Bland/Ewing Cycle Improvements

Description

CROSS-REFERENCE TO RELATED APPLICATIONS

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.

PRIOR ART

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.

BACKGROUND

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:

• • (1) An isochoric increase in temperature and concomitant increase in pressure.
  • (2) An isothermal expansion and concomitant decrease in pressure.
  • (3) An isochoric decrease in temperature and concomitant decrease in pressure.
  • (4) An isothermal compression and concomitant increase in pressure, which returns the vapor or gas to the state it started from.

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:

• • (1) As shown in the move from FIG. 1A to FIG. 1B, without moving the upper piston, moving the lower piston through its (ideally isothermal) compression stroke;
  • (2) as shown in the move from FIG. 1B to FIG. 1C, moving both pistons simultaneously and at the same rate over an equal distance in order to displace the gases through the cooler/regenerator/heater matrix at constant volume;
  • (3) as shown in the move from FIG. 1C to FIG. 1D, immediately pausing the lower piston while the upper piston undergoes an (ideally isothermal) expansion; and
  • (4) as shown in the move from FIG. 1D to FIG. 1A, cooling the gases at constant volume by movement of both pistons simultaneously and at the same rate over an equal distance in order to displace the gases through the heater/regenerator/cooler matrix at constant volume.

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.

The Valved Cell and Closed Cycle Valved Cell Concepts

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, FIG. 7 and FIG. 8 for a photograph and FIG. 9 for 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.

The EISG CCVC Testbed as a Hybrid Ericsson/Stirling Closed Cycle Heat Engine

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.

Ducted/Ported and Valved Regeneration

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 FIG. 6. A more prosaic schematic of a VSDSR is shown in U.S. patent application Ser. No. 18/362,951, FIG. 5a and FIG. 5b, 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 as FIG. 1 through FIG. 4 in 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, FIG. 6a and FIG. 6b.

Ducted/Valved Regeneration Applied to a CCVC Engine

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.

The Exhaust Recuperator-Regenerator Concept

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:

• • A. The modification to the existing EISG CCVC testbed is relatively simple.
  • B. For either the existing EISG CCVC testbed layout or the proposed modification to the EISG CCVC testbed layout, the exact same amount of work is required to both compress the working fluid within the displacer/compressor and to exhaust the pressurized working fluid out of the engine, since this second process occurs at the same exact pressure in both cases. Ergo, there is no additional thermodynamic cost to the change.
  • C. By exhausting back through (1) the cooler and (2) the recuperator, the exhaust portion of the recuperator is converted into a kind of SSR, since the exhausting working fluid will absorb some of the thermal energy within the expander exhaust regenerator half of the regenerator/recuperator that was deposited during the expander exhaust displacement stroke.
  • D. The exhausted working fluid will thus be raised in temperature but not necessarily in pressure during its final exhaust from the displacer/compressor.
  • E. As a result of the “regeneration” of heat, higher quality waste heat energy is made available for use, for example in otherwise improving the thermal efficiency and/or power density of the EISG CCVC testbed. Following such a use, the ultimately exhausted working fluid could be cooled and compressed.
    a CCVC Engine with Complete Isochoric Waste Heat Regeneration

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.

• • A. In the EISG CCVC prototype, at TDC, the pressure and temperature within the small second displacer, the source heater, the recuperator VDSR, and any connecting manifolding, is at a maximum. Following the opening of the transfer valve at approximately TDC, the various cylinders undergo a simultaneous action:
    • 1. The expander piston travels towards BDC, expanding the gas within the second small displacer, the heat source heater, and the various connecting manifolds. The small second displacer intake check valve immediately shuts, for example via a spring bias towards closed, as pressure equalizes across it.
    • 2. The small second displacer piston likewise begins to travel towards BDC, displacing the hot high pressure working fluid back through the heat source heater, past the transfer valve, and into the expander, dropping pressure.
    • 3. The small first displacer piston likewise begins to travel towards BDC, lowering the pressure in the regenerator and its manifolds through the still-open small first displacer actuated exhaust valve and thus in the small first displacer cylinder.
    • 4. The displacer/compressor piston likewise begins to travel towards BDC. The lower pressure, lower temperature working fluid following the displacement exhaust process begins to recompress backwards out of the displacer/compressor, closing the cooler intake check valve, for example via a spring bias towards closed.
  • B. As the displacer/compressor compression process continues, increasing pressure opens the displacer/compressor exhaust check valve and begins raising the pressure in the SSR and, via the cooler intake check valve, the cooler. Eventually, pressure equalizes across the recuperator VDSR hot side intake check valve, connecting the recuperator VDSR, and, via the still-open small first displacer actuated exhaust valve, the small displacer. Simultaneously, the small first displacer actuated exhaust valve is closed and the main engine exhaust valve is opened.
  • C. The expansion/displacement process continues between the small second displacer, the heater, and the expander, the refilling process continues in the small displacer, and the ideally constant pressure exhaust process continues from the displacer/compressor until BDC begins to be approached. This constant pressure exhaust process from the displacer/compressor, through the exhaust regenerator, through the recuperator VDSR, and to the radiator through the main engine exhaust valve serves to “charge” the recuperator VDSR with otherwise-waste thermal energy, preparing it for the next isochoric displacement of colder pressurized working fluid.
  • D. In the instance that the working fluid is exhausting at a lower pressure than that of the small first displacer working fluid, the main engine exhaust valve closes early enough to allow increasing compression in the displacer/compressor to build the pressure in the SSR, the recuperator VDSR, and the connecting manifolds to match the pressure in the small first displacer prior to BDC.
  • E. Just as BDC is reached, the displacer/compressor reaches the end of its exhaust stroke, the expander exhaust valve opens, the transfer valve closes, and the recuperator VDSR intake check valve is closed, for example via a spring bias towards closed. Ideally, the expander has expanded its charge of working fluid via the transfer valve to exactly the pressure in the regenerator, the displacer/compressor, the expander main exhaust manifold, and the small first displacer cylinder, such that pressures throughout the entire engine system equalize at the pressure of the new charge of pressurized working fluid within the small first displacer.
  • F. Following the closing of the transfer valve at approximately BDC, the various cylinders undergo a simultaneous action:
    • a. The expander piston travels towards TDC, exhausting the gas past the now-open exhaust valve, past the exhaust check valve, through the SSR, through the cooler, past the cooler exhaust check valve, and into the displacer/expander cylinder.
    • b. Simultaneously, the small first displacer piston actuated exhaust valve opens, and the small first displacer piston likewise begins to travel towards TDC, displacing its pressurized working fluid through the (reheated) recuperator VDSR, past the small second displacer intake check valve, through the heat source heater, and into the small second displacer piston cylinder, increasing the displaced working fluid's temperature and thus pressure in the process.
    • c. Simultaneously, the small second displacer piston receives the isochroically-heated pressurized working fluid.
    • d. Simultaneously, the displacer/compressor piston likewise begins to travel towards TDC, while it simultaneously receives the isochorically-cooled expander exhaust.
  • G. As the exhaust stroke is approaching completion at TDC, the expander exhaust valve is closed, allowing recompression of remnant gas in the expander to match the pressure of the working fluid now reaching maximum temperature and pressure on the other side of the transfer valve, thus allowing it to open and matching the system to step A. above, thus completing the cycle.

Advanced Fully-Regenerated Engines

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).

Regenerating Heaters and Regenerating Coolers

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, FIGS. 1A through 1D: 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:

• • A. At TDC, the intake valve from the cooler VSDR to the displacer/expander actuated valve is closed.
  • B. Following that closure, at or near BDC a small compressor attached to the hot side of the cooler VSDR quickly compresses a sufficient volume of fluid adiabatically into the cooler VSDR to the higher pressure of the cooling fluid. The compressor would probably be solenoid-actuated, but possibly cam-driven. The volume compressed would depend on the size of the cooler VSDR and the pressure differential created during the expander exhaust displacement process.
  • C. When the cooler VSDR internal pressure reaches the transfer cooling fluid pressure via the actions of the small compressor, the cooler VSDR cold side cooling fluid inlet transfer valve automatically opens, as by a spring bias, much like the expansion cylinder transfer valve.
  • D. When pressures equalize, the cooler VSDR hot side cooling fluid exhaust transfer valve also automatically opens, as by a spring bias, allow the cooling fluid flow through the cooler VSDR to commence.
  • E. Just before BDC, the intake and exhaust cooling fluid transfer valves are physically shut, probably by solenoid, but also possibly cam-driven.
  • F. The small compressor will be returned to its recharged state.
  • G. As the pressure difference increases approaching TDC again, the physical forces holding the intake and exhaust cooling fluid transfer valves closed are removed, and the valves are held closed by pressure differential. And so on, starting at A. above.

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

Such a CCVC isochoric engine will predictably greatly increase the thermal efficiency and the power density of the engine process over that of the EISG CVCC testbed.

A Rankine/Stirling CCVC Engine

In U.S. patent application Ser. No. 17/746,848, the Bland/Ewing chemo-thermodynamic cycle (B/E cycle) disclosed in U.S. Pat. No. 3,225,538 is modified to operate as a kind of B/E Rankine cycle. Applying that concept to the concept proposed by the modified EISG CCVC testbed suggests yet another interesting CCVC embodiment; a kind of combined Rankine/Stirling engine can be created by using the final exhaust from the EISG CCVC testbed to vaporize or partially vaporize a liquid working fluid prior to passing said vaporized working fluid through the modified EISG CCVC testbed. Note that, as in normal steam engines, the exhausted steam can then be cooled to liquid temperature and recycled. That in turn obviates any need for a gaseous compressor, while maintaining thermal efficiency by using otherwise-waste thermal energy for the vaporizer.

(Note: In the system described above under the heading “A CCVC engine with true isochoric waste heat regeneration”, a “compression” is proposed by the so-called displacer/compressor at step “A.d.” above. However, that small compression occurs only until pressures equalize between the exhausting working fluid in the displacer/compressor and the small first displacer, after which the exhaust process can be isobaric. As an alternative, the main engine may, following closure of the small first displacer actuated exhaust valve, be exhausted “explosively” to a lower pressure with the opening of the main engine exhaust valve, thus creating a lower pressure displacer/compressor isobaric exhaust stroke. Then, at the end of the displacer/compressor exhaust stroke, an early closure of the main engine exhaust valve can allow a recompression within the regenerator VDSR, by means of a recompression of remnant gas/vapor by the displacer/compressor into the regenerator VDSR, the expander exhaust SSR, and any connecting manifolds, back to the pressure within the small first displacer just prior to the opening of the small first displacer actuated exhaust valve. Note the similarity of this recompression stroke to the opening of the transfer valve and the recompression stroke that allows a “valved cell” to function efficiently.)

A Special Use Case Rankine/Stirling CCVC Engine

On the poles of Earth's Moon, special areas exist that never see the light of our Sun. These areas are Permanently Shadowed Regions (PSRs). It is also possible to artificially create shaded areas on the lunar surface that won't seen sunlight which can be called Artificially Shadowed Regions (ASRs). In the depths of some polar craters that never see sunlight, extremely low temperatures approaching 100 Kelvins. The existence of these pools of cold matter constitute a meaningful lunar resource for things like heat engines.

Some potential CCVC working fluids can be found to be extremely useful, thanks to the existence of PSRs and, potentially, of ASRs, and that would be working fluids that are very close to being a volatile vapor at the temperatures there.

One example is CO2. In a PSR, CO2 can be easily converted from a gas to a solid (so-called dry ice). CO2 sublimates into a solid or “boils” to a vapor at 1 atmosphere (14.7 psi) and 195 K (−109 deg F.). A CCVC CO2 working fluid can thus be pressurized to a higher pressure of, for example, 5 atmospheres (73 psi), by essentially placing CO2 in a “pressure cooker” and applying a temperature of, for example 283 K (50 deg F.) to it. That heated, pressurized CO2 can then be fed into a CCVC engine. Between isochoric regenerative waste heat recycling and isochoric heating with a heat source, the CO2 can be easily superheated and super-pressurized to a temperature of 1,000 deg F. (810 K), potentially creating diesel engine-like pressure for expansion in a fully regenerating CCVC heat engine with no compression phase.

Within the confines of a lunar PSR, ASR, or combined PSR and ASR, the exhausting working fluid will still contain enough waste heat to power the Rankine cycle-type vaporization of pressurized solid CO2, likely to a temperature of at least 50 deg F.

The theoretical Carnot thermal efficiency ((Th—Tc)/Th) of a heat engine with a source temperature of 810 K and a theoretical sink temperature of 195 K is ((810−283)/810=) 0.65 or 65%. Coupled with the extremely high delivered thermal efficiency of a compressor-less engine and the very low relative peak temperature, and it's likely that a delivered thermal efficiency could approach 50%. For solar energy, that means possibly 40% of the focused solar power can be converted into electricity, possibly matching if not exceeding the best that a solar-powered Stirling engine has ever achieved.

Open Cycle or Semi-Open Cycle CCVC-Based Internal Combustion Engines

One of the most interesting “valved cell” engine capabilities is its ability to operate as an open-cycle engine, thus easily permitting the use of internal combustion as a heat source. For an open cycle based on the EISG CCVC testbed, or Open Cycle Valved Cell (OCVC) isochoric fully regenerating heat engine, the working fluid can be compressed air, the heat input can be the internal combustion of a fuel, and the final exhaust can be vented to the atmosphere.

In a particular instance, the pressurized vapor can be steam. The steam-powered semi-open cycle CCVC engine would be manifested through application of the above-proposed modifications to the EISG CCVC testbed. Just prior to the pressurized steam being taken into the small first displacer, H2 at an equal pressure and temperature may be added, creating H2-enriched steam. Following the displacement of the H2-enriched steam working fluid from a small first displacer through a heat source and into a small second displacer of equal volume, an increase in pressure would be manifested by the isochoric raising of the temperature of the enriched mixture by the addition of otherwise-waste exhaust heat, as shown in the above-proposed modifications to the EISG CCVC testbed. At the beginning of and/or during the following expansion, a valved cell process may then add either preheated and compressed O2 or preheated and compressed H202 vapor to the H2-enriched steam, thus adding thermal energy to the CCVC engine by internal combustion. The exhaust of such an engine ideally being pure H2O, said exhaust may then be captured by cooling and condensing and eventually recycled by electrolysis back into H2 and O2. That is, in a semi-open cycle, the final exhaust can be ultimately recycled back to its original constituents.

Alternatively, just prior to preheated and pressurized steam being taken into the small first displacer, either O2 or H2O2 at an equal pressure and temperature may be added to the steam, creating either an O2-enriched steam mixture or an H2O2-enriched steam mixture. Following the displacement of the O2-enriched steam or H2O2-enriched steam working fluid from a small first displacer through a heat source and into a small second displacer of equal volume, an increase in pressure would be manifested by the isochoric raising of the temperature of the enriched mixture by the addition of otherwise-waste exhaust heat, as shown in the above-proposed modifications to the EISG CCVC testbed. At the beginning of and/or during the following expansion, a valved cell process may then add compressed and preheated H2 to the O2-enriched or H2O02-enriched steam, thus adding thermal energy to the CCVC engine by internal combustion. The exhaust of such an engine ideally being pure H2O, said exhaust may then be captured by cooling and condensing and eventually recycled by electrolysis back into H2 and O2. That is, in a semi-open cycle, the final exhaust can be ultimately recycled back to its original constituents.

The CCVC isochoric engine as a means of approaching a true Stirling cycle engine

As noted above, a true Stirling cycle requires:

• • (1) An isochoric increase in temperature and concomitant increase in pressure. This can be accomplished with a fully regenerating CCVC heat engine. Alternatively, it may be accomplished with a fully regenerating OCVC heat engine and internal combustion.
  • (2) An isothermal expansion and concomitant decrease in pressure. By using a fuel injector, for example a valved cell as was proposed in U.S. Pat. Nos. 4,817,388 and 5,179,839, heat from internal combustion can be added to the working fluid during an expansion stroke and arranged to maintain temperature during said expansion.
  • (3) An isochoric decrease in temperature and concomitant decrease in pressure. This is accomplishable with a fully regenerating CCVC heat engine.
  • (4) An isothermal compression and concomitant increase in pressure, which returns the vapor or gas to the state it started from. There are two possible approaches to achieving an isothermal compression. In the first, a series of compressions and inter-coolings, as in the well-known Ericsson cycle, will essentially “approach” an isothermal compression. In the second, it is possible to (a) exhaust a vaporous working fluid from the engine, cool it to its temperature of condensation, increase it's pressure with a pumping process, and (d) re-vaporize the working fluid. This is, of course, the well-known Rankine cycle. Note that a liquid may be compressed at essentially constant temperature, and that, once vaporized, the vapor can be increased in temperature isochorically. At lower pressures, the “line of vaporization” passing from liquid to vapor approximates an isothermal compression. Assuming otherwise-waste heat is available to supply the energy required for vaporization, there is little impact on such an engine's thermal efficiency.

The Advanced Fully-Regenerated CCVC Heat Engine as Applied to B/E Chemo/Thermodynamic Cycles.

A B/E heat engine cycle, by increasing the “endothermic fluid” (product) mol count for the expander to be higher than the “exothermic fluid” (reactant) mol count for the compressor, reduces the work put in by the compressor relative to the work put out by from the expander, known as the heat engine's mechanical efficiency. Since the compressor is usually the biggest part of the engine, that decreases many of the unavoidable losses in heat engines, potentially allowing the Wout (n)/Hin (t) of said heat engines to more closely approach the (Th—Tc)/Th. It is for this reason that it has been proposed in U.S. patent application Ser. No. 18/095,463 that a B/E cycle be separated into at least two half/cycle heat engines. The first heat engine is used to create stored Hin (t) as well as to generate Wout (n). The second heat engine is used to generate work from the release of the stored Hin (t). The overall efficiency is therefore the sum of the two Wout (n) and the two (Hin (t), with the summed quantities then used in the equation Wout (n)/Hin (t).

In U.S. Pat. No. 3,871,179, issued posthumously to Reginald B. Bland, the use of a B/E cycle within a Stirling cycle is considered, including the use of a regenerator. To effect the catalytic change required, a reducing catalyst heat exchanger and a reconstituting catalyst are proposed, with the reducing catalyst between the heater and the “hot” side of the Stirling engine, and the reconstituting catalyst between the cooler and the “cold’ side of the Stirling engine.

There are several problems with this approach, including a potential increase in the added internal volume by adding the catalytic chambers, although this might be avoided by somehow integrating the heater with the reducing catalyst and the cooler with the reconstituting catalyst. More importantly, however, in the classic B/E Brayton cycle, both reduction and reconstitution occur at constant pressure and constant temperature. As will be shown, in a true constant volume Stirling cycle, neither pressure nor temperature would be constant:

• • A. Assuming a product completely converted to reactant in the cold side cylinder, the piston would force colder working fluid through the reconstituting catalyst bed with no molecular change, through the regenerator, through the heater, and through the reducing catalyst bed with a potentially significant change in the number of molecules. As a result of the increase in the number of molecules, the pressure would increase. However, as is clearly shown in U.S. Pat. No. 3,225,538, FIG. 1 and FIG. 2, in order for the conversion to continue, with higher pressure, a higher source temperature is required, which will itself raise pressure in a constant volume environment, necessitating even higher temperatures, and so on.
  • B. Assuming a reactant completely converted to product, the process now reverses, the hot side piston would force hotter working fluid through the reducing catalyst bed with no molecular change, through the heater, through the regenerator, and through the reconstituting catalyst bed. As a result, the number of molecules would decrease, reducing the pressure. However, as is clearly shown in U.S. Pat. No. 3,225,538, FIG. 1 and FIG. 2, in order for the reconversion to continue, with lower pressure, a lower cooling temperature is required, which will itself reduce pressure in a constant volume environment, necessitating even lower temperatures, and so on.

It would therefore appear that a classic Stirling cycle is not compatible with a classic B/E cycle. Note, however, that a Stirling engine is not following a classic Stirling cycle, but is a mixture of adiabatic, isochoric, isobaric, and even isothermal processes. It is therefore possible that some amount of chemothermodynamic conversion can cyclically be made to occur in a Stirling engine. Note, however, that modern Stirling engines use a different approach to achieve useful power output, and that is the use of highly pressurized gaseous working fluids. Unfortunately, pressurizing a chemo/thermodynamic fluid such as is used in a B/E cycle will drive up the temperature required for conversion in either direction, and fairly quickly. On balance, a B/E-Stirling engine would add complexity and potentially increase internal volume, and not necessarily increase power density over that of the standard Stirling engine.

More importantly, there is the issue of the amount of thermal energy that would need to be chemo/thermodynamically transferred. In a closed cycle engine, a massive amount of energy will need to be transferred from a high temperature source into a lower temperature sink. As a result, the engine would almost certainly be less thermally efficient than existing Stirling engines.

Semi-Open Cycle CCVC-Based Combined B/E Half-Cycle Heat Engines

U.S. patent application Ser. No. 18/095,463 disclosed the concept of creating a unique full B/E chemo/thermodynamic cycle as disclosed in U.S. Pat. No. 3,225,538 by connecting two semi-open chemo/thermodynamic half-cycles; a semi-open endothermic half-cycle, and a semi-open exothermic half-cycle. The chemo/thermodynamic heat engine cycle analyzed herein is the classic cyclohexane <=>benzene+H2 cycle or C6H12<=>C6H6+3H2 cycle analyzed in U.S. Pat. No. 3,225,538. For the two half-cycles, the endothermic cycle is C6H12=>C6H6+3H2, and the exothermic cycle is C6H12<=C6H6+3H2. Note that in either direction, the exact same amount of thermal energy is stored as is released. However, the endothermic half-cycle occurs at a much higher temperature for a given pressure than the exothermic half-cycle. That is, the endothermic half-cycle requires higher quality heat than the exothermic half-cycle returns.

Both half-cycles have one purpose in common; generate net work out. But for the endothermic half-cycle, a second purpose is to store thermochemical energy, and for the exothermic half-cycle, it's main purpose is to use that stored energy efficiently to generate net work. Note that by far the largest amount of energy used to power the endothermic half-cycle is stored chemically. That in turn means that the largest amount of net work that will be produced independently by the two half-cycles will be produced by the exothermic half-cycle. It also means that maximizing the thermal efficiency of the exothermic half cycle is to maximizing the efficiency of both half-cycles combined.

A Rankin/Stirling Fully-Regenerating CCVC-Based Endothermic Semi-Open Half-Cycle Heat Engine

The C6H12=>C6H6+3H2 endothermic half-cycle has the ability to create a kind of chemical expansion process. Note that C6H12=>C6H6+3H2 literally turns a single molecule into 4 molecules; one molecule of benzene, and 3 molecules of hydrogen gas.

By pressurizing the C6H12 before it is converted into C6H6+3H2, 4 times as many molecules are generated at any given conversion pressure. The same amount of thermal energy is chemically stored, whether the conversion is at low pressure or high pressure. The only requirement is that, as the pressure increases, so does the temperature to feed the conversion. In other words, the C6H12=>C6H6+3H2 endothermic half-cycle may be seen as a kind of chemical H2 pressurization system.

Therefore, the approach being proposed for the endothermic half-cycle is to essentially use the endothermic conversion to produce pressurized H2 gas without a mechanical compressor. For example:

• • 1. Pressurize the liquid C6H12 reactant to some desired pressure.
  • 2. Heat the reactant to the vaporization point.
  • 3. Vaporize the reactant.
  • 4. Raise the pressure of the vaporized reactant slightly with a small amount of compression. The purpose of this, as will be shown, is to increase the temperature at which the eventual vapor product (C6H6) will condense over the temperature at which the C6H12 liquid at the pressure given it in step 1 above will evaporate. C6H12 and C6H6 have temperatures and heats of condensation/vaporization that are very close to one another (at 14.7 psi, C6H12 boils at 343.9 K with a standard heat of vaporization of 380 KJ/kg, while C6H6 boils at 353.2 K with a standard heat of vaporization of 433 KJ/kg). Consequently, condensing slightly higher pressure C6H6 will theoretically supply more than the required heat of vaporization of C6H12, removing the thermal cost of vaporizing the C6H12 from the required source heat in to the proposed process, as will be shown.
  • 5. Preheat the reactant by passing it through a first heat exchanger to the temperature at which the reactant, at the given pressure, will convert the reactant to product in an endothermic catalytic converter.
  • 6. Convert the vaporous C6H12 reactant at constant pressure and temperature into the product C6H6+3H2. Note that, at 100% conversion of C6H12 reactant to C6H6+3H2 product, 1,180 kJ of heat are absorbed chemically per 0.4536 kg (1 pound) of the resulting product. Note for comparison purposes that the vaporization requirement for about 0.45 kg of C6H12 equals (380×0.45=) about 170 kJ, or only about 14% of the required thermal input to create pressurized C6H6+3H2.
  • 7. Exhaust the product from the endothermic reactor and cool the product down with the first heat exchanger to just above the temperature at which the C6H6 will be condense into a liquid. The molar heat capacity of H2 is 28.84 Joules per an increase in temperature of 1 K (mol K). The molar heat capacity of vaporous C6H12 is 105 J/(mol K). The molar heat capacity of C6H6 is 135 J/(mol K). Thus, the molar heat capacity of three mols of H2 plus one mol of vaporous C6H6 equals ((28.84×3)+135=) 221.52 J. That is, the reactant has (135/221.52=) 61% of the heat capacity of the product. Assuming a perfect heat exchange, that will require passing only about 61% of the product through the first heat exchanger to preheat the C6H12 vapor to the temperature of the endothermic catalytic reactor.
  • 8. In a second heat exchanger, cool the 39% portion of C6H6+3H2 down with some process to just above the temperature at which the C6H6 will become a liquid.
  • 9. Recombine the ⅔ and ⅓ streams of C6H6 vapor+3H2 gas.
  • 10. Pass the relatively low temperature C6H6+3H2 product mixture through a fully-regenerating CCVC engine.
  • 11. For the source heat, use the heat separated out in the second heat exchanger. If desired, supply additional source heat to the C6H6+3H2 mixture to superheat it.
  • 12. After the C6H6 vapor and the H2 gas product mix is finally exhausted from the fully-regenerated CCVC engine, cool it down to the point where the C6H6 condenses, thus separating the product into liquid C6H6 and H2 gas.
  • 13. If possible, use any remaining latent heat in the CCVC engine exhaust to further reduce the heat requirements of the CCVC engine, for example, to supply the heat to increase the temperature of the pressurized C6H12 liquid reactant to just below the temperature required for vaporizing it at the given pressure.
  • 14. Based on only the latent heat requirements of the proposed endothermic heat engine, such a fully-regenerated CCVC engine can be expected produce delivered thermal efficiencies in the range of 50% or more.

A Rankin/Stirling Fully-Regenerating CCVC-Based Exothermic Semi-Open Half-Cycle Heat Engine

For a semi-open exothermic half-cycle, the primary purpose is to generate thermal energy to be used in a heat engine operating at maximum efficiency. Per U.S. Pat. No. 3,225,538, FIG. 1, at 1 atmosphere the conversion of C6H6+3H2 will occur at about 540 K (512 deg F.). That will produce, with negligible work in, 1,180 kJ of heat per 0.4536 kg (1 pound). In addition, in U.S. patent application Ser. No. 18/095,463, it is proposed that a small compressor and negligible work be applied to C6H6 vapor for a similar purpose as that proposed in step 4 above for the endothermic half-cycle engine. That is, it is proposed to supply much of the required heat of vaporization for the C6H6 by the condensation of higher-pressure C6H12. In other words, with such an approach, both the work in to produce heat at 540 K and the heat cost required are negligible compared to the fairly high temperature heat produced.

SUMMARY

This application proposes several new embodiments of a Closed Cycle Valved Cell (CCVC) heat engine, including means to create a fully-regenerated isochorically-heated CCVC heat engine. It further proposes the application of such a heat engine to Bland/Ewing (B/E) chemo/thermodynamic half-cycles.

BRIEF DESCRIPTION OF THE DRAWINGS

The invention will be illustrated in greater detail by description in connection with specific examples of the practice of it and by reference to the accompanying drawings, in which:

FIG. 1 is partially cutaway and partially transparent solid model of the existing EISG CCVC testbed. FIG. 1 shows the EISG CCVC testbed at Top Dead Center (TDC). The numbers under the boxed labels represent information found within corresponding boxed labels within FIG. 3

FIG. 2 is similar to FIG. 1 and shows the EISG CCVC testbed at Bottom Dead Center (BDC).

FIG. 3 is a schematic of the processes that occur within the EISG CCVC testbed. The numbers under the boxed labels represent the various means employed with the EISG CCVC testbed. The prime mover can be seen as the standard crankcase and single-cycle camshaft below the engine in FIG. 1, FIG. 2, and elsewhere in other figures.

FIG. 4 is a proposed modified version of the EISG CCVC testbed. FIG. 4 shows the EISG CCVC testbed at Top Dead Center (TDC). The numbers under the boxed labels represent additional information found within corresponding boxed labels within FIG. 6 and FIG. 7.

FIG. 5 is similar to FIG. 4 and shows the modified EISG CCVC testbed at Bottom Dead Center (BDC).

FIG. 6 is a schematic of the processes that occur within a modified version of the EISG CCVC testbed. The numbers under the boxed labels represent the various means employed within a modified version of the EISG CCVC testbed.

FIG. 7 is a schematic of the processes that occur within another modified version of the EISG CCVC testbed. The numbers under the boxed labels represent the various means employed within another modified version of the EISG CCVC testbed.

FIG. 8 is another proposed modified version of the EISG CCVC testbed. FIG. 8 shows the EISG CCVC testbed at Top Dead Center (TDC). The numbers under the boxed labels represent additional information found within corresponding boxed labels within FIG. 12, FIG. 13, and FIG. 14.

FIG. 9 is a rotated closeup of the modified EISG CCVC testbed shown in FIG. 8.

FIG. 10 is similar to FIG. 8 and shows the modified EISG CCVC testbed at Bottom Dead Center (BDC).

FIG. 11 is a rotated closeup of the modified EISG CCVC testbed shown in FIG. 10.

FIG. 12 is a schematic of the processes that occur within another modified version of the EISG CCVC testbed. The numbers under the boxed labels represent the various means employed within a modified version of the EISG CCVC testbed.

FIG. 13 is a schematic of the processes that occur within another modified version of the EISG CCVC testbed. The numbers under the boxed labels represent the various means employed within a modified version of the EISG CCVC testbed.

FIG. 14 is a schematic of the processes that occur within another modified version of the EISG CCVC testbed. The numbers under the boxed labels represent the various means employed within a modified version of the EISG CCVC testbed.

DETAILED DESCRIPTIONS

Description-First Embodiment

This is a description of the unmodified EISG CCVC testbed, aka hybrid Ericsson/Stirling closed cycle heat engine. For reference, see paragraph 12 and 13 above and FIG. 1, FIG. 2, and FIG. 3 of the drawings.

The EISG CCVC testbed was designed to be a closed system with a sealed and pre-pressurized working fluid, similar to the standard sealed and pre-pressurized working fluid for Stirling engines. In the EISG CCVC testbed test cases, nitrogen (N2) was used as the working fluid. Pre-pressurization of the testbed was held to about 90 psi, with peak measured pressure of about 190 psi. Peak temperature was about 511 K. Heat energy was supplied by an electric coil heater and an electric cartridge heater physically in contact with the heater heat exchanger. Cooling was supplied by water flowed through coils of copper tubing exterior to and interior to the cooler heat exchanger and exterior to the walls of displacer #1 and the combination displacer #3/compressor.

The piston is a single contiguous, hollow aluminum 2″ diameter tubular piston “rod” with a 2.5″ diameter “pancake” piston on either end. The piston is cyclically driven through a 2.75″ stroke by a prime mover below the engine proper. A ⅜″ diameter titanium rod connects the bottom of the contiguous piston to the prime mover. A 2.5″ outer diameter teflon-and-stainless spring steel seal is mounted in each of the 2.5″ diameter pancake pistons, and a single 2″ inner diameter teflon-and-stainless steel spring seal is mounted in a 2″ diameter tube between displacer cylinder #1 and the displacer cylinder #2. The teflon seals are not lubricated and ride on a vapor-deposited nickel substrate that is coating the single 2″ diameter piston connecting rod and the two 2.5″ diameter cylinders. Excluding manifold and heat exchanger spaces, the four internal spaces of the engine are comprised of the two 2.5″ diameter pancake pistons with piston seals, the 2″ diameter connecting rod, the 2″ diameter cylinder with 2″ diameter rod seal, the two 2.5″ diameter cylinders, and the ⅜″ diameter connecting rod, thus forming the 13.5 cubic inch expander space, the 13.2 cubic inch displacer #3/compressor space, and the 4.86 cubic inch #1 and #2 displacer spaces. The piston is driven by the single converted crankshaft with a 2.75″ throw from a converted single-cylinder four-stroke gasoline prime mover with an oil-lubricated crankcase. The crankshaft is contiguous with a modified single-lobe camshaft, which pushes a cam follower and connecting rod to operate a rocker arm mounted on the expander cylinder head, which opens and closes the poppet-type exhaust valve. The exhaust valve is spring-biased towards closed and opens at or just prior to the upstroke engine's upstroke at BDC and closes just prior to the downstroke at TDC. An arm mounted just below a teflon-and-stainless steel spring seal exhaust valve stem seal extends perpendicularly from the exhaust valve stem. The exhaust valve stem-mounted arm is used to (A) physically close the transfer valve and (B) compress the transfer valve opening spring when the exhaust valve closes as BDC is approached. The reverse poppet-type transfer valve stem has an o-ring seal. The transfer valve stem and and o-ring seal are oil-cooled and oil-lubricated with lower pressure oil, as is the exhaust valve stem. The upper portion of the transfer valve head is designed to seal against an upper seat and thus physically protect the transfer valve stem o-ring seal during the passage of the higher pressure hot gases flowing past the transfer valve head and into the expander. The transfer valve head upper seat is also partially oil-cooled, thus providing partial cooling of the transfer valve head. Also, the transfer valve stem is purposefully designed with a slight narrowing in the area where the transfer valve stem contacts the transfer valve o-ring seal during the travel towards opening of the transfer valve, temporarily reducing friction between the stem and the seal for that fractional moment and allowing the transfer valve head to adequately seal against the upper transfer valve head upper seat. Finally, the transfer valve head is provided with a slight physical projection into the expansion chamber, which allows it to operate similarly to that of a steam engine “bash-valve”, thus ensuring that the transfer valve does in fact break free and open at TDC. A second teflon-and-stainless steel spring seal exhaust valve stem seal ensures that the pressurized oil will not contaminate the higher pressure working fluid.

The converted prime mover's connecting rod is stock, but the original piston has been replaced with a titanium “follower” device and “force plates” mounted in the piston cylinder that converts the rotary motion of the crankshaft into linear motion. The ⅜″ diameter titanium rod passes through a teflon-and-stainless steel spring seal at the top of the prime mover that can seal against oil on the one side and substantial pressure on the other. The ⅜″ diameter titanium rod is attached on the lower end to the titanium follower device and on the upper end to the single contiguous piston, which is thus cyclically driven linearly from TDC to BDC.

All “coiled tube heat exchangers” used within the EISG CCVC testbed use lathe-machined multi-spiraled helical ribbing constructed in the manner illustrated and described in U.S. patent application Ser. No. 18/362,951, FIG. 9 and FIG. 10.

Operation—First Embodiment

Following the flow process illustrated in FIG. 1, FIG. 2, and FIG. 3, starting at BDC in the engine (shown in FIG. 2), ideally all the gases within the engine are at the same pressure, as will be shown, which is the pressure of the feed gas exiting the gaseous radiator (boxed label 1 in FIG. 1 and FIG. 3). A full pressurized charge of working fluid has just passed through the displacer #1 intake check valve (boxed label 2 in FIG. 1 and FIG. 3) and been taken into displacer #1 (boxed label 3 in FIG. 1 and FIG. 3). During the move in the engine to TDC (shown in FIG. 1), the pressurized charge of working fluid is displaced through the displacer #1 exhaust check valve (boxed label 4 in FIG. 1 and FIG. 3), through the heat input side of the coiled tube exhaust heat exchanger (boxed label 5 in FIG. 1 and FIG. 3), through the electrically heated coiled tube heater (boxed label 6 in FIG. 1 and FIG. 3), and into displacer #2 (boxed label 7 in FIG. 1 and FIG. 3). This is an isochoric (constant volume) process that creates an increase in the temperature and thus pressure of the working fluid thus “captured” at constant volume between displacer #1 (3) and displacer #2 (7).

Simultaneously at BDC, the expander transfer/intake valve (boxed label 8 in FIG. 1 and FIG. 3) is physically closed by the closing action of the exhaust valve's (boxed label 10 in FIG. 1 and FIG. 3) stem-mounted arm, compressing the transfer valve's (8) stainless steel spring (not shown) which is biased towards open. During the following engine move to TDC, the hot, low pressure charge in the expander (boxed label 9 in FIG. 1 and FIG. 3) is displaced past the now-open expander exhaust valve (10), forcing open and displacing past the expander check valve (boxed label 11 in FIG. 1 and FIG. 3), which is biased closed by a stainless steel spring (not shown), displaced through the heat output side of the coiled tube exhaust heat exchanger (5), displaced through the water-cooled coiled tube cooler (boxed label 12 in FIG. 1 and FIG. 3), and finally displaced into the combination displacer #3/compressor (boxed label 13 in FIG. 1 and FIG. 3). This is an isochoric process that creates a decrease in temperature and thus pressure of the working fluid thus “captured” at essentially constant volume between the expander (9) and the combination displacer #3/compressor (13).

Starting slightly before TDC is reached, the expander exhaust valve (10) is closed. The transfer valve's (8) stainless steel spring (biased towards open) then propels it open, thus connecting the expander (9) to the hot, high pressure working fluid captured between the displacer #1 exhaust check valve (4) and the expander exhaust valve (10).

Starting at TDC, displacer #2 (7) displaces hot, high pressure working fluid back through the heater (6), past the transfer valve (8), and into the expander (9). Simultaneously, the pressure drops due to expansion within the expander (9), overcoming some working fluid expansion via the heater (6) as the fluid from displacer #2 (7) flows through it. Some additional flow (and some additional working fluid expansion) also proceeds as pressure drops within the various connecting manifolds (not shown) and within the heat input side of the coiled tube exhaust heat exchanger (5).

Simultaneously at TDC, the displacer #1 exhaust check valve (4), biased towards closed by a stainless steel spring (not shown), closes as the pressure is equalized across it. The remnant gas still trapped in displacer #1 (3) expands, eventually equalizing pressure across the displacer #1 intake check valve (2) and overcoming that valve's stainless steel spring (not shown) bias towards closed. Consequently, relatively cold ideally constant pressure working fluid flows into displacer #1 (3), eventually fully charging it by BDC, at which time the displacer #1 intake check valve (2) will close.

Finally, near TDC, as noted above, the expander exhaust valve (10) will be closed, freeing the transfer valve to open when the pressure differential is equal on both sides of the transfer valve (8), thus permitting the stainless steel spring bias to open the transfer valve (8). Simultaneous to the closing of the expander exhaust valve (10), the expander check valve (11) will be closed due to its stainless steel spring bias towards closed. The displacer #3/compressor (13) will now begin compressing the colder, lower pressure working fluid backwards into the volume between the displacer #3/compressor (13) the expander check valve (11), and the compressor check valve (boxed label 14 in FIG. 1 and FIG. 3). Eventually, the pressure inside that volume will equalize to the pressure in the gaseous radiator (1). The compressor check valve (14) will then open as its stainless steel spring (not shown) bias towards closed is overcome. The displacer #3/compressor (13) will then begin exhausting re-pressurized working fluid past the compressor check valve (14), through the gaseous radiator (1), past the displacer #1 intake check valve (2), and into displacer #1 (3), ideally at constant pressure.

And so on, beginning again at paragraph 66 above.

Description—Second Embodiment

Modifications to the EISG CCVC testbed, aka hybrid Ericsson/Stirling closed cycle heat engine are herein proposed that will increase the testbed's potential and delivered thermal efficiency. These modifications will leave much of the EISG CCVC testbed unmodified, and may thus be seen both as a practical means for testing the usefulness of the modifications prior to the development of eventual production machinery and as a means for describing and detailing herein the proposed modifications. Changes beyond the First Embodiment labels will therefore be limited, and will be described below. Where no changes are incurred, FIG. 1 and FIG. 3 are referred to for unchanged labeling.

Replacing a Part of the EISG CCVC Testbed Recuperation Process with Regeneration

The existing EISG CCVC testbed uses a device called a “recuperator” to capture otherwise-waste exhaust heat. There is an alternative to recuperation called regeneration. A regenerator is advantaged over a recuperator in its ability to permit an exchange of heat with both a greatly-reduced internal volume and a much faster heat exchange rate. Decreasing internal volume for isochoric heat transfers will increase the theoretical thermal efficiency of a heat engine. A faster heat exchange rate will increase the rpm for delivering power output and thus increase the power output for a given engine mass.

Two basic types of regeneration are presently utilized. The first may be termed “single stream regeneration” (SSR). The second may be termed “dual-stream regeneration” (DSR).

DSR is generally accomplished by either switching the two counter-flowing streams through a single regenerator core, which is accomplished by either ducts/ports, which may be termed a “ducted DSR” or DDSR, or by valves, which may be termed a “valved DSR” or VDSR.

In addition, there is a variant in which may be termed a “valve-switched DSR” (VSDSR) or possibly a “duct/port-switched DSR” (DSDSR). In such a variant, there are at least two and possibly more regenerator cores which are “switched out” by ducts/ports such that one regenerator core or set of regenerator cores is “thermally charged” while the other regenerator core or set of regenerator cores is “thermally depleted”. One design for a possible VSDSR was proposed in U.S. patent application Ser. No. 17/746,848, FIG. 6.

As a means for partially reducing the volume of the present EISG CCVC testbed recuperator and thus increasing the potential thermal efficiency, a kind of combined exhaust heat SSR/cooler is herein proposed.

Replacing the Hybrid Ericsson/Stirling Closed Cycle Heat Engine with a Hybrid Rankine/Stirling Closed Cycle Heat Engine

A unique characteristic of the valved cell process that underlies the EISG CCVC testbed is the ability to connect previously disconnected processes. An Ericsson-like or a Rankine-like process may be integrated with the Stirling-like processes of the EISG CCVC testbed, creating a hybrid Ericsson/Stirling engine that uses constant volume regeneration or a hybrid Rankine/Stirling engine that will increase the net work delivered by the expander. In the case of the proposed modified EISG CCVC testbed, the existing combined displacer/compressor can be converted to a displacer only. That in turn increases delivered thermal efficiency.

Note that the two modifications included in the proposed second embodiment are not mutually exclusive.

Operation—Second Embodiment

For a hybrid Rankine/Stirling closed cycle engine, following the flow process illustrated in FIG. 1 through FIG. 7, and starting again at BDC in the engine (FIG. 5), it is assumed that the pressure and temperature of the feed vapor exiting the steam generator (boxed label 15 in FIG. 4 and FIG. 6, which replaces boxed label 1) is the same as for the gaseous radiator (1). A full pressurized charge of working fluid has just passed through the displacer #1 intake check valve (2) and been taken into displacer #1 (3). During the move in the engine to TDC, the pressurized charge of working fluid is displaced through the displacer #1 exhaust check valve (4), through the heat input side of the coiled tube exhaust heat exchanger (5), through the electrically heated coiled tube heater (6), and into displacer #2 (7). However there has been a change in the source of heat for the exhaust heat exchanger (5). Instead of the source of exhaust heat coming directly from the exhaust of the expander (9) as it passes through to the displacer #3/compressor (13), the source of heat derives from an SSR/cooler (boxed label 16 in FIG. 4 and FIG. 6, which replaces boxed label 12). In addition, the displacer #3/compressor (13) has been replaced by displacer #3 (boxed label 17 in FIG. 4 and FIG. 6, which replaces boxed label 13), with the SSR/cooler (16) sitting between displacer #3 (17) and the ECV (11). Finally, a new valve, called a “main engine exhaust valve” (boxed label 18 in FIG. 4 and FIG. 6) has been added to the EISG CCVC testbed. The new main engine exhaust valve (18), being between the the expander check valve (11) and the new SSR/cooler (16), also subtly changes the “plumbing” between the displacer #1 exhaust check valve (4) and displacer #2 (7), which no longer shows a route between the two that also passes through the heater (6). Instead, the hot exhaust working fluid exhausts through the main engine exhaust valve (18), and through the heater (6) before entering the exhaust heat exchanger (5). This subtle change “shortens the path” of the working fluid displacing from displacer #1 exhaust check valve (4) to displacer #2 (7), reducing the volume between and thus increasing the potential isochoric thermal increase in temperature and pressure.

After passing through the repositioned exhaust heat exchanger (5), the exhaust gases may still be of sufficient heat content to aid in the generation of steam vapor, shown by the exhaust gases being passed through said steam generator (15). The further cooled exhaust gases are now passed through a condenser (boxed label 19 in FIG. 4 and FIG. 6), into a fluid storage tank (boxed label 20 in FIG. 4 and FIG. 6), then through a liquid pump (boxed label 21 in FIG. 4 and FIG. 6). And so on.

FIG. 7 indicates an alternative place for the main engine exhaust valve (boxed label 22 in FIG. 7, which replaces boxed label 18).

For an hybrid Ericsson/Stirling closed cycle engine, substitute a gaseous radiator (1) for the condenser (19) and a compressor (13) for the liquid pump (21).

Description—Third Embodiment

Modifications to the EISG CCVC testbed are herein proposed that will increase the testbed's potential and delivered thermal efficiency. These modifications will leave much of the EISG CCVC testbed unmodified, and may thus be seen both as a practical means for testing the usefulness of the modifications prior to the development of eventual production machinery and as a means for describing and detailing herein the proposed modifications. Changes beyond the First and Second Embodiment labels will therefore be limited, and will be described below. Where no changes are incurred, FIG. 1, FIG. 3, FIG. 4, FIG. 6, and FIG. 7 are referred to for unchanged labeling.

Replacing the EISG CCVC Testbed Exhaust Heat Exchanger with a VDSR

The existing EISG CCVC testbed uses a recuperator to exchange heat isochorically between the expander exhaust and the displacer #1 (3) to displacer #2 (7) displacement. Even in the proposed Second Embodiment, a recuperator (5) was still show. However, by replacing the recuperator (5) with an SSR/cooler combination (16), as shown in the Second Embodiment, the syncopation between the two isentropic displacement processes found in the existing EISG CCVC testbed is changed dramatically, such that the final exhaust from displacer #3 (17) now occurs out of phase with the displacement process between displacer #1 (3) and displacer #2 (7).

That means it is now possible, as stated above, to tap into the waste heat flowing out of the SSR to directly thermally charge a single regenerator core “valved dual-stream regenerator”, or VDSR heat exchanger, greatly simplifying the process of converting the EISG CCVC prototype to replace the existing recuperator (5) with (a) an exhaust gas regenerator and (b) an isochoric displacement regenerator. That in turn will greatly reduce the “dead space” volumes within the EISG CCVC prototype, further increasing the potential thermal efficiency.

Exhaust Cylinder Displacement Versus Exhaust Compounding.

In replacing the gaseous compression process with the Rankine process, it becomes possible to extend the expansion process a great deal over the standard Stirling expansion process. One means for accomplishing this is via exhaust compound-expansion. For example, the present EISB CCVC testbed expansion chamber, as mentioned above equals 13.5 cubic inches, and the displacer volumes equals the 4.86 cubic inches. Even assuming dead space, that would limit the expansion ratio to 13.5/4.86 or 2.78X.

If the “displacer” were replaced with a 27 cubic inch expander, then the expansion ratio would potentially double to 5.5X, creating significantly more net power output per crank cycle. That would, of course, also significantly lower the post-expansion exhaust pressure. However, with the SSR exhaust system, a great deal of the thermal content “left behind” in the regenerator during the exhaust stroke can be recovered in the following secondary exhaust stroke, allowing for both additional net power out but also a significant amount of waste heat recovery. Finally, the “exhaust resistance” of pumping the expanded and thus lower pressure gases out of the 2nd expander would mean less exhaust work in is required, also aiding the net work out per cycle.

It is also possible for an isochoric displacer #3 (17) to exhaust to a compounding expander as an alternative to changing out displacer #3 (17) for an expander. Again, this is made easier when the modified EISG CCVC testbed uses the proposed modified Rankine process.

Replacing the EISG CCVC Testbed Heater with a VSDSR Heater

The existing EISG CCVC testbed uses a recuperator heater (6) to add source heat to the engine during isochoric heating. It is herein proposed that a “valve-switching dual-stream regenerator (VSDSR) heater be used in the EISG CCVC testbed to replace the existing recuperator heater (6).

Replacing the EISG CCVC Testbed Cooler with a VDSR Cooler

The existing EISG CCVC testbed uses a recuperator cooler (12, 16) to remove heat from the engine during isochoric cooling. It is herein proposed that a “valved dual-stream regenerator (VDSR) cooler be used in the EISG CCVC testbed to replace the existing recuperator cooler (12, 16).

These changes “shorten the paths” of the working fluids between (a) displacer #1 (3) and displacer #2 (7) and (b) the expander (9) and displacer #3 (17) or expander #2 (30), reducing the volume between them and thus increasing the potential isochoric thermal changes in temperature and concomitant pressure.

Operation—Third Embodiment

Following the flow process illustrated in FIG. 1 through FIG. 7, and starting again at BDC in the engine (FIG. 5), a full pressurized charge of working fluid has just passed through the displacer #1 intake check valve (2) and been taken into displacer #1 (3). Looking backwards to the previous move from TDC to BDC, it is assumed to be the case that the pressure within the expander (9) was significantly above the pressure within the steam generator (15) at TDC, and it is likewise assumed to be the case that, when the expander exhaust valve (10) opens at BDC, the pressure within the expander (9) is at about the same pressure as the pressure within the steam generator (15).

That being the case, if an isochoric displacement from BDC to TDC occurs through the SSR/cooler (16) and into displacer #3 (17), then the pressure will be significantly lower than the pressure in the steam generator (15) when displacer #3 (17) almost reaches TDC and the expander exhaust valve (11) closes slightly early.

However, because the engine is now using a pressurized liquid-to-pressurized-vapor combined Rankine/Stirling process, it is perfectly permissible for the newly-added main engine exhaust valve (18) to open and release the working fluid either at that lower pressure or even permit some “explosive decompression” through the air engine exhaust valve (18, 22) to an even lower pressure. Assuming that to be the case, displacer #3 (17) now exhausts at either the pressure following displacement or some lower pressure.

Exhaust Cylinder Displacement Versus Exhaust Compounding.

Likewise, it is perfectly permissible that, instead of a pure isochoric displacement process, an exhaust gas compound expansion process is created by, essentially, replacing displacer #3 (13) with a second expander (boxed label 30 in FIG. 8, FIG. 12 and FIG. 13, replacing boxed labels 13 and 17) takes place that lowers the gas pressure even more, as mentioned above.

Note that, per FIG. 7, in either case the exhaust flows back through the SSR/cooler (16), back through the heater (6), and through the exhaust heat exchanger (regenerator) (5), where it exchanges heat with the cool higher pressure gas on the other side of the recuperator (5) flowing between displacer #1 (4) and displacer #2 (7).

Replacing the EISG CCVC Testbed Exhaust Heat Exchanger with a VDSR

In FIG. 8 and FIG. 12, however, something else has been added: The recuperator (5) is replaced with a “valved dual stream regenerator” or VDSR heat exchanger (boxed label 24 in FIG. 8, FIG. 12, FIG. 13, and FIG. 14, replacing boxed label 5). Note that the main engine exhaust valve (boxed label 32 in FIG. 8, FIG. 12, FIG. 13, and FIG. 14, that replaces boxed label 18 and boxed label 22) is shown on the “cold side” of the VDSR heat exchanger (24). That means that, if desired, the main engine exhaust valve (32) can be made to exhaust the gases at TDC upstream of expander check valve or ECV #2 (boxed label 31 in FIG. 8, FIG. 12, and FIG. 13), and ECV #5 (boxed label 40 in FIG. 14) into a much lower pressure environment, including that of ambient atmospheric pressure.

This may require some “explosive decompression” of trapped working fluid in the VDSR heat exchanger (24), for the following reason: As previously stated, the highest temperature and concomitant pressure occurs at approximately TDC at the end of the isochoric displacement process between displacer #1 (3) and displacer #2 (7). Note the addition in FIG. 8 and FIG. 12 of a “displacer #2 intake check valve (boxed label 23 in FIG. 8, FIG. 12, FIG. 13, and FIG. 14) which is biased towards closed by a stainless steel spring (not shown). That implies that, at TDC, the displacer #2 intake check valve (23) will close. Also at TDC, the main engine exhaust valve (32) will be closed, as will ECV #2 (31) or ECV #5 (40). If the displacer #1 actuated exhaust valve (23) is also closed, then high pressure gas will be effectively trapped inside the VDSR heat exchanger (24).

Hence, displacer #1's exhaust valve (23) has been changed from a check valve (3) to permit a reverse expansion of gases trapped in the VDSR heat exchanger (24) back into displacer #1 (3). That is also why it is essential that the VDSR heat exchanger (24) be as compact as possible: The larger the internal volume of the VDSR heat exchanger (24), the more re-expansion of its trapped high pressure gas back into displacer #1 (3) will be required. That in turn will impact the amount of charge that the engine can receive into displacer #1 (3) during its intake stroke, which in turn will either increase or reduce the molal count of the gas that circulates per cycle, which in turn will increase or reduce the power density that the modified EISG CCVC testbed can achieve per cycle, which may positively or negatively impact potential thermal efficiency.

Determining the optimal balance that will generate maximum thermal efficiency for the modified EISG CCVC testbed between re-expansion back into displacer #1 (3) of trapped VDSR heat exchanger (24) gas versus “explosive decompression” out of the VDSR heat exchanger (24) via the main engine exhaust valve (32) will need to be determined experimentally.

Starting once again at BDC in the engine (FIG. 10), a full pressurized charge of working fluid has just passed through the displacer #1 intake check valve (2) and been taken into displacer #1 (3). The displacer #1 activated exhaust valve, being closed by a biased closed stainless steel spring (not shown), can maintain pressure substantially higher than exists within the recently-thermally-charged VDSR heat exchanger (24). However, to match pressure within the regenerator and within the volumes between displacer #3 or expander #2 (30) and ECV #1 (11), the main engine exhaust valve is closed early, permitting some recompression into said volumes between displacer #3 or expander #2 (30) and ECV #1 (11) to match the pressure within displacer #1 (3), as well as the pressure within the expander (9) and the volume between the expander (9) and the displacer #2 intake check valve (25). At BDC, therefore, pressure has been ideally equalized across all internal spaces within the modified EISG CCVC testbed. Note that the transfer valve (8) has just been closed as the expander exhaust valve (10) has opened.

Replacing the EISG CCVC Testbed Heater with a VSDSR Heater

In FIG. 12, a constant pressure heat transfer vapor system, which is a pressurized quantity of the same vapor being used in the engine and at the same pressure as the vapor in the engine at BDC, is connected to a thermally discharged regenerator core of a VSDSR heater (boxed label 26 in FIG. 8, FIG. 12, FIG. 13, and FIG. 14, replacing the heater 6, and referencing for an example of a VSDSR device U.S. patent application Ser. No. 17/746,848, FIG. 6). Starting at the previous BDC of the engine, a heat source (boxed label 27 in FIG. 8, FIG. 12, FIG. 13, and FIG. 14) begins thermally charging said thermally discharged regenerator core by (A1) flowing at constant pressure said high temperature constant pressure heat transfer fluid (A2) from said heat source (27), through a heat transfer fluid intake valve connected to the hot side of said thermally discharged regenerator core, (A3) through said thermally discharged regenerator core, (A4) through a heat transfer fluid exhaust valve attached to said regenerator core's cold side, and (A5) back to said heat source be reheated (27).

At the following TDC of the engine, (A6) the now thermally-recharged regenerator valve-switches to disconnect from the constant pressure heat transfer system (27) and connects to ECV #2 (31). As it moves away from TDC, displacer #3 (30) begins recompressing the working fluid within the volumes between it, ECV #1 (11), and ECV #2 (31), while simultaneously the VSDR head exchanger (24) drops in pressure either due to expansion back into displacer #1 (3) via the displacer #1 actuated exhaust valve (23) or by explosive decompression from opening of the main engine exhaust valve (32) or both. When pressures equalize, displacer #3 (30) begins exhausting back through the cooler (29), back through the exhaust SSR (28), past ECV #2 (31) through the thermally-recharged regenerator of the VSDSR heater (26), through the VSDR heat exchanger (24), and out of the main engine exhaust valve (32), eventually reaching BDC in the engine.

Meanwhile, simultaneous to the move from the previous BDC to the TDC of the engine, with the main engine exhaust valve (32) closed, the ECV #2 (31) closed, the displacer #2 intake check valve (25) about to open, the displacer #1 actuated exhaust valve (23) about to open, and the pressures all the same throughout the engine, it is now possible to simultaneously valve-switch out yet another thermally discharged regenerator core, this one for use following the isochoric displacement from displacer #1 (3) into displacer #2 (7), finishing at TDC. Further travel from BDC towards TDC will initiate an isochoric heat input process as working fluid is passed through the VDSR heat exchanger (24) fully charged regenerator core. The second charged regenerator core would then be available by valve-switching to use during the displacement/expansion process, where working fluid is displaced out of displacer #2 (7), through the cold side and out the hot side of the second charged regenerator core (26), past the transfer valve (8) and into the expander (9). In other words, the valve-switching would involve swapping four regenerator cores at BDC, and the VSDSR heater (26) would thus be made up of those four cores, their switching valves, and connecting manifolds.

In FIG. 13, an alternative is shown where heater #2 (boxed label 33 in FIG. 8 and FIG. 13) serves to supply source heat to the working fluid during the isochoric displacement from displacer #1 (3) to displacer #2 (7), and the first heater is used only to supply heat to the working fluid during the displacement-expansion process from displacer #2 (7) through the transfer valve (8) and into the expander (9). Note that heater #2 (33) can be a second VSDSR heater.

FIG. 12 and FIG. 13 differ primarily in the pathway for adding source heat to the proposed modified EISG CCVC testbed. An advantage of the FIG. 12 system is that the VDSR heat exchanger (24) only needs to transfer otherwise/waste exhaust heat. Also, the displacer actuated exhaust valve (23) will be subjected to lower temperature heat. Also, the exhaust gas will not have to undergo a recompression to equalize pressure. The disadvantage is the addition an additional VSDSR heater's volume. However, this advantage may disappear if post-displacement heating is done differently, for example by using internal combustion.

Replacing the EISG CCVC Testbed Cooler with a VDSR Cooler

Further travel from BDC towards TDC will likewise initiate an isochoric heat removal process as working fluid is passed through the exhaust SSR (28) and VSDR cooler (boxed label 35 in FIG. 8 and FIG. 14, replacing boxed label 12 and 29), as shown in the proposed modified EISG CCVC testbed move from FIG. 8 through FIG. 11 and as shown in the schematic of the processes in FIG. 14.

Proceeding on with the flow process, FIG. 8 through FIG. 11 and FIG. 14 assume the exhaust process is an isochoric displacement process. Consequently, the desire is to reach the lowest possible temperature by the end of the process with the least amount of “dead space”, i.e., space that isn't found in either the expander (9) or displacer #3 (boxed label 38 in FIG. 8 and FIG. 14, replacing boxed label 13, 29, and 30). That is partially achieved replacing the recuperator/exhaust heat exchanger (5) with an exhaust SSR (28), and partially achieved by replacing the exhaust heat exchanger (5) with the VDSR heat exchanger (24). The final means to reducing “dead space” is to replace the cooler (12, 29) with the VDSR cooler (35).

As can be seen in FIG. 8 and FIG. 14, the only changes required to replace the EISG CCVC testbed cooler with a VDSR cooler are limited to the isentropic displacement process between the expander (9) and displacer #3 (38). From BDC, as the exhaust flows out of the expander (9), past the expander exhaust valve (10), past ECV #1 (11), it enters and flows through the exhaust SSR (28). from there it passes through a new regenerator-to-regenerator check valve (boxed label 34 in FIG. 8 and FIG. 14) and enters the VDSR cooler (35). The VDSR cooler as shown may flow directly into displacer #3 (38) or it may flow through another check valve, ECV #3 (boxed label 37 in FIG. 8 and FIG. 14). Displacer #3 (38) in the exiting EISG CCVC tested exhausts through “compressor check valve (14), now called ECV #4 (14).

In essence, during the move from BDC to TDC, the working fluid exhausted from the expander (9) will enter displacer #3 (38) via either the VDSR cooler (35) or ECV #3 (37), and during the move from TDC to BDC, the working fluid from displacer #3 (38) will bypass the VDSR cooler (35) and go to the cold side of the exhaust SSR (28). This will have the effect of reducing the cooling requirement for the VDSR cooler (35) and increasing the temperature exhausted out of the hot side of the exhaust SSR (28). Since reducing the cooling requirement will potentially reduce the internal volume of the SSR for a given amount of temperature drop, any increase in manifolding due to the bypass may be accounted for.

For an hybrid Ericsson/Stirling closed cycle engine, substitute a gaseous radiator (1) for the condenser (19) and a compressor (13) for the liquid pump (21).

Description—Fourth Embodiment

Modern Rankine cycles operate at supercritical H20 steam levels to attain decent efficiencies. On the poles of Earth's Moon, special areas called Permanently Shadowed Regions (PSRs) exist that never see sunlight. In the depths of some polar craters that never see sunlight, extremely low temperatures approach 100 K.

One usefulness of Rankine/Stirling CCVC engines may be their ability to operate with very high power density and good efficiency at low temperatures. With careful design, a peak temperature of 810 K (1,000 deg F.) may be sustainable with teflon seals and bearings and no requirement for lubricant or coolant.

Some potential Rankine/Stirling CCVC working fluids can be found to be extremely useful in proximity to PSRs. One example is CO2, which can attain supercritical levels at temperatures of about 300 K (80 deg F.) and about 1,000 psi. A heat engine that can exhaust at a final waste heat of perhaps 100 deg F. can thus easily turn solid CO2 (dry ice) at 1,000 psi into a supercritical gas. Also, at an exhaust pressure of even 1 atm, CO2 gas can be converted back to dry ice at as high a temperature as 195 K (−78 deg C.), which should be attainable in most PSRs.

The theoretical Carnot thermal efficiency ((Th—Tc)/Th) of a heat engine with a source temperature of 810 K (1,000 deg F.) and a theoretical sink temperature of 195 K is ((810-283)/810=) 0.65 or 65%. Coupled with the extremely high delivered thermal efficiency of a compressor-less engine and the very low relative peak temperature, it's likely that a delivered thermal efficiency could approach 50%. For solar energy, that means possibly 40% of the focused solar radiant energy can be converted into electricity, possibly matching if not exceeding the best that a solar-powered Stirling engine has ever achieved.

Operation—Fourth Embodiment

It is herein proposed that, in a Rankine/Stirling CCVC engine such as has been described above, working fluids that are liquids and/or solids be (A) compressed, (B) converted to supercritical vapors, (C) passed through the CCVC isochoric processes to (1) permit the efficient recycling of otherwise-waste exhaust heat, and (2) be superheated, (D) be expanded to produce (1) net work out and (2) high quality waste heat for recycling purposes, and (E) any remaining waste heat in the final exhaust from the engine be used to help vaporize said liquids and/or solids in the first place.

It is herein further proposed that, where access to sufficiently low temperatures is available, liquids and/or solids which can be converted to supercritical vapors at those low temperatures be utilized in Rankine/Steriling CCVC engines as a means of expanding the usefulness of low temperature engines such as non-lubricated and/or non-cooled engines.

Description—Fifth Embodiment

Open Cycle CCVC-Based Internal Combustion Engines

Unlike other closed-cycle hot gas engines, a valved cell engine can be open cycle, closed cycle, or even both open and closed cycle. That is because the concept of the valved cell essentially is a means to attach seemingly disparate heat engine processes together. In the original valved cell patents, valved cells were used to connect an externally-pressurized gas to an internally pressurized gas. In essence, converting a closed cycle valved cell (CCVC) to an open cycle valved cell (OCVC) is as simple as taking compressed air into displacer #1, and later exhausting it from displacer #3. Note that even then, all heat can be added from an external heat source in any of the manners proposed above. In fact, some heat can be added that way, and additional heat can be added by injecting fuel into the preheated, pressurized gas, at any of several places. Also, the fuel can be atomized and injected liquid, as is presently done with diesel engines, or it can be vaporized and valved cell-injected gas, as was proposed in the original valved cell patents.

For an open cycle based on the EISG CCVC testbed, or Open Cycle Valved Cell (OCVC) isochoric fully regenerating heat engine, the working fluid can be compressed air, the source heat input can be the internal combustion of a fuel, and the final exhaust can be vented to the atmosphere. Note that a fully regenerating OCVC with isochoric internal combustion can be thought of as a fully regenerating Otto/Stirling hybrid cycle engine, or with isobaric combustion, a fully regenerating Diesel/Stirling hybrid cycle engine, or with isothermal combustion, a fully regenerating Carnot/Stirling hybrid cycle engine.

Semi-Open Cycle CCVC-Based Internal Combustion Engines

Two examples of semi-open CCVC-based internal combustion engines have been suggested.

H2—Enriched Steam Engines

Adding H2 to pressurized steam to create a H2-spiked working fluid that will not in itself combust allows such a working fluid to be taken into displacer #1 in a CCVC engine such as described above that has been converted to an OCVC engine. Following or even during the displacement from displacer #1 to displacer #2, injecting a quantity of O2, as with a valved cell, will literally convert 2 molecules of O2 and a molecule of H2 into 2 molecules of H2O and release heat, probably without even requiring a spark plug or glow plug. The exhaust, barring a few molecules that didn't get converted, is thus more steam than was originally put into the engine, plus work out, plus waste heat. It can be called a semi-closed cycle because the H2O thus created can be broken back down into its H2 and O2 constituents, as by electrolysis, and recycled continually.

O2—Enriched Steam Engines

Alternatively, rather than adding H2 to steam, O2 or even H2O2 could be added to steam. H2 would then be injected following or even during isochoric displacement, once again converting 2 molecules of H2 and 1 of O2 into steam, allowing the eventual breaking of the resulting H20 into separate H2 and O2 gases, as with electrolysis, thus creating a semi-closed cycle.

Operation—Fifth Embodiment

In a conversion of a regenerating CCVC engine into an air-breathing regenerating OCVC engine, the use of internal combustion as a means of increasing the thermal efficiency of said air-breathing regenerating OCVC engine.

In a conversion of a regenerating CCVC engine into an air-breathing regenerating OCVC engine, the use of internal combustion by means of liquid or solid fuel injection as a means of increasing the thermal efficiency of said air-breathing regenerating OCVC engine.

In a conversion of a regenerating CCVC engine into a regenerating OCVC engine, the use of internal combustion by means of liquid or solid oxidizer injection as a means of increasing the thermal efficiency of said regenerating OCVC engine.

In a conversion of a regenerating CCVC engine into a regenerating OCVC engine, the use of internal combustion by means of both liquid and/or solid and/or gaseous fuel and liquid and/or solid and/or gaseous oxidant injection as a means of increasing the thermal efficiency of said air-breathing regenerating OCVC engine.

In a conversion of a regenerating CCVC engine into a regenerating OCVC engine, the use of internal combustion by means of gaseous fuel injection by valved cell means and/or gaseous oxidizer injection by valved cell means for increasing the thermal efficiency of said regenerating OCVC engine. As an example of how a valved cell means can be used to inject a gas into a regenerating OCVC, see FIG. 8 through FIG. 11 for an example. As shown, a valved cell gaseous injector (boxed label 41 in FIG. 8 and FIG. 9) with a closing solenoid and a stainless steel spring (not shown) that is biased to open when pressures equalize at TDC in the OCVC engine, has been opened at TDC in such a manner as to present from the valved cell a measured quantity of either a gaseous fuel such as H2 or a gaseous oxidizer such as O2 to, respectively, either a working fluid mix with a quantity of gaseous fuel or a working fluid mix with a quantity of gaseous oxidizer. Said quantity of gaseous fuel or gaseous oxidizer may then be actively forced into the expander mix, as with a plunger, or passively displaced into the expander mix, as by a constant pressure addition as the expander mix is both expanded and dropped in pressure, in the manner disclosed in U.S. Pat. Nos. 4,817,388 and 5,179,839. Following said injection of gaseous fuel or gaseous oxidizer, as shown in FIG. 10 and FIG. 1.11, said valved cell gaseous injector (boxed label 42 in FIG. 10 and FIG. 11) shall be closed to permit or finalize the recharging of the valved cell, in the manner disclosed in U.S. Pat. Nos. 4,817,388 and 5,179,839.

Description—Sixth Embodiment

Semi-Open Cycle CCVC-Based Combined B/E Half-Cycle Heat Engines

For the purposes of describing these half-cycles, the The chemo/thermodynamic heat engine cycle analyzed herein is the classic reversible cyclohexane: benzene+H2 cycle or C6H12<=>C6H6+3H2 cycle analyzed in U.S. Pat. No. 3,225,538.

As noted earlier, U.S. Pat. No. 3,067,594 proposed an open-cycle Bland/Ewing chemo-thermodynamic process, U.S. Pat. No. 3,225,538 proposed a closed-cycle Bland/Ewing chemo-thermodynamic process, and U.S. Pat. No. 3,871,179 proposed the application of the B/E cycle to the classic Stirling cycle. U.S. patent application Ser. No. 18/095,463 disclosed the concept of creating a unique full B/E chemo/thermodynamic cycle as disclosed in U.S. Pat. No. 3,225,538 by connecting two semi-open chemo/thermodynamic half-cycles; a semi-open endothermic half-cycle, and a semi-open exothermic half-cycle. For the two half-cycles, the endothermic cycle is C6H12=>C6H6+3H2, and the exothermic cycle is C6H12<=C6H6+3H2. Note that in either direction, the exact same amount of thermal energy is stored as is released. However, the endothermic half-cycle occurs at a much higher temperature for a given pressure than the exothermic half-cycle. That is, the endothermic half-cycle requires higher quality heat than the exothermic half-cycle returns.

Both half-cycles have one purpose in common; generate net work out. But for the endothermic half-cycle, a second purpose is to store thermochemical energy, and for the exothermic half-cycle, it's main purpose is to use that stored energy efficiently to generate net work. Note that by far the largest amount of energy used to power the endothermic half-cycle is stored chemically. That in turn means that the largest amount of net work that will be produced by the two half-cycles will be produced by the exothermic half-cycle. It also means that maximizing the thermal efficiency of the exothermic half cycle is what is going to determine just how efficient both half-cycles are combined.

There is one unique result of the C6H12=>C6H6+3H2 endothermic half-cycle that can be seen as beneficial for converting heat into work, and that is this particular endothermic half-cycle's ability to create a kind of chemical expansion process. Note that C6H12=>C6H6+3H2 literally turns a single molecule into 4 molecules; one molecule of benzene, and 3 molecules of hydrogen gas.

In a heat engine, it is useful to find a way to pressurize a gas, and it turns out that, by pressurizing the C6H12 before it is converted into C6H6+3H2, 4 times as many molecules are generated at any given conversion pressure. In other words, one way to look at the C6H12=>C6H6+3H2 endothermic half-cycle is as a kind of chemical H2 pressurization system.

Operation—Sixth Embodiment

A Rankin/Stirling Fully-Regenerating CCVC-Based Endothermic Semi-Open Half-Cycle Heat Engine

The approach being proposed for the endothermic half-cycle essentially uses the endothermic conversion to produce pressurized H2 gas. For example:

• • 1. Pressurize the liquid C6H12 reactant to some desired pressure.
  • 2. Heat the reactant to the vaporization point.
  • 3. Vaporize the reactant.
  • 4. Raise the pressure of the vaporized reactant with a compressor. The purpose of this is to increase the temperature at which the eventual vapor product (C6H6) will condense over the temperature at which the C6H12 liquid at the pressure given it in step 1 above will evaporate. C6H12 and C6H6 temperatures and heats of condensation/vaporization that are very close to one another (at 14.7 psi, C6H12 boils at 343.9 K with a standard heat of vaporization of 380 KJ/kg, while C6H6 boils at 353.2 K with a standard heat of vaporization of 433 KJ/kg). Consequently, condensing slightly higher pressure C6H6 will theoretically supply more than the required heat of vaporization of C6H12, removing the thermal cost of vaporizing the C6H12 from the required heat in to the proposed process, as will be shown.
  • 5. Preheat the reactant by passing it through a first heat exchanger to the temperature at which the reactant, at the given pressure, will convert the reactant to product in an endothermic catalytic converter.
  • 6. Convert the vaporous C6H12 reactant at constant pressure and temperature into the product C6H6+3H2. Note that, at 100% conversion of C6H12 reactant to C6H6+3H2 product, 1,180 kJ of heat are absorbed chemically per 0.4536 kg (1 pound) of the resulting product. Note for comparison purposes that the vaporization requirement for about 0.45 kg of C6H12 equals (380×0.45=) about 170 kJ, or only about 14% of the required thermal input to create pressurized C6H6+3H2.
  • 7. Exhaust the product from the endothermic reactor and cool the product down with the first heat exchanger to just above the temperature at which the C6H6 will be condense into a liquid. The molar heat capacity of H2 is 28.84 Joules per an increase in temperature of 1 K (mol K). The molar heat capacity of vaporous C6H12 is 105 J/(mol K). The molar heat capacity of C6H6 is 135 J/(mol K). Thus, the molar heat capacity of three mols of H2 plus one mol of vaporous C6H6 equals ((28.84×3)+135=) 221.52 J. That is, the reactant has (135/221.52=) 61% of the heat capacity of the product. Assuming a perfect heat exchange, that will require passing only about 61% of the product through the first heat exchanger to preheat the C6H12 vapor to the temperature of the endothermic catalytic reactor.
  • 8. In a second heat exchanger, cool the 39% portion of C6H6+3H2 down with some process to just above the temperature at which the C6H6 will become a liquid.
  • 9. Recombine the ⅔ and ⅓ streams of C6H6 vapor+3H2 gas.
  • 10. Pass the relatively low temperature C6H6+3H2 product mixture through a fully-regenerating CCVC engine.
  • 11. For the source heat, use the heat separated out in the second heat exchanger. If desired, supply additional source heat to the C6H6+3H2 mixture to superheat it.
  • 12. After the C6H6 vapor and the H2 gas product mix is finally exhausted from the fully-regenerated CCVC engine, cool it down to the point where the C6H6 condenses, thus separating the product into liquid C6H6 and H2 gas.
  • 13. If possible, use any remaining latent heat in the CCVC engine exhaust to further reduce the heat requirements of the CCVC engine, for example, to supply the heat to increase the temperature of the pressurized C6H12 liquid reactant to just below the temperature required for vaporizing it at the given pressure.
  • 14. Based on the latent heat requirements of the proposed endothermic heat engine (the chemical storage heat requirements being only relevant to the exothermic half-cycle), such a fully-regenerated CCVC engine can be expected produce delivered thermal efficiencies well in excess of 50% or more.

A Rankin/Stirling Fully-Regenerating CCVC-Based Exothermic Semi-Open Half-Cycle Heat Engine

For a semi-open exothermic half-cycle, the primary purpose is to generate thermal energy to be used in a heat engine operating at maximum efficiency. Per U.S. Pat. No. 3,225,538, FIG. 1, at 1 atmosphere the conversion of C6H6+3H2 will occur at about 540 K (512 deg F.). That will produce, with negligible work in, 1,180 kJ of heat per 0.4536 kg (1 pound). In addition, in U.S. patent application Ser. No. 18/095,463, it is proposed that a small compressor and negligible work be applied to C6H6 vapor for a similar purpose as that proposed in step 4 above for the endothermic half-cycle engine. That is, it is proposed to supply much of the required heat of vaporization for the C6H6 by the condensation of higher-pressure C6H12. In other words, with such an approach, both the work in to produce heat at 540 K and the heat cost required are negligible compared to the fairly high temperature heat produced. Therefore:

• • 1. It is proposed that this exothermic heat from the C6H12<=C6H6+3H2 half-cycle be generated where there is both a source of useful high grade heat and a source of H2 gas, as for example is being produced by the electrolysis of H20.
  • 2. The C6H6 liquid be converted into pressurized vapor just above the temperature of condensation.
  • 3. The C6H6 vapor be compressed to a slightly higher pressure.
  • 4. Gaseous H2 produced at the site be compressed to the same pressure and temperature.
  • 5. The pressurized, heated C6H6 vapor and the similarly pressurized and heated H2 then be mixed to form a chemically-balanced C6H6+3H2 mix.
  • 6. The C6H6+3H2 mix be passed through a first heat exchanger, raising the mix to the temperature close to the working temperature of the catalytic reaction chamber.
  • 7. The C6H6+3H2 mix then be heated to exactly the working temperature of the catalytic reaction chamber (for 1 atmosphere pressure, about 540 K).
  • 8. The C6H6+3H2 mix be passed through the exothermic reactor at constant pressure and temperature and converted back into hot, pressurized C6H12 reactant vapor.
  • 9. The hot, pressurized C6H12 reactant vapor may then be cooled in the first heat exchange to just above the temperature of condensation by helping to preheat the C6H6 heat exchanger. As noted above, assuming a 100% conversion back to C6H12 reactant, the latent heat in the reactant can ideally only supply about 61% of the thermal energy required to heat the C6H6+3H2 mix to the temperature of the reactor. Therefore, the remainder, or about 87 kJ/(mol K) will need to be supplied by another heat source, for example, the heat being exothermically released by the conversion of the product back into the reactant. For C6H6 at 1 atmosphere, that temperature is 353.2 K. Therefore, a temperature difference of (540-353.2=) about 187 K would require (187×87=16.269 Kilojoules. 39% of that would equal (16.269×0.39=) 6.345 KJ/mol. At STP, C6H6 (liquid) has a mass of 78.11 g/mol. 0.4536 kg (1 pound) of C6H6 would equal (453.6/78.11=) 5.807 mols, and the total heat required would thus equal (6.345×5.807=) about 37 kJ. That leaves a delivered heat output of (1−(37/1,180)=) 0.969 or 97% or 1,173 kJ remaining of the total exothermically-produced heat at about 540 K.
  • 10. Note that a thermal carrier fluid such as H2 or He gas could be used to cool the exothermic catalytic reactor, then be superheated with solar energy. That superheated carrier fluid can then be used to power a fully-regenerating CCVC engine.

Since the overall efficiency of a heat engine is used to determine the thermal efficiency of any heat used in that engine, the “recycled” exothermic heat is being utilized at the overall thermal efficiency of its heat engine. It is expected that such a fully-regenerating CCVC-based exothermic semi-open half-cycle heat engine can produce delivered thermal efficiencies in the range of 50% or more.

Accordingly, the total thermal efficiencies of both the endothermic half-cycle fully-regenerated CCVC engine and the exothermic half-cycle fully-regenerated CCVC engine can be expected to generate delivered thermal efficiency in the range of 50% or more.