 states that heat engines are defined as work-generating devices that operate as a result of a temperature difference in their working fluid. The Carnot theorem for maximum theoretical efficiency of a heat engine, mathematically expressed by the equation (Th—Tc)/Th, where Th is the absolute temperature of the hot reservoir or heat source and Tc is the absolute temperature of the cold reservoir or heat sink, specifies limits the thermal efficiency that any heat engine can obtain to the absolute temperature difference between those two thermal reservoirs.

A heat engine cycle does not materially change (Th—Tc)/Th; that is, a heat engine with a given Th and Tc still has the same (Th—Tc)/Th. However, in real world engines, various unavoidable losses, for example friction losses, resistance to flow losses, radiation losses, and so forth, impact and define the percentage of delivered thermal efficiency versus ideal thermal efficiency. Delivered thermal efficiency is defined herein as net work out or Wout (n) divided by total source heat in or Hin (t), or Wout (n)/Hin (t).

In theory, either regeneration or counter-flowing separate stream heat exchange (standard heat exchange) can approach a perfect exchange of thermal energy, where the cold working fluid flowing in from the “cold side” can be made to equal the temperature of the hot working fluid flowing in from the “hot side” and vice versa. In reality, the heat exchange is not perfect for either device. In part, complete heat exchange is a function of the length of the heat exchange device, in part it is a function of the cross section of the heat exchanger's internal passages (smaller passages allow faster heat transfer), in part it is a function of the increase in friction and thus “pumping losses” with any increase in internal volume or decrease in internal passage cross section, in part it is a function, in the case of the recuperator, of the thickness of the internal walls that separate the two flows, thus reducing heat transfer, and in part it is a function of the time the counter-flowing gas streams are given to exchange heat. Given enough volume/length and a slow enough flow rate, and assuming excellent heat insulation, recuperation can come close to regeneration's capability for heat exchange, but for a given degree of heat exchange at a given flow rate for a given internal volume and a given pumping loss over a given length of time, a regenerator is vastly superior to a recuperator.

This difference in flow rate, internal volume, and heat transfer time is particularly meaningful for isochoric processes. Consider the drawings of the ideal Stirling cycle in U.S. Pat. No. 3,871,179,: It is clear that between the two pistons there is a physical space, made up of a heater heat exchanger, a regenerator, and a cooler heat exchanger. If these spaces were not present, and somehow the working gas could be transported directly from one cylinder to the other while still effecting the desired heat transfers, then the volume being transferred from the one piston to the other piston would be exactly equal to the total volume of working fluid. Instead, there is an additional volume in between the two pistons; a kind of “dead space”. The size of that additional volume will determine the amount of thermal energy transferred (into and out of the working fluid) to elicit a given increase or decrease in pressure and temperature.

That is, for a given isochoric thermal input per stroke, replacing the regenerator with a recuperator in a Stirling cycle will drastically reduce a rise in both pressure and temperature for a given quantity of thermal input. The EISG CCVC testbed, with its very long recuperator and its long heater and cooler heat exchangers, indicated that quite clearly.

In other words, the use of a regenerator rather than a recuperator can improve thermal efficiency in an isochoric process. Note, however, that presently, heat addition and heat removal in Stirling engines still requires the use of heater and cooler heat exchangers, with a regenerator “sandwiched” in between. That added internal volume for the heater and cooler can therefore be seen as a negative for the potential thermal efficiency of isochoric heat input or removal, since increasing the overall volume of the connecting plenum that must be displaced through limits a rise in both pressure and temperature for a given isochoric thermal input.

However, as has been shown specifically in FIG. 6 of U.S. patent application Ser. No. 17/746,848, it is possible to construct a valved and/or ported regenerator that can intermittently pass a gas stream through from the hot end of a valved regenerator, switch the valves and/or ports, then intermittently pass a second gas stream from the cold end of a valved regenerator. In the fourth embodiment shown in U.S. patent application Ser. No. 17/746,848, it was proposed that a valved regenerator be used for a simple high temperature heat exchange rather than a recuperator. Note that a ported regenerator can also be used, especially where the pressure differences between the two streams is small, and in the application proposed in said fourth embodiment, pressure was equal for both gas streams. The advantage in said fourth embodiment to a “valve-switching regenerator” operated as a kind of constant pressure heat exchanger feeding into and out of a high temperature environment was that it was likely to increase the efficiency of the heat exchange process over a standard counter-flow heat exchanger.

Note, however, that a valve-switching regenerator can also heat and/or cool a volume of gas being passed through isochorically, although timing and pressure matching are critical, as will be shown. It is therefore proposed that, in addition to situating a regenerator between a heater and a cooler, as in a present Stirling engine, a CCVC engine be situated with a valved regenerator heater means and a valved regenerator cooler means, the intent of said means being to replace higher volume and less thermally efficient counter-flowing separate stream heat exchangers with much lower volume and more thermally efficient valved regenerators, thus increasing the CCVC engine's overall potential (Th—Tc)/Th. Note that this concept may be applied to existing Stirling engines as well to good effect.

Applying this idea to improving the existing EISG CCVC testbed, the small first displacer may (a) pass its working fluid first through a recuperator VDSR, (b) through a heater VSDSR's first (charged) regenerator core, and (c) into the small second displacer, thus raising the engine to it's peak temperature with a minimum of displaced fluid, thus increasing the amount of thermal energy transferred per displacement stroke, and thus raising the pressure over what would be possible with a standard heater heat exchanger. The small second displacer would then (d) reverse the flow, pass its working fluid back through the heater VSDSR's first (partially charge) regenerator core, through the transfer valve, and into the expander via the present process of displacement expansion, thus taking out work while adding some additional thermal energy to the working fluid throughout the said displacement expansion process. Note that while this occurs, the heater VSDSR's second regenerator core may be thermally charged via a flow of gaseous/vaporous fluid coming from and returning to a primary heat source. Immediately upon closure of the transfer valve, the charged and depleted regenerator cores would be valve-switched. Note that this could also be accomplished with more than two regenerator cores, permitting more charging time, if required.

As is illustrated in FIG. 6 of U.S. patent application Ser. No. 17/746,848, upon disconnection of the depleted regenerator core from direct contact with the engine and the simultaneous connection of a fully charged regenerator core in its place, a gaseous/vaporous heat transfer fluid, for example the engine's working fluid pressurized to the pressure of the engine working fluid following expansion, may be connected to the heater VSDSR's depleted regenerator core via intake and exhaust valves connecting and disconnecting said gaseous/vaporous heat transfer fluid, the flow of said gaseous/vaporous heat transfer fluid then being used to “charge” the heater VSDSR's depleted regenerator core with thermal energy by flowing from the heat source through the valved regenerator core then back to the heat source.

Thus, following each expansion piston intake stroke, the heater VSDSR's “thermally exhausted” regenerator core would be (e) “switched out” for a “thermally fully charged” regenerator core. And so on.

When exhausting from the CCVC expander, the exhausting working fluid may (f) pass through its SSR, then (g) potentially pass through a cooler, further cooling the gas isochorically during the displacement process from the expander to the displacer/compressor and reducing the temperature and the concomitant pressure. As mentioned above, the cooler, lower pressure fluid leaving the displacer/compressor does not necessarily need to be exhausted backwards through the cooler. Instead, the cooler can be bypassed via valving, and the exhaust from the displacer/expander can (h) pass directly back through the SSR. This permits the exhaust from the displacer/compressor to regain a maximum amount of waste heat as it passes back through the SSR, and reduces the cooling required by the cooler.

There is a further advantage. Bypassing the cooler during the exhaust stroke creates an opportunity for an intermittent charge of cooling gas to be passed through the valved cooling regenerator from the cold end to the hot end during the following move from TDC to BDC. Ergo, in the case of the cooler, only a single valved regenerator core may be required or cooler VSDR, rather than a series of “valve-switching” regenerator cores. However, the working fluid in the displacer/expander is at its lowest pressure at TDC, and will need to be reconnected when the displacer/expander is at its highest pressure at BDC. That would appear to require that the gaseous/vaporous heat transfer cooling fluid would also need to be at the highest pressure. As a result, while the valve connecting the cooler VSDR to the SSR could be a one-way check valve, the valve connecting the cooler VSDR to the displacer/compressor would need to resist cooler VSDR pressure while the cooler VSDR is being charged, necessitating an actuated valve. The inlet and exhaust valves connecting the heat transfer cooling fluid would transfer valves, similar to the expansion cylinder transfer valve. In addition, there would be a small recompression device attached to the cooler VSDR, since the volume of the cold side regenerator core will be small.

Therefore, the process for cooling with a cooler VSDR is as follows:

Said charge of engine working fluid thus exhausting backwards through the SSR may then exhaust through the recuperator VSDR.