Synchronized Regenerators and an Improved Bland/Ewing Thermochemical Cycle

Abstract

For efficiently exchanging heat between two streams of fluid at approximately equal pressure while simultaneously reducing the internal volume and general overall mass of the heat exchange means per quantity of heat exchanged over time, a means termed a Synchronized Thermal Regenerator Exchange Pump (STREP) is proposed.

Description

CROSS-REFERENCE TO RELATED APPLICATIONS

This application claims priority to and the benefit of U.S. Provisional Patent Application No. 63/393,960, filed Jul. 31, 2022, and to U.S. Provisional Patent Application No. 63/439,781, filed Jan. 18, 2023, the entire content of each of which is incorporated herein by reference.

REFERENCES

This field is related in part to the invention disclosed in U.S. Provisional Patent Application No. 63/393,960 termed a Synchronized Thermal Regenerator Exchange Pump (STREP), in part to a heat engine cycle invention disclosed in U.S. Provisional Patent Application No. 63/439,781 termed a Synchronizing Displacer (SD) Valved Cell (SD-VC) engine, and in part on U.S. Pat. Nos. 3,225,538, 3,067,594, 3,871,179, 4,817,388, and 5,215,691.

Reference is also made to U.S. patent application Ser. Nos. 17/746,848, 18/095,463, and 18/197,092.

BACKGROUND

The present invention proposes methods and apparatus for improving the efficiency of heat transfer between two fluid streams, particularly as said methods and apparatus relate to improving the technology disclosed in U.S. Pat. Nos. 3,225,538, 3,067,594, 3,871,179, Pending U.S. patent Ser. Nos. 18/095,463, and 18/197,092, wherein techniques are detailed for creating, among other useful methods and apparatus, a unique thermochemical cycle, termed the Bland/Ewing Cycle (B/E Cycle) after the co-inventors behind U.S. Pat. No. 3,225,538, involving “molecular expansion” and “molecular compression”. The advantage of the B/E Cycle is best exemplified in FIG. 3 and FIG. 4 of U.S. Pat. No. 3,225,538, where PN and T/S charts indicate the potential for increased “power density”. This power density is a result of the reduced compression work in (W-in) following exothermic conversion to fewer moles of gas relative to the increased expansion work out (W-out) following endothermic conversion to increased moles of gas.

This invention particularly relates to improvements to methods and apparatus that permit the efficient employment of endothermic chemical reactions and reversible chemical reactions of the endothermic-exothermic type for transfer of heat and/or production of mechanical energy.

The underlying foundational invention takes the form of a unique heat transfer system or STREP in which a counter-flow regenerator can universally replace a counter-flow recuperator to good effect, particularly where a STREP can increase the efficiency when an endothermic chemical reaction and/or an exothermic chemical reaction are utilized. In one aspect, the heat transfer method or system of this foundational invention may be adapted to heat a space or a substance or it may be embodied as a refrigeration system. In another aspect, the heat transfer method may take the form of a method or system for the production of mechanical work. In the application to thermochemical processes, endothermic and exothermic methods or systems may be cyclical, wherein a chemical reactant endothermically reacts to form a product or products and the product or products are then reacted to re-form the initial chemical reactant. Also it is contemplated that a reactant which will undergo an endothermic chemical reaction may be employed to do mechanical W-in to a method which does not involve converting the products back to the initial reactant. Also it is contemplated that the reformation to the initial chemical substance, since it evolves the total amount of thermal energy absorbed endothermically, may itself drive heat engine processes that produce W-out, said exothermically-produced W-out then being summable with the W-out produced endothermically to equal a total or net W-out for a complete Bland/Ewing cycle. Also it is contemplated that said reformation to the initial reactant can be designed to primarily produce thermal energy rather than W-out. Also it is contemplated that the endothermic process may be designed to primarily produce cooling by substantially lowering the temperature of the product of endothermic dissociation prior to expansion. Also is contemplated that the Bland/Ewing Cycle, when viewed as composed of two half-cycles, can be seen as an efficient means of transporting hydrogen in liquid form at ambient pressure and temperature via a process termed a Benzene Battery as described in U.S. patent application Ser. No. 18/197,092.

Stated broadly, this foundational invention utilizes the improved characteristics of what might be broadly termed a “valved regenerator” over the characteristics of a standard counter-flow recuperator vis-a-vis increasing the efficacy of heat transfer between two fluid streams, particularly as concerns the two fluid streams associated with a Bland/Ewing Thermochemical Cycle.

This background section is provided only for purposes of introducing certain background material relating to the present disclosure and, thus, is not an admission of prior art.

SUMMARY

In several embodiments of the STREP method proposed herein, the efficacy of replacing a counter-flow recuperator with a valved regenerator concerns the ability to efficiently change the temperature of two counter-flowing streams with markedly reduced internal volumes. This is well known to benefit what are termed “stirling engines”, which flow a fixed quantity of fluid back and forth between two constantly changing volumes through an intermediate thermal sponge or regenerator. The STREP concept, however, perceives the fluid flow as being composed of two different streams, which only incidentally may be composed of a fixed amount of common fluid within some device. This makes it particularly useful when exchanging thermal energy between an endothermic fluid reactant and an exothermic fluid product.

The STREP also is differentiated from a stirling engine regenerator in being capable of flowing fluid through a regenerator with set parameters, including isobaric (constant pressure), isochoric (constant volume), isothermal (constant temperature), and all the possibilities in between. It has even been found to be capable of flowing one stream with one set of parameters, as for example isobaric, and the second stream with another set of parameters, as for example isochoric.

This Summary section introduces some features of non-limiting and non-exhaustive examples of the present disclosure, and is not intended to limit the scope of the claims.

BRIEF DESCRIPTION OF THE DRAWINGS

The drawings, together with the specification, illustrate non-limiting and non-exhaustive example embodiments of the present disclosure.

FIGS. 1, 2, 3, and 4 each illustrate a cross-sectional view of a synchronized thermal regenerator exchange pump (STREP) during different stages of a synchronized heat exchanger process according to some examples.

FIGS. 5A and 5B schematically illustrate a STREP system according to some examples.

FIGS. 6A and 6B schematically illustrate a STREP system according to some examples.

FIG. 7 is a photograph of a Closed Cycle Valved Cell (CCVC) prototype engine.

FIG. 8 illustrates a cutaway view of a solid model upon which the CCVC prototype engine was based.

FIG. 9 illustrates a solid model of the multi-spiraled helical ribbing which constituted the flow channels for the various CCVC heat exchangers.

FIG. 10 illustrates how the CCVC prototype heat exchangers were generally constructed of multi-spiraled helical ribbing, which constituted the flow channels for the various CCVC heat exchangers.

FIG. 11 illustrates an isometric and cross-sectional view into an example Synchronizing Displacer (SD) CCVC (SD-CCVC) design.

FIG. 12 illustrates a front view of the SD-CCVC design of FIG. 11 at Top Dead Center (TDC).

FIG. 13 illustrates a front cross-sectional view of the SD-CCVC design of FIG. 11 at a 90 degree crank rotation following TDC.

FIG. 14 illustrates a front cross-sectional view of the SD-CCVC design of FIG. 11 at bottom dead center (BDC).

FIG. 15 illustrates a front cross-sectional view of the SD-CCVC design of FIG. 11 at a 90 degree crank rotation following BDC.

FIG. 16 illustrates a front cross-sectional view of a M-SD design at a 90 degree crank rotation following TDC according to an example.

FIG. 17 illustrates a front cross-sectional view of the M-SD design at a 90 degree crank rotation following BDC according to an example.

FIG. 18 illustrates a pressure/volume/temperature/energy/entropy chart based on FIG. 70, “Marks Mechanical Engineers' Handbook”, 1st edition, 9-148, “Internal-combustion engines.”

FIG. 19 illustrates a separated-out tracing of working fluid states utilizing the chart illustrated in FIG. 18.

FIG. 20 illustrates estimated pressure and volume information from FIG. 19 in graphed curves.

FIG. 21 illustrates chart 14 of the July 2005, EISG report.

FIG. 22 illustrates a graph based on FIG. 1 of U.S. Pat. No. 3,225,538.

FIG. 23 illustrates a simple system for storing and delivering C6H12.

FIGS. 24a, 24b, and 24c illustrate a combined liquid C6H12 and liquid C6H6 storage tank system.

FIG. 25 illustrates a simple heat generating system, as shown in FIG. 8 of U.S. patent application Ser. No. 18/095,463.

FIG. 27 illustrates certain traced, separated-out lines using FIG. 18.

FIGS. 29a and 29b each illustrate a cutaway view of a CAD solid model of an approach to converting an existing CCVC prototype engine according to some examples.

FIGS. 30a, 30b, and 30c each illustrate a cutaway view of a CAD solid model of an approach to converting an existing CCVC prototype engine according to some examples.

FIG. 32 illustrates a mixed isobaric/isochoric STREP according to some examples.

FIG. 33 illustrates an example cycle according to some examples.

FIG. 34 illustrates a low pressure cylinder and a high pressure cylinder of a STREP according to some examples.

FIGS. 35, 36, 37, and 38 illustrate working fluid pathways for work producing cycles and a refrigeration cycle using certain traced, separated-out lines using FIG. 18.

FIG. 39 is copied from FIG. 25 and labels a simple exothermic heat generation system schematic with labels from FIG. 27.

FIG. 40 illustrates FIG. 2 from U.S. patent application Ser. No. 18/197,092.

FIG. 41 illustrates FIG. 40 reconfigured as a refrigeration cycle according to some examples.

FIG. 42 illustrates a mixed isobaric/isochoric STREP according to some examples.

DETAILED DESCRIPTION

Adding an SD mechanism to a regenerator is herein proposed as a means for greatly increasing the overall efficiency of any heat exchange process. A regenerator is a “thermal sponge” that absorbs thermal energy from a working fluid when it flows in one direction and releases that thermal energy back to the working fluid in the reverse direction. A basic STREP would essentially be composed of a receiver cylinder and piston means, an SD cylinder and piston means, valving and connecting manifold means, and a regenerator. The STREP heat exchanger process would function as follows and as illustrated in solid modeled and cross-sectioned FIG. 1 through FIG. 4 below (Note: dashed lines and arrows illustrate fluid flow in FIG. 1 through FIG. 4):

• • (1) See FIGS. 1 and 2. Through an intake valve (In1), intermittently pass a stream of working fluid at constant pressure (P1) and at temperature (T1) through a regenerator (A), through a receiver mechanism intake valve (In2) and into a receiver mechanism such as a piston-and-cylinder arrangement (B), thus changing the temperature of the working fluid in the receiver mechanism to a different temperature (T2) via conduction of thermal energy either into or out of the material of the regenerator. Simultaneously, through an intake valve (In3), intermittently pass a second working fluid stream from some external system at constant pressure (P2) and temperature (T3) into a second “synchronizer mechanism”, such as a SD piston-and-cylinder arrangement (C).
  • (2) See FIGS. 3 and 4. At constant pressure, intermittently pass the first working fluid stream at P1 and T2 out of the receiver mechanism piston-and-cylinder arrangement, through a receiver mechanism exhaust valve (Ex1), and into some external system. Simultaneously, at constant pressure, intermittently pass the second working fluid stream at P2 and T3 through an exhaust valve (Ex2), out of the synchronizer mechanism SD piston-and-cylinder, through a regenerator intake valve (In4), through the regenerator, through an exhaust valve (Ex3) on the opposite side of the regenerator near intake valve In1, and into some external system, thus changing the temperature of the working fluid in the receiver mechanism to a temperature that approaches T1 via conduction of thermal energy either into or out of the material of the regenerator.

Since a constant pressure is maintained within the receiver cylinder during intake and exhaust, receiver cylinder W-in cancels W-out, reducing any W-in to that required to overcome any pumping losses. For a similar reason, W-in and W-out for the SD cylinder also cancels out except for pumping losses.

Arrows with solid lines within FIGS. 5a through 6b illustrate fluid flow through connecting manifolding/ducting. Dashed lines illustrate disconnected manifolding/ducting and no fluid flow. For continuous flow, two or more regenerators operating intermittently and in syncopation can be used, one or more filling while the others are depleting. The schematics in FIGS. 5a and 5b show such a system, in this case showing the use of a single fan, two regenerators and a series of valves as an alternative to positive displacement mechanisms. The use of a fan or turbine is particularly useful where the regenerator is relatively large compared to the flow rate. Thus, for example, the exhaust from a fire heating a residence can be blown through a large regenerator from the regenerator “hot” side to the “cold” side, absorbing thermal energy, then be exhausted to the outside atmosphere. When the regenerator is thermally “full”, then clean air, for example from within the residence, can be blown through the regenerator in the opposite direction, exhausting back into the residence and heating the residence with the thermal energy deposited in the regenerator earlier.

The process shown in FIGS. 5a and 5b can be shown to operate in the following manner (see Table 1 below):

Table 1

• • A—Fan
  • B—Inlet air stream two way splitter
  • C—Heat source
  • D—Two way valve #1
  • E—Two way valve #2
  • F—Regenerator #1
  • G—Two way valve #3
  • H—Heat source exhaust
  • I—Two way valve #4
  • J—Two way valve #5
  • K—Regenerator #2
  • L—Two way valve #6
  • M—Heated clean air exhaust

In FIG. 5a:

• • A. A fan receives clean air, as from within a residence.
  • B. The clean air stream produced by the fan enters a two way stream splitter.
  • C. One of the two air streams feeds into a heat source, in this case an enclosed fireplace.
  • D. The hot exhaust from the fireplace passes through two way valve #1, where it is directed to two way valve #2.
  • E. Two way valve #2 directs the hot exhaust to regenerator #1.
  • F. The hot air passes through regenerator #1, charging the regenerator with heat, and the hot air being cooled in the process.
  • G. The cooled air passes through two way valve 3.
  • H. The cooled air is exhausted, in this case outside of the residence.
  • I. Simultaneously, the second stream of clean air proceeding from the fan and through the 2 way splitter (see step B, above) is directed to two way valve #4, which directs the clean air to two way valve #5.
  • J. Two way valve #5 directs the clean air to previously thermally charged regenerator #2.
  • K. The clean air passes through regenerator #2, cooling the regenerator and the cooling air thus being heated.
  • L. The heated clean air passes through two way valve #6.
  • M. Finally, the heated clean air is passed back into the house.

In FIG. 5b:

When regenerator #2 has been “emptied” of its thermal energy, the two way valves #1 through #6 are switched to the alternative setting. Now hot exhaust from the fireplace passes through two way valve #1 (D), through two way valve #6 (L), through regenerator #2 (K), through two way valve #5 (J), and finally exhausts outside the residence (H). Simultaneously, a stream of clean air proceeding from the fan (A), through the two way splitter (B), through two way valve #4 (I), through two way valve #3 (G), through regenerator #1 (F), and finally exhausts inside the residence through two way valve #2 (E).

As an alternative to the valving arrangement shown in FIG. 5, the two regenerators shown in FIG. 5 can be made to intermittently exchange places, as by a rotation around their center, such that the position of regenerator #1 (F) is on the top and the position of regenerator #2 (K) is on the bottom, allowing the stream of air cooling the regenerator to always be directed to the top, and the stream of air heating the regenerator to always be directed to the bottom, thus simplifying the valving. Essentially, FIG. 5b would look exactly like FIG. 5a and the dotted lines representing an intermittent change in stream direction would be eliminated and replaced by an intermittent movement of regenerator K to the position of regenerator F and regenerator F to the position of regenerator K.

It is also possible, where input heat could be turned on or off intermittently, to use a single non-moving regenerator (see the schematics in FIGS. 6a and 6b), to intermittently heat and cool the regenerator. In that case, the exhaust from fan A would be replaced with two way valve I that either (1) feeds fire C and exhausts through two-way valves D and E, through regenerator F, through two way valve G, exhausting outside with cooled combusted air H, or (2) heats the house by passing from fan A through two way valve I, through two-way valve G, back through regenerator F, through two-way valve E and exhausting into the residence with clean heated air M.

A second two way valve (not shown) may be used to select whether air flows to the fan from inside the house or from outside the house. Since air from inside the house would generally be warmer than air outside the house, drawing air for heating the house from the house helps maintain heated air within the house. Air within the house thus essentially recirculates through the regenerator. Note that, if air from the house were used to feed the fire, which is then exhausted outside the house, “makeup” cold air from outside would need to be drawn into the house from somewhere to adjust for the air being removed to feed the fire.

SD Heat Engines

SD heat engines are proposed herein, including variants. The SD concept is predicated on the breakthrough concept of adding a “synchronized thermal regenerator exchange pump” or STREP to the original externally heated Closed Cycle Valved Cell (CCVC) heat engine concept (as conceived and constructed under a California Energy Commission (CEC) Energy Innovation Small Grant (EISG) early in the 21st century. This modification would create an “SD-CCVC” engine. An SD permits the CCVC process to utilize a “regenerator”, significantly improving power density and overall real-world efficiency. A regenerator is a “thermal sponge” that absorbs and releases thermal energy from and to a working fluid. A well-known engine that uses a regenerator is a stirling engine, which is roughly based on the Stirling Cycle. A regenerator can absorb heat available in a stirling engine's working fluid following working fluid expansion and release a substantial amount of that thermal energy back to the working fluid prior to the addition of a charge of “new” thermal energy from an outside heat source, thus reducing the amount of “new” thermal energy required to operate the engine. A regenerator can also remove thermal energy with a charge of “cold” from a cold source. This allows stirling engines to essentially be run in reverse, creating refrigerated working fluid when the fluid is expanded to below ambient temperature. Such a refrigerating process requires a work source to compress the working fluid and overcome pumping losses. Note that an SD-CCVC engine can also operate as either a heat engine or as a refrigerating engine, although valving would have to be extensively modified.

As will be shown, there are several possible variants other than an SD-CCVC heat engine. These include:

An Open Cycle or OCVC or SD-OCVC heat engine that is externally heated and uses compressed air as the working fluid.

An OCVC or SD-OCVC heat engine that is heated by internal combustion (i.c.) and uses compressed air as the working fluid.

A Mixed heat source or M-OCVC, or M-SD-OCVC heat engine utilizing multiple sources of heat, such as solar heat (medium temperature), heat from the external combustion of a fuel (high temperature), and/or i.c.-derived heat (very high temperature).

An OCVC, M-OCVC, SD-OCVC or M-SD-OCVC heat engine with internal heat in (H-in) produced by injection and i.c. of a fuel and an oxidant into a compressed gas or vapor prior to and/or during the expansion process, where the main body of working fluid into which the fuel and oxidant are injected and combusted would be air, where the combusted products plus air are essentially completely removed at the end of each cycle, and a new charge of air is taken in.

A CCVC, M-CCVC, SD-CCVC or M-SD-CCVC heat engine with internal H-in produced by injection and i.c. of a fuel and an oxidant into a compressed gas or vapor prior to and/or during the expansion process, where the combusted products are essentially completely removed at the end of each cycle, such as by liquefaction of H2O, where the main body of working fluid into which the fuel and oxidant are injected and combusted would be non-reactive, such as He, and where the main body of working fluid is continually recirculated.

A “Benzene Battery” (BB) or BB-CCVC, BB-SD-CCVC, or BB-M-SD-CCVC heat engine, where the fuel is H2 delivered by a cyclical hydrocarbon such as C6H6 (benzene) and thus the cyclical hydrocarbon is completely recycled. (See “The BB Closed Loop Process” below.)

A BB-CCVC, BB-SD-CCVC, BB-M-SD-CCVC, where the H2 is from a BB and the oxidant is compressed O2 gas, O2 liquid, or O2 released from a chemical carrier such as H2O2, and both the H2O and the cyclical hydrocarbon such as C6H6 are completely recycled. (See “The BB Closed Loop Process” below.) Such H2+O2-burning engines may be characterized as part of a closed-cycle energy capture and conversion system, the heat engine being supplied H2 by the endothermic dissociation of a cyclical hydrocarbon such as cyclohexane (C6H12) into a “carrier hydrocarbon” such as C6H6, and the heat engine being supplied O2 in either compressed gas form, liquid form, or chemical form such as H2O2, where said H2 and O2 are continually recycled in the form of easily stored and shipped exhausted H2O and C6H6, which are potentially continually reusable. For example, such a process put in place on the lunar surface would ideally only require the original chemical constituents, the mechanisms themselves, a source of high temperature source energy such as concentrated solar energy, a means for removing waste heat such as a radiator or cooler, and various storage and shipping means. Note that all the oxidizer and fuel chemical constituents can be stored indefinitely, thus acting as a kind of “battery” for releasing thermal energy over the two week long lunar night.

The Existing CCVC Design.

As mentioned earlier, FIG. 7 is a photograph of the existing CCVC prototype. FIG. 8 is a cutaway of a solid model upon which the CCVC was based. FIG. 9 is a solid model of the multi-spiraled helical ribbing which constituted the flow channels for the various CCVC heat exchangers. FIG. 10 is a photograph showing how the CCVC prototype heat exchangers were generally constructed, in this instance showing five multi-spiraled ribbed channels. If the multi-spiraled ribbed channel illustrated in FIG. 10 were inserted within a second larger diameter multi-ribbed channel, that would create two counter-flowing multi-channel fluid streams passing in close physical proximity to one another, as is shown in FIG. 9, thus allowing two counter-flowing fluid streams to exchange heat with one another.

The existing CCVC prototype, as shown in FIG. 8, uses a single contiguous hollow piston which creates four distinct chambers within two equal-sized cylinders connected by a smaller diameter third cylinder; the expansion chamber, the compression chamber, the lower displacer chamber, and the upper displacer chamber. These were created by placing two piston discs or heads of 2.5″ diameter on either end of a 2″ diameter connecting tube. Running teflon seals were added to the perimeters of the two heads, and a third stationary teflon seal was in contact with the 2″ diameter connecting tube. A small 0.375″ diameter guided drive rod, also put in contact with a teflon seal at the base of the compression cylinder, connected to a piston guide captured between guide blocks, said piston guide being driven by a standard piston connecting rod connecting to a standard crankshaft with a throw of 2.75″. At Top Dead Center (TDC), the (bottom) compressor space would thus have a volume of 13.39 cu in. At Bottom Dead Center (BDC), the (top) expander space would have a volume of 13.5 cu in. At TDC, the space between the 2.5″ diameter expander cylinder wall and the 2″ diameter connecting tube created a space in the upper displacer cylinder with a volume of 8.64 cu in. At BDC, the space between the 2.5″ diameter compressor cylinder wall and the 2″ diameter connecting tube likewise created a space in the lower displacer cylinder with a volume of 8.64 cu in.

The existing CCVC prototype is designed to use teflon+stainless steel spring seals for its piston rings and tube and guided drive rod seals, permitting essentially non-lubricated, low friction movement of the CCVC piston. The internal engine volumes are pre-pressurized to some desired pressure.

In action, beginning at ˜TDC, a cold working fluid such as helium at some pressure is drawn through a poppet intake check valve into the lower displacer cylinder. At ˜BDC, the working fluid is exhausted from the lower displacer cylinder through a poppet exhaust check valve. The working fluid passes into the recuperator counterflow heat exchanger (shown in outline as dotted lines in FIG. 8), from lower to upper and from cold to hot. The working fluid cyclically passes through an inner multi-spiraled set of helical ribbing, where it is preheated with working fluid cyclically exhausting from the engine expander through a completely separate outer multi-spiraled set of helical ribbing. The fluid exits the hot end of the heat exchanger and passes into the top of the heater heat exchanger.

In the existing CCVC prototype, the heater was electrically heated by an internal cartridge heater and an external coil heater. The heater heat exchanger, like the recuperator, has an inner and outer multi-spiraled set of helical ribbing. Note that the heater was originally constructed with seals such that the inner multi-spiraled set of helical ribbing received working fluid from the recuperator on its way to the upper displacer, and the outer multi-spiraled set of helical ribbing received working fluid from the upper displacer on its way to the expander. The electric cartridge heater was put in physical contact with the inner ribbing and the electric coil heater was put in physical contact with the outer ribbing. However, the design was changed, in an attempt to reduce pumping losses, to run working fluid from the recuperator through both ribbed spirals and into the upper displacer during the upstroke, and through both ribbed spirals and into the expander during the downstroke. Note that doing so had zero impact on the total volume “seen” within the heater by working fluid.

The displacement between the lower displacer and the upper displacer is therefore a constant volume waste heat addition process that completes at ˜TDC and thus raises the pressure of the captured working fluid.

Slightly before TDC, a special poppet-type “transfer valve” (expander intake valve) is opened that connects the working fluid in the upper displacer, recuperator, and heater to the expansion chamber. The valve is actively biased to automatically open, such that, if pressures on both sides of the poppet head are equal, the valve will pop open. To equalize pressure on both sides of the transfer valve poppet head, the poppet-type expander exhaust valve, which is biased towards closed and mechanically driven open by a rocker arm connected to a push rod connected to a cam on the crankshaft, closes slightly before TDC. Dead space at TDC is minimized, which allows any remnant working fluid captured in the expander at the close of the expander exhaust valve (as it approaches TDC) to be re-pressurized to at or above the pressure of the hot, pressurized working fluid in the upper displacer and the heater. Consequently, the transfer valve wants to pop open. A tiny projection at the bottom of the transfer valve poppet head is designed to physically contact the top of the piston just prior to TDC, thus ensuring that the transfer valve will in fact begin to open.

As the expander travels from TDC to BDC, the upper displacer then is able to exhaust the working fluid back through the heater, through the transfer valve, and into the expander. Since the displacer is low volume and the expander is high volume, an expansion process thus occurs. Note that the expanding working fluid includes the volume captured in the recuperator, the upper displacer, the heater, and the various manifolds and plenums connecting these elements.

As the piston approaches BDC, the exhaust valve begins to open into the expander cylinder, where the working fluid pressure has been reduced by volumetric expansion. At the same time, an arm on the exhaust valve pushes the transfer valve towards closed and holds it in place there. Note that, as pressure builds within the upper displacer cylinder during the ensuing displacement “charging” stroke, the transfer valve will eventually hold itself closed by pressure differential. Thus, when the exhaust valve begins to close as it nears TDC, it lifts the arm off of the transfer valve, leaving it prepared to automatically pop open again when pressure access the transfer valve head equalizes in the manner described above.

From BDC, the expander piston now increases pressure, until it reaches sufficient pressure to drive the expanded working fluid past a lightly biased check valve. (Note: It is surmised that one of the major reasons the prototype was unable to produce net W-out was due to a very early closure of the transfer valve, which caused super-expansion and recompression in the expander. The super-expansion is the reason for the addition of the exhaust check valve.) This pressure differential-actuated check valve is constructed integral to the mechanically operated exhaust valve, essentially sliding back and forth along the valve stem. With the opening of the exhaust check valve, the exhausting working fluid is allowed to enter the outer multi-spiraled helical ribbing of the recuperator, thus passing otherwise-waste heat to the inner multi-spiraled helical ribbing, as described above.

The exhausting fluid from the expander then exits the recuperator and enters the multi-spiraled helical ribbing of the cooler, which further drops the working fluid temperature. Finally, the exhausting fluid enters the compressor cylinder near the base of the engine that sits on top of the drive unit. Since the volume of the compressor cylinder at full extension closely approximates the volume of the expander at full extension, the process of moving fluid out of the expander and into the compressor essentially occurs at constant volume, albeit a tiny amount of work is required to overcome the small volumetric difference. (Note: This volumetric difference can be avoided by passing a rod of exactly 0.375″ diameter out of the top of the expander. However, it would require a high temperature seal contacting that rod plus some method of lubrication.)

The process of exhausting from the expander, through the two heat exchangers, and into the compressor can thus be seen as a kind of constant volume displacement process that removes heat from the exhausting working fluid. Since this heat removal occurs at constant volume, the pressure of the working fluid during exhaust is thus continually reduced until TDC is reached.

Following TDC, the direction of the compressor piston is reversed, and the volume composed of the interior of the recuperator, the cooler, and the compressor cylinder begins to climb in pressure from the resulting mechanical compression. (Note: In the original design, the exhaust from the compressor was directed to the displacer intake check valve.) Also at TDC, the volume in the lower displacer cylinder assembly will begin to expand, automatically closing the lower displacer poppet exhaust valve, which is lightly biased towards closed to reduce pumping loss through the valve. That will permit a fresh charge of pressurized and cooled gas to be taken into the lower displacer cylinder, thus returning the engine to its initial state at TDC and completing a full cycle.

Testing of the original CVCC prototype verified that the expected pressure differentials were in fact occurring as predicted. However, no net W-out was ever observed in the existing CCVC prototype. In large degree, that was determined to be the result of four factors:

• • (1) Pre-pressurization of the existing CCVC prototype was too low to generate sufficient power density to overcome the engine's friction and pumping losses.
  • (2) The large internal volumes of the heater+recuperator and the cooler+recuperator greatly reduced the potential pressure differentials.
  • (3) Too-early closure of the expander transfer valve.
  • (4) the likelihood of a too limited temperature spread in comparison to the fixed displacer/expander volume ratio.

Regarding (1), since the prototype required W-in to rotate, a higher pressure was difficult to achieve while the prototype was being rotated with no net W-out during the process. That problem would be solved if net W-out could be increased.

Regarding (2), the SD-CCVC design, in replacing the large internal volume recuperator with a much smaller internal volume regenerator, would solve that problem.

Regarding (3), it is likely that, since the pass through volume from the upper displacer into the expander was at a maximum halfway through the stroke, a “suction” was developed that helped to unseat the transfer valve, due to the maximum speed of the dropping expander piston being achieved at exactly that halfway point. Therefore, converting the transfer valve to a fully physically-actuated valve rather than a partially pressure differential-actuated valve should solve that problem.

Regarding (4), increasing the relative volume of the displacer piston would create a better match due to a decreased displacer-to-expander expansion ratio.

The SD-CCVC design being proposed herein is expected to address all but #3 of these issues.

One Possible SD-CCVC Heat Engine Design.

FIGS. 11 through 15 indicate one possible SD-CCVC heat engine design, in this case utilizing most of the elements of the existing CCVC prototype, albeit rearranged, partially modified, and/or duplicated. The proposed SD-CCVC design, being closed cycle, requires that all source heat be added by some form of external heat exchanger (external heater).

FIG. 11 is an isometric and cross-sectioned view into the proposed SD-CCVC design, and labels various elements of the proposed SD-CCVC design. FIG. 12 is a front view of FIG. 11 at TDC, and additionally labels various elements. FIG. 13 is a front cross-sectioned view at a 90 degree crank rotation following TDC. FIG. 14 is a front cross-sectioned view view at BDC. FIG. 15 is a front cross-sectioned view at a 90 degree crank rotation following BDC (a note in FIG. 15 illustrates the location of a potential H2O removal site, whose usefulness will be made apparent).

Table 2 below describes and defines the parts of the proposed SD-CCVC design as shown in FIG. 12. Table 2 is also referred to under the heading below entitled “Cycle analysis”.

Table 2

• • A. Lower displacer-and-2nd stage compressor cylinder
  • B. Regenerator (STREP)
  • C. Upper displacer cylinder
  • D. Displacer piston connecting tube
  • E. Lower displacer actuated exhaust valve
  • F. Lower displacer-and-2nd stage compressor piston
  • G. Lower displacer intake check valve
  • H. Expander cylinder
  • I. Expander piston
  • J. Expander exhaust valve
  • K. Expander intake transfer valve
  • L. External heater
  • M. Upper displacer piston inlet and regenerator exhaust check valve
  • N. Upper displacer piston
  • O. SD cylinder
  • P. Guided piston rods and guides (3 ea) (guide blocks not shown)
  • Q. SD cylinder actuated inlet valve
  • R. SD cylinder exhaust check valve
  • S. SD piston
  • T. 1st stage compressor-to-SD connecting rod
  • U. 1st stage compressor intake transfer valve
  • V. 1st stage compressor cylinder
  • W. 1st stage compressor piston
  • X. 1st stage compressor cylinder actuated exhaust valve
  • Y. 1st stage compressor cylinder exhaust check valve
  • Z. 1st stage compressor exhaust cooler
  • AA. 2nd stage compressor piston connecting tube
  • AB. 2nd stage compressor intake check valve
  • AC. 2nd stage compressor exhaust check valve
  • AD. 2nd stage compressor exhaust cooler
  • AE. SD inlet check valve

The existing CCVC prototype has (from bottom to top) the compressor, the lower and upper displacer, and the expander in a single vertically-combined assembly operated by a single crank throw. The proposed SD-CCVC design, has three assemblies (see FIG. 11) made up of an expander assembly (on the left), a vertically-combined displacer and 2nd stage compressor assembly (in the middle), and a vertically-combined SD and 1st stage compressor assembly (on the right). It is obvious that other ways to combine the various elements are possible. Finally, note that all three crank throws are in line and perfectly syncopated, thus simultaneously hitting both their respective TDCs and BDCs. Note that, though not shown in this configuration, it is possible for a single crank throw to drive all three assemblies. That may be useful in letting the three separate assemblies be “wrapped around” a central drive rod in a future iteration, thus minimizing manifold spacing.

The SD-CCVC heat engine design shown in FIGS. 11 through 15 can be said to vary in large part from the basic design of the existing CCVC prototype by the addition of an SD cylinder (O) and piston (S). In this iteration, the added SD cylinder and piston match the bore of the expander cylinder (H) and the stroke of the expander piston (I), but operate 180 degrees out of phase to the expander piston. Therefore, since the working fluid exhausted from the expander cylinder is transferred to the SD cylinder at essentially constant pressure, temperature and volume via a manifold (not shown), only pumping losses and thermal losses need to be taken into consideration as affecting the state of the working fluid before and after the transfer.

A second way the proposed SD-CCVC design varies from the existing CCVC prototype concerns the proposed means of recapturing otherwise-waste heat. The existing CCVC prototype relied on classic counterflow external heat exchangers at various points in the cycle. In contrast to the existing CCVC design, a STREP, which is a kind of internal thermal regenerator (B), is used in the proposed SD-CCVC design. The use of a STREP is effectively made possible by adding the SD to the existing CCVC engine design.

A STREP's use of internal thermal regeneration makes it more compact than counterflow thermal exchange/recuperation, and thus more practical for a cycle that relies on constant volume displacement processes, such as the existing CCVC engine. Thermal regeneration is also generally more efficient for heat transfer than thermal recuperation, since it increases the ability to more completely transfer the total heat differential between the counter-flowing streams of working fluid.

In addition to the SD and the STREP, a two stage inter-cooled compressor system has also been added to the proposed SD-CCVC design. Increasing the number of inter-cooled compression stages is well known to assist in approaching an isothermal compression, which will aid overall thermal efficiency.

The 1st stage compressor piston (W) is designed to match the diameter and stroke of the SD piston. Thus, having passed through the STREP, the 1st stage compressor cylinder will receive the working fluid exhausted from the SD cylinder at constant volume. To accomplish this, the 1st stage compressor piston has a small diameter 1st stage compressor-to-SD connecting rod (T) on the opposite side of the piston head from the drive unit. The connecting rod passes through a stationary teflon seal held in the 1st stage compressor cylinder head and attaches to the SD piston head, causing (1) the SD piston to exhaust a charge of working fluid through the STREP and into the 1st stage compressor cylinder when the 1st stage compressor piston takes in working fluid, and (2) the SD piston to take in a fresh charge of working fluid from the expansion cylinder's exhaust manifold when the 1st stage compressor space exhausts its latest charge of working fluid. Thus, this working fluid exchange process from the SD cylinder via the STREP to the 1st stage compressor cylinder occurs at essentially constant volume.

The proposed SD-CCVC design's lower (A) and upper (C) displacer cylinders operate exactly like the existing CVCC lower and upper displacer cylinders. Note that the upper displacer cylinder is exactly the same diameter as the lower displacer-and-2nd stage compressor cylinder, and in this instance is also the same as the expander cylinder, the SD cylinder, and the 1st stage compressor cylinder. That is, a large diameter connecting tube (D) forms the inner surface of both the lower and upper displacer cylinders, and a piston head on either end of that tube (N, F) carry teflon piston seals. As in the existing CVCC prototype, a stationary teflon sealing ring (not labeled) is used on the large diameter connecting rod (D), separating the working fluid in the lower cylinder from the working fluid in the upper cylinder. Note: The stationary displacer cylinder sealing ring is placed appreciably closer to the cooler lower displacer to allow a greater temperature to be tolerated in the upper displacer cylinder.

As noted above, the displacer and 2nd stage compressor assembly is physically separated from the expander cylinder, shown on the left side in FIGS. 11 through 15. That separation allows both the expander cylinder and the upper displacer cylinder to utilize a piston with a “standoff extension”, which in turn keeps the teflon piston seal from running on the (hotter) expander cylinder walls by physically moving it farther away. In addition, separating the two assemblies permits the expander inlet and upper displacer outlet ports (not labelled) to be located much closer to one another, reducing manifold space. In like manner, the SD cylinder, shown on the right in FIGS. 11 through 15, will be receiving hot working fluid at constant volume from the expander, and thus also requires a piston standoff extension that keeps the piston seal from running on the (hotter) SD cylinder walls. By physically removing non-lubricated seals (such as teflon seals) from running directly on the hot expander, upper displacer, and SD cylinder walls, higher peak temperatures are permitted within the engine, which increases potential thermal efficiency. In addition, actively cooling the portion of the cylinder walls that the non-lubricated seals run on may permit even higher peak temperatures. It's also possible to actively cool the area surrounding the (non-stationary) piston seals, permitting even higher peak temperatures.

Note: Because the working fluid passing into the 1st stage compressor (V) will be dramatically cooled and dropped to a lower temperature during the displacement expansion of working fluid out of the SD cylinder through the STREP and into the 1st stage compressor, and the 2nd stage compressor (A, AA, and F) and lower displacer (A, D, and F) will likewise be receiving dramatically cooled working fluid, the 1st and 2nd stage compressors and the lower displacer will not require piston standoffs.

As in the existing CCVC prototype, the lower displacer-and-2nd stage compressor piston head (F) is double-sided. The 2nd stage compressor cylinder assembly is composed of the lower displacer-and-2nd stage compressor piston head, the lower displacer-and-2nd stage compressor cylinder (A), and the compressor connecting tube (AA). The 2nd stage compressor connecting tube connects to the lower displacer-and-2nd stage compressor piston head on the upper end, passes through a static sealing ring (not labeled), and connects to the guided piston rod (P) at the lower end.

Finally, note that “breather holes” (not labelled) are drilled in the plate connecting the lower displacer piston head to the guided connecting rod. As a result, the upper displacer piston assembly and the lower displacer-and-2nd stage compressor piston assembly are seen to essentially be composed of a series of connected tubes with varied diameters. That means the interior of the combined piston assembly can easily pass a fluid through the interior of the piston via the top of the upper displacer piston assembly and the bottom of the 2nd stage compressor piston assembly “breather holes”. This tube arrangement is potentially useful for helping internally cool the displacer piston seals, but also for allowing the elimination of external pressure differential across the piston.

Note that, since the area displaced by the upper displacer piston is significantly larger than the area displaced by the 2nd stage compressor piston, if one assumes a sealed and constant volume upper and lower crankcase, then the pressure in the crankcase will elevate during the upstroke of the 2nd stage compressor piston, and reduce during the downstroke. However, if an inlet check valve were attached to the lower crankcase and an outlet check valve were attached to the upper crankcase, then the crankcase fluid would be exhausted at or near constant pressure from the upper crankcase via the outlet check valve and would be taken into the lower crankcase through the inlet check valve at or near constant pressure in spite of the varying volume. Note that the fluid thus transferred can be used to help cool the interior of the upper displacer piston, again raising the potential peak temperature that a non-lubricated piston ring can tolerate.

Seals (not Shown).

The existing CCVC prototype is designed to use teflon+stainless steel spring seals for its piston rings. Teflon seals will function up to about 555 K (1000 R, 282 deg C., 540 deg F.). However, the SD-CCVC design shown in FIGS. 11 through 17 is designed to potentially allow higher peak temperatures with teflon seals by using standoff extensions and active cooling.

Estimated Volumes.

Since the proposed SD-CCVC heat engine design based loosely on the dimensions of the existing CVCC prototype engine, volumes can be estimated. Note that, in FIGS. 11 thru 17, many standard elements are not shown, such as some connecting manifolds, valve springs, etcetera.

Below are volume estimates for the proposed SD-CCVC heat engine design shown in FIGS. 11 through 17:

• • Engine stroke=2.75″ (70 cm).
  • Expansion cylinder (H)=2.5″ (63.5 cm) dia, 4.91 sq in (31.7 cm2) area*, 13.5 cu in (0.221 L) volume
  • Displacer cylinders (Tot 2) (A, C) volume=2.5″ (63.5 cm) dia, 4.91 sq in (31.7 cm2) area, 13.5 cu in (0.221 L) volume.
  • Displacer piston connecting tube (D)=2″ (5.08 cm) dia, 3.14 sq in (20.27 cm2) area, 8.64 cu in (0.079 L) volume.
  • Total, displacer cylinder volume minus displacer piston connecting tube volume (tot 2)=4.86 cu in (0.0796 L).
  • Total, displacer cylinder area minus displacer piston connecting tube area (tot 2)=1.77 sq in.
  • *Total, expansion cylinder sq in net area=3.14 sq in.
  • SD cylinder diameter and stroke=expansion cylinder diameter and stroke.
  • 1st stage compressor-to-SD piston drive rod (T)=0.225″ (0.572 cm) dia, 0.04 sq in (0.258 cm2) area, 0.11 cu in (0.0018 L) volume.
  • Total, SD cylinder volume minus SD piston drive rod volume=13.39 cu in (0.219 L) volume.
  • Total, SD cylinder area minus SD piston drive rod area=4.87 sq in.
  • 1st stage compressor cylinder (V)=expansion cylinder volume.
  • 1st stage compressor cylinder volume minus SD piston drive rod volume=13.39 cu in (0.219 L) volume.
  • Total, 1st stage compressor cylinder area minus SD piston drive rod area=4.87 sq in.
  • 2nd stage compressor cylinder=expansion cylinder volume.
  • 2nd stage compressor piston connecting tube (AA)=1.58″ (4.016 cm) dia, 1.96 sq in (12.6 cm2) area, 4.86 cu in (0.08 L) volume
  • Total, 2nd stage compressor cylinder volume minus 2nd stage compressor connecting tube=8.65 cu in (0.142 L) volume.
  • Total, 2nd stage compressor cylinder area minus 2nd stage compressor connecting tube area=3.33 sq in.
  • Estimated external heater heat exchanger (L) and manifold internal volume=1.5 cu in (0.0246 L)
  • Estimated exhaust heat regenerator (STREP) (B) and manifold internal volume=1.5 cu in (0.0246 L)

Note: The volumes of the expander exhaust manifold, the 1st stage compressor exhaust manifold, the 1st stage compressor exhaust cooler (Z) and cooler exhaust manifold, the 2nd stage exhaust manifold, the 2nd stage compressor exhaust cooler (AD) and cooler exhaust manifold are not listed since their volumes are deemed inconsequential to the overall cycle, for the following reasons:

The expander cylinder exhausts working fluid into an insulated manifold at constant temperature, pressure, and volume which is taken into the SD cylinder at constant temperature, pressure, and volume. Being a constant pressure process, the physical length of the expander cylinder exhaust manifold is therefore essentially unimportant.

In the cases of 1st and 2nd stage exhaust processes, heat is removed from the working fluid as it is being transferred, potentially down to the temperature of the heat sink. However, the 1st and 2nd stage compressors are designed to exhaust into their respective manifolds and through their respective coolers at approximately constant pressure, since the 2nd stage compressor cylinder assembly and the lower displacer cylinder assembly respectively take in working fluid at essentially constant pressure. Therefore, the physical lengths of the 1st and 2nd stage exhaust manifolds are also essentially unimportant.

Finally, it is anticipated that the walls and likely the pistons of the 1st and 2nd stage compressors will also be actively cooled. As a result, cooling of the 1st and 2nd stage compressor walls and pistons will assist in helping both compressions approach isothermal. A full determination of the impact of compressor and piston cooling will require active testing.

Proposed SD and M-SD Heat Engine Designs.

Proposed SD-CCVC or M-SD-CCVC Designs.

An SD engine would add i.c. source heat only or externally-supplied source heat only. An M-SD would add both i.c. source heat and additional source heat via an external heater. FIGS. 11 through 17 can be used to generally illustrate the working mechanisms for an SD or M-SD engine. In one proposed SD and M-SD design, a means for adding i.c. source heat within the expander cylinder (not shown, see “Injector site” in FIG. 11 and FIG. 16 and “Alternative injector site” in FIG. 17) replaces the source heat from the proposed external heater shown in FIGS. 11 through 17, and would be located in approximately the same area. Adding a means for an i.c. heat source to the SD or M-SD heat engine will allow the creation of true isothermal, isobaric, or isochoric thermal input, as well as isobaric, isothermal, and mixed expansion.

In the case of isothermal heat input (H-in), i.c. can maintain a constant temperature in the expander cylinder throughout expansion, or may be followed with some amount of adiabatic expansion. In the case of an isobaric H-in, i.c. can maintain a constant pressure in the expander cylinder throughout expansion, or may be followed with some amount of adiabatic expansion. In the case of an isochoric H-in, heat would be added within the expander by near-instantaneous i.c., thus adding heat at essentially constant volume. The expansion that follows can then be tailored to anything from isobaric to purely adiabatic by simple timing of a continuing i.c. process, if any.

In the proposed SD and M-SD closed cycle designs, i.c. of H2 and O2 only would be arranged, which would produce only H2O. Thus, the exhaust product requiring removal each cycle could be composed entirely of liquid H2O. Assuming the pre-pressurized primary gaseous working fluid to be pure H2, pure O2, any inert gas (such as He), or a mixture of either H2 or O2 plus an inert gas, the liquid H2O/H2O2 combustion product is easily separated out, with any used H2 or O2 constituent in the working fluid being continually replenished. Note that such an i.c. engine cycle is defined herein as “closed cycle”, since the product of combustion is removed but most of the working fluid is generally recycled. One potential site for such a removal is shown in FIG. 15.

Proposed SD-OCVC or M-SD-OCVC Designs.

In using i.c. heat addition in the apparatus shown in FIGS. 11 through 17, means must be provided for removing the products of combustion. However, with an open cycle design, it isn't possible to do so in the manner undertaken in the closed cycle design, since the products of combustion may be primarily gases and not vapors. One possible removal solution is to exhaust the product of combustion and the working fluid, and take in a new charge of working fluid (for example atmospheric air), as is done with existing Otto Cycle and Diesel Cycle positive displacement engines. For use with the engine designs in FIGS. 11 through 17, exhausting of “used air” could be accomplished, for example, by scavenging out the “used” working fluid (such as air) and replacing it with “new” working fluid, similarly to the scavenging process of existing i.c. two stroke engines. This scavenging would be arranged in the 1st stage compressor cylinder (V), as by arranging the addition of a second intake valve (for the scavenging air and a second exhaust valve (for the exhausting “used air”), or possibly a port, that allowed for scavenging, as is obvious to one skilled in the construction of 2 stroke i.c. engines. Note that the exhaust could be at or near atmospheric pressure, or possibly a higher pressure if a supercharger or turbocharger were used.

Alternatively, the 1st stage compressor shown in FIGS. 11 through 17 could be used as a displacement receiver only, that simply exhausts the “used air” at or near atmospheric pressure. A separate standard 1st stage air compressor would then be used to inject intercooled air into the 2nd stage compressor.

Proposed SD and M-SD Design

In the proposed SD and M-SD designs, the means for adding i.c. heat within the expander cylinder follows and supplements source H-in from the proposed external heater, for example as shown in FIGS. 16 and 17. One important usefulness of the proposed M-SD process is in potentially increasing the efficiency of converting external heat such as solar energy into useful work. Recall that the Carnot theorem, or (T1−T2)/T1, where T1 equals the engine's heat source temperature and T2 equals the engine's heat sink temperature, tells us that any process that increases the source temperature above a given heat sink temperature will increase overall thermal efficiency. Generally, adding i.c.-generated heat following heat addition by an external heater will increase T1, and will therefore increase thermal efficiency. Thus, even though solar thermal energy may be added at a lower temperature than T1, the efficiency at which that added solar thermal energy is converted into work equals the net thermal efficiency of the heat engine.

For the proposed M-SD design variant, i.c. would take place following the external heater H-in, as shown in FIGS. 16 and 17. As with the SD variant, the M-SD variant would be capable of isochoric, isobaric, isothermal, or mixed H-in, and isobaric, isothermal, or mixed expansion.

An Alternative Configuration.

FIGS. 16 and 17 propose an alternative M-SD design to FIGS. 11 through 15. FIG. 16 is comparable to FIG. 13, and FIG. 17 is comparable to FIG. 15. The major physical difference is the movement of the external heater from between the upper displacer cylinder and the expander to between the expander exhaust manifold and the SD cylinder.

By positioning the source heater in front of the displacer cylinder, the M-SD process becomes a means for creating what might be termed a kind of constant volume thermal supercharger. In function, the source heat being supplied by the thermal supercharger is placed “on top” of waste heat coming out of the expander, thus raising the temperature of the working fluid transferred into the SD cylinder. On the following SD cylinder exhaust stroke, that higher temperature heat is then temporarily stored in the internal STREP, where it will in turn be transferred into the upper displacer cylinder. In the process illustrated in FIGS. 16 and 17, the thermal supercharger is shown to be an external heater, and i.c. is shown able to add additional heat between the upper displacer and the expander. Other arrangements for i.c. heat addition are clearly possible.

As stated, the change in the externally heated M-SD design's expander exhaust process illustrated in FIGS. 11 through 15 to the internally heated process illustrated in FIGS. 16 and 17 involves moving from a constant temperature/pressure/volume displacement process to a changing temperature/constant volume displacement process. However, the manifold volume between the two displacers can play a part in any increase in pressure. If the manifold volume between the expansion cylinder exhaust valve and the SD cylinder intake valve is relatively large, the expander-to-SD cylinder process approaches isobaric and the working fluid temperature within the external heater and in the external heater outlet increases as the molecular volume increases. If volume is relatively small, the displacement process approaches isochoric.

In a pure isochoric displacement, during displacement of working fluid from the expander cylinder into the SD cylinder, increasing pressure will raise the temperature of the discrete portions entering the heat exchanger, thus continually reducing the amount of heat per mol each following discrete portion can absorb as the displacement proceeds. Simultaneously, as the displacement proceeds, for each discrete portion exiting the heat exchanger, the constantly raising pressure of each previous discrete portion continually raises the temperature of each following discrete portion over the temperature originally supplied by the heat exchanger. As a result, at the end of the displacement process (at TDC), all the working fluid captured within the SD cylinder will be appreciably above the temperature supplied by the heat exchanger. That is, a kind of constant volume compression process has occurred. (Note: This is also essentially the case in the proposed SD and M-SD designs for the working fluid moved from the lower displacer to the upper displacer via the STREP. That is, the temperature in the upper displacer could theoretically end up higher than the temperature in the STREP, due to the internal compression that occurs during displacement.)

Thus, to the degree that the expander-to-SD cylinder process is isochoric, it will raise the pressure in the expander exhaust manifold, external heater, and SD cylinder intake manifold following displacement. That higher pressure needs to be taken into account in the expander's following exhaust stroke, since the working fluid in the expander at BDC will want to be at a lower pressure than the working fluid on the other side of the exhaust valve.

A similar phenomenon occurred with the existing CCVC prototype. In the existing CCVC prototype, “blow-back” into the expansion cylinder from the expander exhaust manifold made it difficult to pump sufficient working fluid through the engine. As in the CCVC prototype, a check valve would need to be added in close proximity to the expander exhaust valve. In the existing CCVC prototype, a poppet sliding check valve with a light spring bias towards closed was added to prevent blow-back. The sliding check valve that was developed essentially slid over the valve stem of the expansion cylinder exhaust valve to seat in the cylinder head, preventing flow backwards into the expander when the expander pressure dropped below the exhaust manifold pressure. In action, following expansion to BDC, the existing CCVC prototype was allowed to super-expand and then recompress its contents until the pressure difference across the check valve was eliminated, causing the lightly biased poppet sliding check valve to open, thus permitting the working fluid in the expander to exhaust.

Such a combined actuated exhaust valve/exhaust check valve design (Y) is illustrated in FIGS. 11 through 17 for use in avoiding blow-back into the 1st stage compressor, and, while not shown, should be assumed added to the expander exhaust valve as well for FIGS. 16 and 17. Note that other exhaust valve/check valve designs are possible, including wholly actuated designs that don't depend on pressure differential for actuation.

A Possible CCVC Prototype to SD-CCVC Conversion.

Finally, it is possible to “convert” the CCVC prototype to an SD-CCVC prototype by converting the existing compressor to an SD cylinder, adding two external inter-cooled compressors similar to the dual compressor system shown in FIGS. 11 through 17, and adding a “valved regenerator” similar to the valved regenerator system shown in FIGS. 11 through 17. The valved regenerator intermittently and cyclically connects (1) the converted SD cylinder to the 1st stage compressor in a manner that allows the SD cylinder to exhaust at constant volume through the regenerator and into the 1st stage compressor similarly to the system shown in FIGS. 11 through 17, and (2) intermittently and cyclically connects the lower displacer to the regenerator in a manner that allows the existing lower displacer to exhaust at constant volume through the regenerator and into the existing upper displacer similarly to the system shown in FIGS. 11 through 17. Note that the existing recuperator and cooler would in essence be replaced with an insulated working fluid transfer tube that connects and intermittently and cyclically creates constant pressure and constant temperature flow from the expander exhaust into a valving system added to the existing compression cylinder similar to the valving system shown connecting the exhaust manifold to the SD cylinder in FIGS. 11 through 17. This working fluid transfer tube between the expander and the converted SD cylinder would be highly insulated as much as possible to permit the SD cylinder to receive the transferred fluid at constant pressure and temperature.

Cycle Analysis for the SD-CCVC Design.

FIG. 18 is based on a pressure/volume/temperature/energy/entropy chart from FIG. 70, “Marks Mechanical Engineers' Handbook”, 1st edition, 9-148, “Internal-combustion engines”. Basic cycle line drawings such as FIG. 19 will be based on FIG. 18. For clarity purposes, FIG. 19 and other figures throughout have separated out lines that are traceable on FIG. 18. Basic cycle line drawings such as FIG. 19 represent the general thermodynamic states of the heat engine working fluid such as pressures, volumes, and temperatures for various heat engine cycles proposed herein and assist in preparing a first order estimate of the thermodynamic potential of said cycles. Some calculations of estimated states of gases will be conducted through analysis of FIG. 18. However, for additional accuracy, calculations of theoretical states of gases will primarily be conducted through use of The Omni Combined Gas Law Calculator (CGL calculator) found at www.omnicalculator.com. Final efficiency calculations will be based on CGL calculator results. (Note: The existing CCVC prototype was originally designed to allow low-temperature energy sources to be converted into useful work at decent thermal efficiencies. Later, the concept was considered for use on the Moon's surface at the lunar poles. A lightweight flat plate thermal collector on the lunar surface can easily sustain 600 K (1,080 R, 327 C, 620 F) throughout the 2 week long lunar day. Collecting thermal energy in proximity with a Permanently Shadowed Region (PSR) or an artificial PSR, where temperatures in PSR's can approach 100 K (180 R, −173 deg C., −280 deg F.), is seen as a means to create a very low mass solar-powered thermal engine when used in conjunction with a high efficiency low temperature engine such as a CVCC engine.)

The following analysis will assume a working fluid of hydrogen gas (H2), a peak heat engine temperature (T1) of ˜1000 R (555 K, 282 deg C., 540 deg F.) and an exhaust temperature (T2) of 400 R (222 K, −51 deg C., −60 deg F.). The ideal thermal efficiency of a low temperature engine operating between a T1 of 555 K and a T2 temperature of 222K is equal to the well-known Carnot theorem or ((T1−T2)/T1); that is, (555−222)/555, or 60%. Note that real-world engines typically only manage a fraction of this theoretical thermal efficiency. In fact, very few if any engines can operate within a 333 K temperature regime with any meaningful thermal efficiency at all.

As in FIG. 18, FIG. 19 temperatures are in degrees Rankine, pressures are in pounds per square inch, energy and enthalpy are in btu's. Volumes are in cubic feet, but will be corrected for scale, as discussed below. In FIG. 18, the working fluid is a 1 lb (0.454 kg) mixture of air, atomized fuel, and clearance gas. At 650 R (361 K) and 1 atm (101.3 kPa), the mixture occupies about 17 cu ft (481.4 L).

Since an intake volume of 17 cu ft would represent a massive engine, FIG. 18 is clearly representing a much smaller engine undergoing multiple identical cycles. In the proposed SD-CCVC design, a similar multi-cycle representation can thus be assumed, based on the mass of working fluid per cycle.

In some proposed designs the working fluid would be pre-pressurized, possibly to 1,000 psi (68 atm, 6,894 kPa). Pressurizing the working fluid, as is done in full scale stirling engines, is a technique for increasing power density. Note that the impact on thermal efficiency of pre-pressurizing the working fluid is minimal, but the impact on losses from friction, pumping, etcetera, is highly positive. However, to keep the information generally contiguous with the boundaries defined by FIG. 18, the proposed SD-CCVC design is assumed to be pressurized to about 3.5 atm (˜50 psi, 355 kPa) at an inlet temperature of 222 K (400 R). It is assumed that doing so will not impact idealized thermal efficiency.

At a temperature of 650 R (351 K, 88 deg C., 190 deg F.), a pressure of 1 atm (101 kPa), and a volume of 17 cu ft (481 L) per minute, dry air equals ˜29 g/mol. Assuming 28.5 g/mol to account for water vapor, vaporized fuel (C8H18), and remnant gases, total mol count would equal ˜16.9 moles. Using an ideal gas calculator (for example, https://www.meracalculator.com/chemistrv/ideal-gas-law.php), since H2 is nearly n “ideal” gas, H2 volume would equal 17.0 cu ft (481 L) at a similar pressure, temperature, and mol count (16.6 moles for H2), which is close to the estimate found for FIG. 19, and supportive of the use of FIG. 19 as an aid to a first order analysis.

The steps of a typical cycle are described below, with letters in the description representing various elements of the proposed design, as described in Table 2 earlier and as described in FIG. 11 and FIG. 12. The working fluid is assumed to be H2. An analysis of the various pressure/volume/temperature states theoretically achieved in these processes will indicate the theoretical potential efficiency of the design, as traced in FIG. 19. Note that the cycles shown in FIG. 19 are idealized (no pumping, friction, or thermal leaking losses are presumed). However, pumping, friction, and thermal losses are expected to be minimal in a mature design, implying that real world engines of this type should approach fairly closely the thermal efficiency of their idealized cycles.

BDC to TDC—Lower Disolacer to Upper Displacer:

Per the cycle, starting at BDC, the lower displacer cylinder (A) will exhaust, in an isochoric process, cold, pre-pressurized H2 through the regenerator (B) and into the upper displacer cylinder (C). Note that the lower displacer cylinder is composed of the lower displacer cylinder walls and the displacer piston connecting tube (D), while the upper displacer cylinder is composed of the upper displacer cylinder walls and the displacer piston connecting tube.

To keep the information generally contiguous with the boundaries defined by FIG. 19, at BDC the lower displacer cylinder H2 working fluid is assumed pressurized to about 3.5 atm (˜50 psi, 355 kPa) at an inlet temperature of 400 R (222 K) (see Point [1] on FIG. 19). At BDC, the lower displacer exhaust valve (E) would have been mechanically opened. By TDC, the lower displacer piston (F) would have pumped the H2 working fluid past the open lower displacer exhaust valve, through the regenerator, past the regenerator exhaust check valve (M) and into the upper displacer cylinder. Since the total volume would have been held constant, the H2 working fluid would have increased in both pressure and temperature. Assuming 100% regeneration of heat, temperature increased in the displaced mix from 400 R (222 K) in the lower displacer to approximately the temperature exhausting from the expander, or 720 R (400 K) (see Point [2] on FIG. 19).

Per FIG. 19, at the beginning of source H-in at or near TDC, pressure equals approximately 80 psi (551 kPa). Total volume of the lower displacer, the regenerator, and the connecting manifolds are estimated to equal about 6.4 cu in (0.1 L). Per the ideal gas calculator, at a temperature of 400 R (222 K), the mol count would equal about 0.02 mols, which seems about right. At or slightly before TDC, the lower displacer exhaust valve will have closed. The lower displacer exhaust valve is a cam-actuated poppet valve biased by a spring (not shown) towards closed, and is used to maintain a higher pressure within the lower displacer during its intake stroke, as will be shown. Note that a small quantity of heated, pressurized H2 working fluid will remain in the lower displacer following closure of the lower displacer exhaust valve at TDC. With lower displacer piston expansion past TDC, the pressure of the remnant gas in the lower displacer will immediately drop, eventually matching the 345 kPa (50 psi) pressure of the H2 at 400 R (222 K) in the lower displacer inlet manifold (not shown). The cold H2 lower displacer intake check valve (G) will thus be primed to overcome its light spring bias and open, filling the lower displacer with H2 at 400 R (222 K) and 50 psi (345 kPa). Note that, while the lower displacer cylinder volume equals 4.86 cu in (0.08 L), the quantity of new H2 working fluid taken into the lower displacer will be reduced by the amount of any remnant working fluid in the lower displacer expanded back to 222 K and 50 psi. Assuming no remnant, maximum mol count per the ideal gas law calculator would approximate 0.015 moles of H2 per cycle. Per the CGL calculator, following a pure isochoric displacement of 0.08 L (4.86 cu in) of H2 at an initial temperature of 222 K (400 R) and an initial pressure of 345 kPa (50 psi), for a final temperature of 400 K (720 R), internal energy change and heat would equal 0.0545 kJ (0.0516 BTU), and final pressure would equal 622 kPa (90.2 psi).

Near TDC—Remnant Gas Recompression and Pressure Equalization:

Assuming source heat is added by isochoric combustion of H2 and oxygen, just previous to combustion, the previous charge of i.c.-heated H2+H2O (and/or H2O2 if H2O2 is used as the oxidizer) gaseous/vaporous mix will have been exhausted from the expander cylinder (H). During the exhaust, the expander piston (I) moves from BDC towards TDC and the gaseous/vaporous mix will be exhausted past the expander exhaust valve (J) at approximately constant pressure and temperature, as will be shown. The expander exhaust valve is a cam-activated poppet valve with a spring bias towards closed (not shown), with any high pressure differential across the expander exhaust valve head thus sealing the valve closed.

Just prior to TDC, the expander exhaust valve will be closed by the valve's spring bias. This will result in the expander piston rapidly driving any remnant working fluid captured in the expander to a higher pressure and temperature, eventually matching the pressure (80 psi (551 kPa)) on the other side of the expander intake transfer valve (K), thus freeing the expander intake transfer valve to pop towards open via a return spring bias (not shown), and thus connecting the expander, the internal volume of any external heater (L.), and the various connecting manifolds between the transfer valve and the regenerator exhaust check valve. For insurance, a slight mechanical “bash” by the expander piston top is arranged just before TDC to ensure the expander intake transfer valve does in fact open in a timely fashion just prior to i.c. H-in.

At or near TDC, isochoric i.c. will occur, for example by the instantaneous connection, injection, and combustion of a sufficient quantity of pure O2 gas, for example at approximately 150 psi (1,034 kPa). The combustion will instantly increase chamber pressure to approximately 117.5 psi (810 kPa). Note that the back-flowing pressure wave will instantaneously seal the regenerator exhaust check valve (M), thus isolating the regenerator (B) from the pressure surge.

TDC to BDC—Expansions (Expander):

Thermal inputs other than isochoric and expansions other than adiabatic are clearly possible and will be discussed below. That includes what can be termed “displacement heating and expansion” processes. In the instance of a pure isochoric i.c. source heat addition and the instantaneous combustion of H2 and O2 immediately following the charging of the upper displacer cylinder, and assuming no H-in via an external heater and direct injection of the upper displacer into the expander, the volume of working fluid prior to expansion will equal the upper displacer volume (4.86 cu in, 0.0796 L) and the volume of working fluid following expansion will equal the expander volume (13.5 cu in, 0.221 L). 1 g of H2 has a low heat value of 120 kJ or 120,000 J. Increasing the internal energy change of the H2+O2 mix by 51.1 J thus requires the combustion of 0.000426 g or 0.426 mg of H2. One mol of H2 has a mass of 2.0158 g, and the H2 at TDC prior to combustion would equal approximately 0.02 moles, or a mass of 40.32 mg. The mass of H2 combusted will thus equal about 1% of the total mass of H2 in the charge. The mass of O2 combusted per charge would be 8 times the mass of the H2 combusted, or about 3.3 mg of O2.

In one expansion iteration (FIG. 6 through FIG. 9). Assuming zero thermal input following TDC, the following upper displacer piston (N) and simultaneous expander piston movements past TDC and towards BDC will create an ideally adiabatic “displacement expansion”, with expanding gas being continuously moved from the lower displacer into the expander during the expander piston downstroke. Total expandable volume at TDC will thus equal the combined volume of the upper displacer cylinder (0.079 L (4.86 cu in)), and the internal volume of the external heater and manifolding (0.016 L (1 cu in)), or 0.095 L (5.85 cu in) (see “Known and estimated volumes” above). Following combustion and still at TDC, temperature will vary throughout the mix, but ideal peak temperature will be held to approximately 1000 R (555 K) (see Point [3] on FIG. 19).

Note: Permissible peak temperature for an adiabatic expansion process may permissibly be considerably higher than for an isobaric process, which may permissibly be considerably higher than for an isothermal process, since the average temperature over the expansion will be lower. An average working fluid temperature that is lower may permit use of non-lubricated bearing surfaces at a higher peak temperature. The extents of this will need to be determined experimentally.

A. Isochoric Source Heat Input/Adiabatic Expansion

In the instance of a pure isochoric i.c. source heat addition via the instantaneous combustion of H2 and O2 following the charging of the upper displacer cylinder, and assuming no H-in via an external heater, a following adiabatic/isentropic expansion can be made to take place (see Point [4] on FIG. 19). The volume of working fluid prior to expansion will equal the external heater (˜1 cu in, 0.0164 L), the external heater connecting manifolds (˜0.5 cu in, 0.008), and the upper displacer volume (4.86 cu in, 0.0796 L), or a total of 4.88 cu in (0.08 L). The volume of the working fluid following expansion will be equal to the external heater (˜1 cu in, 0.0164 L), the various connecting manifolds (˜0.5 cu in, 0.008), and the expander volume (13.5 cu in, 0.221 L), or a total of 13.65 cu in (0.224 L), or an expansion ratio of ˜2.8/1.

An adiabatic/isentropic expansion of 1 to 2.8 equals an expansion from ˜3.4 cu ft to ˜9.5 cu ft (269 L). Per FIG. 19, for a temperature drop from 1000 R to ˜720 R (400 K), pressure drops to ˜30 psi (207 kPa). Per the gas law calculator, H2 mol count cycle per minute would equal 16.65 moles at 1 atm (101 kPa) and 650 R (351 K) and a volume of 17.0 cu ft (488 L), which closely matches the 17 cu ft, and estimated mol count of 16.9 moles of air plus clearance gases, in FIG. 18. Recall also that a maximum of 0.015 moles of H2 at a temperature of 400 R (222 K) and a pressure of 50 psi (345 kPa) is injected into the cycle via the lower displacer. Finally, recall also that it was determined that FIG. 18 may represent an engine undergoing multiple identical cycles.

Therefore, assuming that 16.65 moles are cycled per minute, a process that cycles ˜0.015 moles can be said to be representing an engine with a cycling speed of ˜1,130 cycles/revolutions per minute (rpm).

In FIG. 20, an estimate of the pressure and volume information shown in FIG. 18 was estimated and inserted into a spreadsheet. The result creates the graphed curves shown in FIG. 20. Total work out for the equivalent of 17 cu ft of H2 cycled over 1 minute is then calculated in foot-pounds/minute for the adiabatic expansion following isochoric (CV) process shown in the top chart, for the isobaric (CP) source heat input process shown in the middle chart, and for the combined isothermal (CT) expansion/source heat input process shown in the bottom chart. Pressures are entered in 2.5 psi differences between the estimated 117.5 psi peak pressure and the estimated 30 psi minimum pressure following expansion.

For the isochoric/adiabatic expansion process, total work out is calculated in FIG. 20 to equal ˜46,270 ft lb/min, or 59.46 BTU/min or 3,568 BTU/hr, or about 1.40 ho (1.046 kW, 3,764 kJ).

Note: 59.46 BTU/hr can be compared to FIG. 18 and the Btu output shown there for the isochoric/adiabatic expansion. Assuming 17 cu ft flow per minute, or 0.015 moles of H2 per revolution at 1,130 rpm, since FIG. 18 indicates Btu drops from ˜100 to ˜40 or a difference of ˜60 Btu/min or ˜3,600 BTU/hr. Per the CGL calculator, for an adiabatic expansion of H2 from (4.86 cu in, 0.0796 L) at an initial temperature of 555 K (1,000 R) and an initial pressure of 863 kPa (125 psi) to a final volume of 0.221 L (13.5 cu in ( ), temperature would equal 370 K (666 R) (a little lower than estimated) and pressure would equal 207 kPa (30 psi). In addition, internal energy change and W-out would both equal 0.058 kJ. At 1,130 rpm over an hour, total H-in and W-out would equal 3,932 kJ (1.092 kWh, 1.465 HP/hr).

B. Isobaric Source Heat Input/Adiabatic Expansion

In the isobaric/adiabatic SD CCVC engine, as in the isochoric/adiabatic SD CCVC engine, the start point for the isochoric waste heat regeneration process is BDC, where the H2 working fluid in the lower displacer cylinder and the upper displacer volume (4.86 cu in, 0.0796 L) is assumed pressurized to about 3.5 atm (˜50 psi, 355 kPa) at a temperature of 400 R (222 K) (see Point [1] on FIG. 19). Assuming perfect isochoric regeneration of exhaust waste heat from the previous isobaric source heat addition and adiabatic expansion, per FIG. 19, isochoric waste heat regeneration will raise the approximately 0.015 moles of H2 to about 790 R (439 K). (Note that this is a higher temperature at which source heat is added than for the isochoric/adiabatic SD CCVC engine mapped above. That is because: (1) All the proposed cycle models are (a) set to a given peak temperature and (b) use the same compression process; and because, (2) as is shown in FIG. 19, an isobaric thermal input inherently drives the adiabatic exhaust process to the right in comparison with an isochoric thermal input, forcing the adiabatic line to connect with the volume limiting line of the isochoric regeneration process at a higher peak temperature.) Per the CGL calculator, that higher input temperature from the isobaric/adiabatic SD CCVC engine exhaust will increase the pressure following isochoric thermal regeneration to about 702 kPa (102 psi). In FIG. 19, an isobaric source heat input/adiabatic expansion process is illustrated, beginning at the point where i.c. isobaric source H-in is begun, that is, at TDC (see Point [5] on FIG. 19). In the instance of a pure isobaric i.c. source heat addition and the instantaneous commencement of combustion of H2 and O2 immediately following the charging of the upper displacer cylinder, and assuming direct injection of the upper displacer into the expander, the volume of working fluid prior to expansion will equal the upper displacer volume (4.86 cu in, 0.0796 L) and the volume of working fluid following expansion will equal the expander volume (13.5 cu in, 0.221 L). The isobaric expansion thus begins at the pressure, temperature, and volume of the H2 working fluid in the upper displacer at TDC, or ˜702 kPa (102 psi), 790 R (439 K), and (4.86 cu in, 0.0796 L).

Per the CGL calculator, the isobaric expansion completes at ˜702 kPa (102 psi), 1000 R (555 K), and ˜0.101 L (6.16 cu in), with H-in equal to 0.0511 kJ (0.0480 BTU) and W-out equal to 0.0148 kJ (0.0140 BTU). (see Point [6] on FIG. 19) At 1,130 rpm over an hour, that equals 1,003 kJ (0.279 kWh, 0.374 HP). Total H-in at 1,130 rpm over an hour equals 3465 kJ (3,284 BTU).

Immediately following isobaric expansion, adiabatic expansion continues until volume equals 0.221 L (13.5 cu in) (see Point [7] on FIG. 19). Per the CGL calculator, the adiabatic expansion of H2 from 0.101 L (6.16 cu in), 555 K (1,000 R) and 702 kPa (102 psi) to a volume of 0.221 L (13.5 cu in) results in a final pressure of 234 kPa (34.9 psi) and a final temperature of 404 K (727 R). The internal energy change and W-out equal 0.0475 kJ (0.045 BTU). At 1,130 rpm over an hour, that equals 3,221 kJ (0.895 kWh, 1.20 HP). Total work out would thus equal 4224 kJ (1.17 kW, 1.57 HP).

C. Isothermal Source Heat Input/Expansion

Perhaps the most intriguing aspect of the proposed SP-OCVC and MCSP-OCVC designs is the ability to make possible a highly efficient isothermal expansion process. In FIG. 19, 1000 R (555 K) was “set” as the peak temperature that a non-lubricated engine can sustain. Future experimentation will of course determine the true sustainable peak temperature. But even at 1000 R, the ability to regenerate heat from the exhaust into the working fluid following expansion will have a profound impact on developing highly efficient heat engines.

In FIG. 19, an isothermal expansion from 0.0796 L (4.86 cu in) is illustrated (see Point [3] on FIG. 19). Note that this temperature is achieved by the theoretical 100% thermal regeneration of the 555 K (1,000 R) exhaust. Isothermal expansion, which is assumed to result from combustion of H2 and injected O2 internally, will then continue to (13.5 cu in, 0.221 L). Per the CGL calculator, an isochoric heat input into 0.0796 L (4.86 cu in) of H2 at an initial temperature of 222 K (400 R) and pressure of 345 kPa (50 psi) to an final temperature of 555 K will result in a final pressure of 863 kPa (125 psi). Per the CGL calculator, an isothermal expansion from 0.0796 L (4.86 cu in) at a temperature of 555 K to a final volume of 0.221 L (13.5 cu in) would result in a final pressure of 311 kPa (45.1 psi) and H-in and total W-out would equal 0.070 kJ. Assuming 1,130 rpm, W-out and H-in over 1 hour for the isothermal expansion process equals 4,746 kJ (1.32 kWh, 1.77 HP/hour).

D. Displacement Expansion

A typical stirling engine uses displacement heating, displacement cooling, displacement compression, and displacement expansion. The proposed SD-CCVC engine can be seen to use the same kind of processes, albeit in separate and discrete segments. Testing of the CCVC prototype gave some data indicating the probable effect of displacement heating in that design.

From General Report on the Close Cycle Valved Cell (CCVC) heat engine test program to the California State Energy Innovation Small Grant (EISG) program administrator, July, 2005 (reference to colors will be clarified below):

“For all tests, a driver motor drove the prototype. As in previous tests, by rotating the engine with a driver motor, pressure transducers were able to “map” some of the internal pressure fluctuations of the engine . . . . These maps show the basic pressure fluctuations occurring within the engine at various temperature inputs and rpm's. Note that the pressure lines are formed by pressure readings captured from the transducers 5000 times per second. Rpm is thus determined by measuring the number of readings per complete cycle (for example, a complete cycle that occurs in 2500 readings indicates a cycle operating at 120 rpm). The red line represents the pressure fluctuations occurring in the engine's heater manifold (the upper displacer/expander “side” of the engine) and the blue line represents the pressure fluctuations occurring in the engine's cooler manifold (the compressor/cooler displacer “side” of the engine). The lines running from top to bottom represent approximate TDC and BDC in the engine.

It is important to keep in mind that the physical makeup of upper displacer and compressor “sides” of the engine changes during an engine cycle, as various valves connect or disconnect various volumes. As a result, the expander will be connected to the upper displacer chamber during expansion and to the compressor during exhaust, while the lower displacer chamber will be connected to the compressor during compression and to the upper displacer chamber during exhaust. This means that the pressure transducers only observed part of what was happening within the engine. The rest must be deduced, as will be shown.”

FIG. 21 is from Chart 14 of the July, 2005 EISG report. (NOTE: This is a greyscale conversion of this data which de-colorizes it. Only the “red” and “dark blue” lines are actual transducer data. The jagged line starting at the top left and ending second from the top right is the “red” transducer data line and the jagged line starting at the bottom left and ending at the bottom right is the “blue” transducer data line.) FIG. 21 indicates that the point was reached within the CCVC prototype where W-in equaled W-out, assuming zero friction, pumping, and thermal losses. It also suggests that the expansion process was not adiabatic, to wit:

On the left top quadrant, the “red” descending line indicates the observed expansion process, up until the transfer valve closed at about 135 psi. (The red re-ascending line moving to BDC represents the re-compression by the hot displacer into the dead space between the expansion cylinder transfer (intake) valve and the cold displacer exhaust check valve, which includes the heater internal volume, the recuperator internal volume, and the various connecting manifold volumes. Note that, as the cold displacer displaces into and through the recuperator at constant volume, the pressure increases as TDC is approached, as predicted.)

On the right top quadrant, the “dark blue” line descending from BDC and about 170 psi indicates an nearly adiabatic compressor gas re-expansion process, up until the exhaust check valve opens at about 115 psi. (The “dark blue” descending line from that point indicates the constant volume displacement of hot gas out of the expander, into and through the expander inlet side of the recuperator, through the cooler, and into the compressor. Note that, as the hot gas displaces into and through the recuperator and cooler at constant volume, the pressure decreases as TDC is approached, as predicted.)

It is visually obvious that the displacement expansion process is not adiabatic, but that heat is being added to the fluid displacing from the displacer, through the heater, and into the expander. Therefore, while it is not possible with just this small amount of data to predict exactly how a displacement expansion curve would track, it is clear that the peak temperature and pressure are quite a bit higher than the final expansion temperature and pressure. That suggests that, as with the isochoric/adiabatic expansion process (and the isobaric/adiabatic expansion process), the peak temperature can be elevated substantially above the overall temperature limit required to operate without lubricant.

It can therefore be assumed that peak temperature for an SD-CCVC displacement expansion engine as modeled above could approximate that of an isochoric engine as modeled above and stay within this limit, although more work would be generated and more input heat would be required with a displacement expansion process than with an isochoric/adiabatic expansion process. Actual determination of thermal efficiency and power output for a pure SD-CCVC displacement expansion engine would thus approximate that of an SD-CCVC isochoric/adiabatic expansion engine, but can't be accurately determined until actual testing is carried out.

One further interesting takeaway from FIG. 21: Consideration was not given at the time to the potential of reducing the sink temperature, as by running the CCVC engine within a PSR on the lunar surface. If the sink temperature were reduced by 200-300 K, that would be effectively the same as increasing the peak temperature by the same amount. Pressure differential would be accordingly raised, and expansion would proceed more deeply than seen in FIG. 21.

BDC to TDC—Expander to SD Cylinder:

At BDC (on FIG. 19, see Point [4] for the isochoric/adiabatic expansion, see Point [7] for the isobaric/adiabatic expansion, and see Point [8] for the isothermal/adiabatic expansion), closing the upper displacer exhaust/expander intake/transfer valve would decrease the volume available for the following exhaust processes to 13.5 cu in (0.221 L). Per FIG. 20, the pressure following expansion equals ˜27.5 psi for the isochoric/adiabatic process (see Point [3] on FIG. 19), ˜32.5 psi for the isobaric/adiabatic process (see Point [6] on FIG. 19), and ˜40 psi for the isothermal process (see Point [8] on FIG. 19). Assuming the exhaust valve completely closes with ⅛″ of travel, since area equals 4.91 sq in across the piston head, volume equals 0.61 cu in (0.010 L). Per FIG. 19, temperature equals 720 R (400 K) and pressure equals ˜30 psi (207 kPa). Per the ideal gas calculator, at a temperature of 720 R (400 K), a volume of 13.5 cu in (0.221 L), and a pressure of 207 kPa, molal mix equals 0.0006 moles. If 0.015 moles transfer, then at maximum expansion, the expander holds 0.0156 moles of working fluid mix. Per the ideal gas calculator, pressure equals 241 kPa (35 psi). The exhaust process discussed below will be similar for all three expansion processes, only differing in the pressure and temperature of the exhausting working fluid mix, but will be assumed for the purposes of calculation to be isobaric at 30 psi (207 kPa). (Note: The pressure results following expansion per the CGL calculator are also different. However, for purposes of generating a 1st order calculation and defining the processes, 30 psi (207 kPa) is deemed acceptable.

As stated above, an SD cylinder enables use of a stirling cycle-type regenerator. At TDC, the SD cylinder (O), via an insulated expander exhaust manifold (not shown), will have received a full charge from the expander cylinder (H) at approximately constant temperature, pressure, and volume, with a small volumetric reduction relative to that of the expander cylinder resulting from the 1st stage compressor-to-SD cylinder connecting rod (T).

Thus, for the isochoric/adiabatic expansion:

Expander cylinder volume equals 13.5 cu in (0.221 L) and expander piston area equals 4.91 sq in. Force of exhaust thus equals 147 pounds over 2.75 inches or 0.227 feet or 33.4 ft lb. At 1,130 rpm, that equals 37,784 ft lb/minute or 48.5 BTU/minute or 2,913 BTU/hr, or 1.14 HP/hr (0.854 kWh, 3.073 kJ/hr).

The SD cylinder internal volume is slightly less than the volume of the expander, equaling 13.39 cu in (0.219 L). SD piston (S) area thus equals 4.87 sq in. The intake W-out thus equals 146 pounds over 0.227 feet or 33.2 ft lb. At 1,130 rpm, that equals 34,353 ft lb/minute, or 1.04 HP/hr (0.776 kWh, 2,792 kJ). The expander exhaust-to-SD cylinder displacement process therefore requires approximately 0.01 HP/hr of W-in. That is a negligible amount, and for that reason calculations for the isobaric and isothermal expansion processes will also be assumed to be negligible.

Exiting the expander, the exhaust passes through the expander exhaust valve, into the expander-to-SD manifold (not shown), past the SD inlet check valve (AE), past the SD inlet actuated valve (Q), and into the SD cylinder. The SD inlet check valve in this instance is a simple poppet valve lightly biased towards closed. That is, a small pressure differential across the valve head will open the valve and allow flow from the expander exhaust manifold to pass, but any higher pressure on the SD cylinder side of the valve head will cause the valve to firmly shut.

The SD cylinder inlet actuated valve is a poppet valve (mechanically closed by a crankshaft-mounted camshaft/pushrod/rocker arm-actuated assembly) that is biased towards open, but blocking gas flow in the opposite direction to the SD cylinder inlet check valve. That is, when the SD cylinder inlet actuated valve is closed, working fluid mix from the expander exhaust manifold cannot go into the SD cylinder.

Note: When pressure in the SD cylinder increases over the pressure across the SD cylinder inlet actuated valve head, (which occurs once per cycle), some leakage can occur. That is the reason for the SD cylinder inlet check valve's presence; that is, the SD cylinder inlet check valve will not allow back-pressure flow into the expander exhaust manifold due to leakage past the SD cylinder inlet actuated valve head.

The SD cylinder inlet actuated valve will manually connect and disconnect the SD cylinder and the expander-to-SD manifold when the pressure in the SD cylinder approximately equals the pressure and temperature of the gas exhausting from the expander, which occurs at two distinct points per cycle, as will be shown below. As a result, the SD cylinder inlet check valve will maintain the pressure in the expander-to-SD manifold at times when the SD cylinder pressure is higher than the expander exhaust manifold pressure, and the SD cylinder inlet actuated valve will maintain the pressure in the expander-to-SD-manifold at times when the SD cylinder pressure is lower than the exhaust manifold pressure. The SD cylinder inlet actuated valve opens (that is, is actuated against the bias) slightly before BDC, and the SD cylinder inlet actuated valve closure occurs at or slightly before TDC, as will be shown below.

TDC to BDC—SD Cylinder to 1st Stage Compressor Via Regenerator.

Starting at TDC (on FIG. 19, see Point [4] for the isochoric/adiabatic expansion, see Point [7] for the isobaric/adiabatic expansion, and see Point [8] for the isothermal/adiabatic expansion), the SD piston, with approximately a full charge of working fluid mix, will travel downward. At TDC, the SD cylinder exhaust poppet-type check valve (R) is being held closed by pressure differential across the valve head. Recall that, by the time TDC has been reached, the working fluid in the lower displacer, regenerator, and upper displacer has reached the maximum pressure just prior to i.c. initiation. SD piston travel from TDC thus begins compressing the captured working fluid mix within the SD cylinder, raising its pressure. Pressure differential across the SD cylinder exhaust check valve (R) will keep that valve from opening, while the SD cylinder intake check valve (AE) will keep any pressurizing working fluid mix within the SD cylinder from entering the expander exhaust manifold.

Simultaneously at TDC, the 1st stage compressor piston (W) will begin expanding working fluid out of the regenerator via the 1st stage compressor cylinder (V) intake transfer valve (U), since the 1st stage compressor intake transfer valve will have previously opened, as will be shown below. This 1st stage compressor piston expansion will thus simultaneously drop pressure in the 1st stage compressor cylinder and the connected regenerator while the pressure is being raised in the SD cylinder. Consequently, pressure will quickly equalize across the SD cylinder exhaust check valve, allowing a flow of hot working fluid mix from the SD cylinder to pass through the SD cylinder exhaust check valve, past the regenerator, past the previously opened 1st stage compressor intake transfer valve (as will be shown below), and into the 1st stage compressor cylinder.

Note: When the SD cylinder exhaust check valve opens as a result of the pressure equalization, the SD cylinder's adiabatic pressurization process and the 1st stage compressor cylinder adiabatic depressurization process will be instantly convert to an isochoric displacement process. Note that the volume of the SD cylinder and the 1st stage compressor both exactly equal 13.39 cu in (0.219 L). As isochoric displacement is initiated, heat will be removed at constant volume from the working fluid mix exiting the SD cylinder by the regenerator, causing the working fluid mix pressure and temperature to instantly begin to drop. Note that this isochoric displacement will begin when the pressure within the SD cylinder drops approximately to that of the working fluid mix in the regenerator+regenerator manifold+1st stage compressor.

In the regenerator at TDC, pressure equals 80 psi (551 kPa) and volume equals 1.5 cu in (0.0246 L). Per FIG. 18, the ideal temperature of the working fluid entrapped in the regenerator would equal ˜720 R (400 K) (that is, the temperature of the gas exhausting from the expander). However, the average temperature of the entrapped working fluid across the regenerator at 80 psi would equal the average difference between the peak temperature and the lower displacer inlet temperature (400 R, 222 K), which would equal 560 R (311 K). Recall that the estimated regenerator and manifold internal volume would equal 1.5 cu in (0.0246 L). From the ideal gas calculator, mol count in the regenerator and manifolding at 311 K (560 R) and 551 kPa (80 psi) will thus be estimated at 0.0052 moles. Assuming an expansion of densified working fluid at the low end temperature of ˜400 R (222 K), very little expansion of the 1st stage compressor cylinder and compression of the SD cylinder will be required to equalize pressure across the SD cylinder exhaust check valve to something less than 80 psi and greater than 30 psi. Moving past TDC, the pressure within the SD cylinder will increase adiabatically over the beginning pressure of 30 psi. At the same time, the pressure within the 1st stage compressor will drop adiabatically under 80 psi. Assuming the combined 1st stage compressor and regenerator space volume doubles from 1.5 cu in (0.0246) to 3 cu in (0.049 L), all in the low temperature side of the regenerator, and assuming an average temperature drop in the combined 1st stage compressor and regenerator space volume to ˜500 R (278 K) at the end of the pressure equalization process, and further assuming 0.0052 moles of gas, then per the ideal gas calculator, pressure would drop from 551 kPa (80 psi) to 241 kPa (35 psi), and the SD cylinder exhaust check valve would have opened. Per the ideal gas calculator, 0.015 moles of H2 at 400 K (720 R) in the SD cylinder with a volume of 0.219 L (13.39 cu in) would equal a pressure of 228 kPa (33.1 psi), or slightly higher than earlier estimates. Per the CGL calculator, increasing pressure from 228 kPa (33.1 psi) to 241 kPa (35 psi) will decrease volume to 0.211 L (12.8 cu in). That is generally in the ball park of what would be required to equalize pressure across the SD cylinder exhaust check valve. Over a 1st stage compressor piston area of 4.87 sq in, and a cylinder volume increase of an estimated 1 cu in, or a change in volume of 1/13.39th or 7.47%, that would equal a travel distance of ˜0.205″. Note that the same ˜0.205″ would be traveled by the SD cylinder as it compresses the working fluid mix entrapped within it. This would amount to a negligible amount of W-in by the SD cylinder before pressure is equalized across the SD cylinder exhaust check valve. Also, W-out by the 1st stage compressor is expected to more than balance W-in to the SD cylinder.

As the 1st stage compressor approaches BDC (for all SD CCVC models, see Point [9] on FIG. 19) during isochoric displacement, pressure and temperature will have dropped. Per FIG. 19, pressure in the SD cylinder, the regenerator, and the 1st stage compressor would drop to about 25 psi (172 kPa), assuming a complete displacement. Per the CGL calculator, isochoric expansion starts at 241 kPa (35 psi) with the vast majority of gas in the SD cylinder at a temperature of ˜400 K (720 R) and a volume of 219 L (13.39) cu in (not inclusive of the regenerator and manifolding). Assuming a final temperature within the 1st stage compressor of 234 K (421 R), final pressure would equal 141 kPa (20 psi). Pressure is expected to be slightly higher, however, since there is expected to be an additional 0.001 moles of H2 within the 1st stage compressor due to clearance gas carrying over from the previous stroke. (Note: The SD cylinder actuated inlet valve (Q), being at this point closed, prevents a “blowdown” in gas from the expander-to-SD manifold (not shown) into the SD cylinder due to the drop in pressure below 30 psi.) However, it is essential that pressure be equalized across the lower displacer exhaust valve (E) before that valve opens at or near BDC. Recall that pressure on the displacer side of the lower displacer exhaust valve is at ˜50 psi. In order to increase the pressure on the regenerator side of the lower displacer exhaust valve, the 1st stage compressor intake transfer valve (U) is closed early, as BDC is approached. This will immediately cause the SD piston to recompress any working fluid mix now captured between the SD piston, the lower displacer exhaust valve, and the 1st stage compressor intake transfer valve. In other words, the 1st stage compressor intake transfer valve will be designed to close at a point that allows movement of the SD piston to BDC to reach the pressure found in the lower displacer cylinder, or ˜50 psi.

Earlier, it was calculated that, in the regenerator and its manifolds at TDC, mol count equals 0.0052 moles at an average temperature across the regenerator of ˜311 K and 80 psi. As noted above, per FIG. 19, pressure in the SD cylinder, the regenerator, and the 1st stage compressor would drop to about 25 psi (172 kPa), assuming a complete displacement. Assuming a pressure nearing full displacement of about 25 psi (172 kPa), at an average temperature of 311 K (per above) in the regenerator and its manifolding (equal to 1.5 cu in or 0.0246), mol count in the regenerator and its manifolding, per the ideal gas law calculator, will drop to about 0.00164 moles. Following closure of the 1st stage compressor transfer valve at an average temperature of 311 K and 50 psi (345 kPa), per the ideal gas law calculator, mole count in the regenerator+manifolds or 1.5 cu in (0.0246) would equal 0.00328 moles. Since 0.00328 moles are required, approximately 0.00164 moles must remain in the SD cylinder following displacement and prior to reaching BDC, or just sufficient “dead space” moles to allow recompression back to ˜50 psi at BDC, thus allowing the lower displacer exhaust valve (E) to open without a pressure difference.

Per FIG. 19, the temperature of the entrapped working fluid in the SD cylinder just following 1st stage compressor transfer valve closure at about 25 psi (172 kPa) is about 525 R (292 K). Per the ideal gas law calculator, with a mol count of 0.00164, that would equal a volume of 0.00023 L. Since the total volume of both the SD cylinder and the 1st stage compressor cylinder equals 0.219 L (13.39 cu in), total volume of the 1st stage compressor cylinder at the time of the 1st stage compressor intake transfer valve is essentially unchanged at 0.219 L (13.39 cu in).

At the instant the 1st stage compressor transfer valve closes, assuming 0.015 moles pass through to the 1st stage compressor, and 0.001 moles remained in the 1st stage compressor cylinder clearance space at TDC, approximately 0.016 moles remain in the 1st stage compressor cylinder. From above, pressure would have dropped to about 25 psi (172 kPa), and volume would equal 13.39 cu in (0.219 L). Per the ideal gas calculator, temperature would thus equal ˜234 K (421 R), which is a close estimate of the temperature at the point in FIG. 19 (for all SD CCVC models, see Point [9] on FIG. 19).

Since the displacement of the working fluid mix from the SD cylinder and into the 1st stage compressor occurred at constant volume, only a negligible amount of net W-in or W-out occurs.

Mechanically, as BDC is approached, the SD cylinder actuated inlet valve actuator (not labeled) will allow the SD cylinder actuated inlet valve (Q) to open, since it is biased towards open. Note that a small amount of pressurization into the space between the SD cylinder actuated inlet valve and the SD cylinder intake check valve (AE) will then take place, increasing to about 50 psi. Further note that the SD cylinder intake check valve will not permit back flow into the expander-to-SD manifold (not shown). As shown in FIGS. 11 thru 17, the SD cylinder actuated inlet valve actuator is an arm physically touching the stem of the 1st stage compressor actuated exhaust valve (X). That SD cylinder actuated inlet valve actuator, which is physically connected to the 1st stage compressor actuated exhaust valve, closes the SD cylinder actuated inlet valve when the 1st stage compressor actuated exhaust valve closes slightly before TDC, as will be shown. Since the 1st stage compressor actuated exhaust valve opens just prior to BDC, the SD cylinder actuated inlet valve, which is biased towards open, will also be allowed to open just prior to BDC, assuming pressure differential allows it. However, note that, prior to the closing of the 1st stage compressor transfer valve, there will be a pressure differential across the SD cylinder actuated inlet valve head as a result of the lower pressure (˜25 psi) in the SD cylinder versus in the manifold space (˜30 psi) between the SD cylinder actuated inlet valve and SD cylinder inlet check valve (AE). As a result, the SD cylinder actuated intake valve itself will remain closed until pressure builds (<30 psi) in the SD cylinder sufficient to trigger the bias towards open, at which point the valve will automatically open. (Note: It is possible that a small amount of pressurized working fluid will escape past the SD cylinder check intake valve, but it should be minimal.)

With further movement of the SD piston to BDC, the working fluid mix remnant (within the SD cylinder, the regenerator and its manifolds, and the manifold connecting the SD cylinder intake check valve) will climb in pressure above the pressure acting to open the SD cylinder inlet check valve (lightly biased towards closed), closing that valve, if open. As pressure in the cylinder grows above ˜30 psi to ˜50 psi, the pressure in the expander exhaust manifold will be kept constant (at ˜30 psi) by virtue of the SD cylinder's inlet check valve.

As a result of recompression of remnant gas in the SD cylinder as the SD piston moves to BDC, the pressure in the SD cylinder will reach approximately the pressure in the lower displacer cylinder (˜50 psi), allowing the lower displacer actuated exhaust valve to be mechanically opened without incurring a pressure drop.

(Note: The push rod that operates the lower displacer actuated exhaust valve (not labeled) is “shortened” (that is, some cam lift is allowed to occur before the push rod contacts the lower displacer actuated exhaust valve) to allow the compression by the SD piston to largely go to completion at BDC before the lower displacer cylinder exhaust valve is actuated towards open. It is considered obvious that other means for successfully operating the lower displacer actuated exhaust valve are possible.

BDC to TDC—1st Stage Compressor Compression:

Note: The 1st stage compression is assumed to use purely adiabatic compression coupled with a purely isobaric exhaust processes. However, to the degree that the adiabatic process actually approaches isothermal, less work will be required for a given pressure to be achieved. It is expected that the walls and internal piston of the compressor will be chilled to the “ambient” temperature of 400 R, and thus some degree of heat transfer out of the compressing working fluid mix is expected during compression.

Per the CGL calculator, for an adiabatic compression of 0.219 L (13.39 cu in) of H2 from an initial temperature of ˜234 K (421 R) and an initial pressure of 25 psi (172 kPa), for a final pressure of 207 kPa (30 psi), final temperature would equal 247 K (445 R) and would equal a volume of 0.192 L (11.7 cu in) (for all SD CCVC models, see Point [10] on FIG. 19). In addition, internal energy change and required W-in would equal 0.0051 kJ. Assuming 1,130 rpm, W-in per hour for the compression process equals 346 kJ (0.096 kWh, 0.13 HP/hour).

In addition, as noted above, 0.001 moles of the 0.016 moles in the 1st stage compressor cylinder is “reserved” for re-pressurizing the STREP regenerator (B), meaning total exhausted volume is equal to 15/16ths of 0.192 L or 0.180 L (11 cu in). Further, the temperature of the 0.015 mol of exhausted charge, which equals 247 K (445 R), will be cooled to 222 K in the following isobaric cooler, reducing the exhausted volume, per the ideal gas calculator, to 0.134 L (8.18 cu in)

An isobaric exhaust process at −30 psi proceeds from 0.180 L (11 cu in) until slightly before TDC (for all SD CCVC models, see Point [11] on FIG. 19), when the 1st stage compressor actuated exhaust valve is closed. A small amount of remnant working fluid is thus captured within the 1st stage compressor, allowing that remnant fluid to be recompressed to the pressure on the regenerator side of the 1st stage compressor transfer valve (˜80 psi). The remnant fluid is assumed to equal 0.001 moles, thus leaving 0.015 moles to be exhausted per stroke of the 1st stage compressor.

Note that the 1st stage compressor intake transfer valve, similarly to the expander intake transfer valve, is opened by a combination of pressure differential equalization and spring bias, then mechanically closed, then kept closed by pressure differential. Also similarly, a slight mechanical “bash” by the 1st stage compressor piston top may be arranged just before TDC to ensure the 1st stage compressor intake transfer valve opens in a timely fashion.

Total volume of the 1st stage compressor piston equals 0.219 L (13.39 cu in). Following adiabatic compression, 0.180 L (11 cu in) remain, or 82.2% of total piston travel. Total piston travel is 2.75″, therefore, total travel at constant pressure equals 2.26″. Exhausting the working fluid mix at 30 psi over a 4.91 sq in area equals 147.3 lb force. Traveling along a stroke of 2.26″ or 0.188′ creates a W-in of 27.7 ft lb per stroke. Total 1st stage compressor isobaric W-in at 1,130 rpm equals 31,292 ft lb/minute or 0.948 HP/hr (0.71 kWh, 2,545 kJ). Total work in for the 1st stage compressor equals 346 kJ (0.096 kWh, 0.129 HP/hr) of adiabatic compression plus 2,545 kJ (0.707 kWh, 0.948 HP/hr) of isobaric exhaust, or 2891 kJ (0.803 kWh, 1.07 HP/hr).

BDC to TDC—2nd Stage Compressor Isobaric Intake:

The 2nd stage compressor (A, AA) is also assumed in this analysis to use purely adiabatic compression, in this case coupled with both an isobaric intake and an isobaric exhaust process. However, to the degree that the adiabatic compression process approaches isothermal, as by cooling of the cylinder and piston walls, less work will be required for a given pressure to be achieved.

The 2nd stage compressor cylinder receives its charge as the 1st stage compressor piston moves from BDC (for all SD CCVC models, see Point [9] on FIG. 19) to TDC (for all SD CCVC models, see Point [11] on FIG. 19), said first stage compressor piston simultaneously compressing and exhausting the H2 cooled by the regenerator through the 1st stage compressor actuated exhaust valve (X), through the 1st stage compressor exhaust check valve (Y), through the 1st stage compressor exhaust cooler (Z), through the 2nd stage compressor intake check valve (AB) and into the compressor cylinder. It therefore has a full charge at TDC at a pressure of −30 psi and a temperature of ˜400 R (222 K).

Importantly, the working fluid mix, upon exiting the 1st stage compressor exhaust cooler system, or some early portion of said system, can be seen to be below the temperature at which any H2O in the mix will have liquified and condensed out of the H2 working fluid. It is a simple matter to then completely separate the H2O liquid and the H2 working fluid. It is also a simple matter to add at some point sufficient new H2 working fluid at ˜400 R (222 K) and 30 psi to achieve the ˜0.015 moles of H2 required to continue the overall cycle. One potential H2O removal site is indicated in FIGS. 15 and 17. The 2nd stage compressor volume and piston area is formed by the space between the 2.5″ dia, 4.91 sq in (31.7 cm2) area, 13.5 cu in (0.221 L) volume of the 2nd stage compressor cylinder (A) and the 1.6″ (4.06 cm) dia, 2 sq in (12.9 cm2) area, 5.5 cu in (0.09 L) volume of the 2nd stage compressor piston connecting tube (AA). That equals a net volume of 8 cu in (0.131 L) and a net piston area of 2.91 sq in.

Per the ideal gas calculator, at 222 K, 30 psi (206.84 kPa), and 0.015 moles, volume will equal 0.134 liters (8.17 cu in) The 2nd stage compressor will thus closely match the 8.18 cu in 30 psi constant pressure output from the 1st stage compressor. (Note: the 1st stage compressor exhaust cooler ensures that the output remains at 222 K (400 R).) Note that the 2nd stage compressor cylinder piston (F) is on the lower side of the lower displacer-and-2nd stage compressor piston, while the lower displacer cylinder piston is on the upper side of the lower displacer-and-2nd stage compressor piston. Total piston travel of the 2nd stage compressor piston equals 2.75″, therefore, total travel at constant pressure equals 2.75″ or 0.229′. Receiving the working fluid mix at 30 psi over a 2.91 sq in area equals 87.3 lb force. Traveling along a stroke of 0.23′ creates a W-out of 20.0 ft lb per stroke. Total 1st stage compressor isobaric W-out at 1,130 rpm equals 22,600 ft lb/minute or 0.685 HP/hr (0.510 kWh, 1,840 kJ).

TDC to BDC—2nd Stage Compression and Isobaric Exhaust:

Per the CGL calculator, an adiabatic compression of 8.17 cu in (0.134 L) of H2 from an initial temperature of 400 R (222 K) and an initial pressure of 30 psi (207 kPa) (for all SD CCVC models, see Point [11] on FIG. 19), to a final pressure of 345 kPa (50 psi) and a final temperature of 257 K (463 R), would equal a volume of 0.093 L (5.68 cu in) (for all SD CCVC models, see Point [12] on FIG. 19). In addition, internal energy change and required W-in would equal 0.0108 kJ. Assuming 1,130 rpm, W-in per hour for the compression process equals 732 kJ (0.20 kWh, 0.27 HP/hour).

Exhaust W-out can be calculated by MEP over area over length of expansion. The MEP equals 50 psi. Area equals 2.91 sq in. Force equals 146 lb over an expansion of 0.229′. Force thus equals 33.32 ft lb. At 1,130 rpm, that equals 37,651 ft lb/min, or 1.141 HP/hr (0.851 kWh, 3,063 kJ). Total W-in of 2nd adiabatic expansion and isobaric exhaust thus equals 1,223 kJ (0.34 kWh, 0.46 HP/hr) (for all SD CCVC models, see Point [1] on FIG. 19). Further, the temperature of the 0.015 mol of exhausted charge, which equals 257 K (463 R), will be cooled to 222 K in the following isobaric cooler, reducing the exhausted volume, per the CGL calculator, to 0.080 L (4.86 cu in) (for all SD CCVC models, see Point [1] on FIG. 19).

BDC to TDC—2nd Stage Compressor to Lower Displacer:

From earlier, the upper and lower displacer cylinder volumes minus the displacer piston connecting tube volumes equals 4.86 cu in (0.0796 L) each. Since the stroke equals 2.75″, piston area equals 1.77″. With intake pressure equal to 50 psi, force equals 88.4 pounds. over 0.229′, W-out per stroke equals 20.2 ft lb. Over 1,130 rpm and 60 minutes, total ft lb/hr equals 1,371,951 ft lb, or 1,860 kJ (0.52 kWh, 0.693 HP/hr) (for all SD CCVC models, see Point [1] on FIG. 19).

Typical SD-CCVC Cycle Net Work and Thermal Efficiency:

• • SD-CCVC Isochoric/Adiabatic Cycle Expansion W-Out:
    • 1.465 HP/hr (1.092 kWh, 3,932 kJ)
  • SD-CCVC Isobaric/Adiabatic Cycle Expansion W-Out:
    • 1.89 HP/hr (1.41 kWh, 5,071 kJ)
  • SD-CCVC Isothermal Cycle Expansion W-Out:
    • 1.85 HP/hr (1.38 kWh, 4,970 kJ)
  • 1st Stage Compressor Compression W-In:
    • 0.129 HP/hr (0.096 kWh, 346 kJ)
  • 1st Stage Compressor Exhaust W-In:
    • 0.948 HP/hr (0.707 kWh, 2,545 kJ)
  • 2nd Stage Compressor Expansion W-Out:
    • 0.685 HP/hr (0.510 kWh, 1,840 kJ)
  • 2nd stage compressor compression W-in:
    • 0.27 HP/hr (0.20 kWh, 732 kJ)
  • 2nd Stage Compressor Exhaust W-In:
    • 1.141 HP/hr (0.851 kWh, 3063 kJ)
  • Lower Displacer Expansion W-Out:
    • 0.693 HP/hr (0.52 kWh, 1,860 kJ)
  • Total 1st and 2nd stage compressors W-in, all cycles:
    • 1.11 HP/hr (0.828 kWh, 2,979 kJ)
  • Total W-Out, SD-CCVC Isochoric/Adiabatic Cycle:
    • 1.465 HP/hr (1.092 kWh, 3,932 kJ)
  • Net W-Out, SD-CCVC Isochoric/Adiabatic Cycle:
    • 0.355 HP/hr (0.265 kWh, 953 kJ)
  • Thermal Efficiency, SD-CCVC Isochoric/Adiabatic Cycle:
    • Thermal source isochoric input: 3,726 BTU (3,932 kJ)
    • Thermal efficiency: 24.2%
  • Total W-Out, SD-CCVC Isobaric/Adiabatic Cycle:
    • 1.57 HP/hr (1.17 kWh, 4,204 kJ).
  • Net W-Out, SD-CCVC Isobaric/Adiabatic Cycle:
    • 0.464 HP/hr (0.346 kWh, 1,245 kJ)
  • Thermal Efficiency, SD-CCVC Isobaric/Adiabatic Cycle:
    • Thermal source isobaric input: 3,284 BTU (3465 kJ)
    • Thermal efficiency: 35.9%
  • Total W-Out, SD-CCVC Isothermal Expansion Cycle:
    • 1.77 HP/hr (1.32 kWh, 4,746 kJ)
  • Net W-Out, SD-CCVC Isothermal Expansion Cycle:
    • 0.65 HP/hr (0.49 kWh, 1,767 kJ)
  • Thermal Efficiency, SD-CCVC Isothermal Cycle:
    • Thermal source isothermal input: 4,498 BTU (4,746 kJ)
    • Thermal efficiency: 37.2%

Typical SD-CCVC Cycle H2 and O2 Mass Flows:

SD-CCVC Isochoric/Adiabatic Cycle:

1 mole H2 equals 2.0 grams. 0.015 moles equals 0.03 grams. Assuming the low heat of combustion, 1 kg of H2 has a combustion value of ˜120,000 kJ (33.33 kW-h), or 120 kJ/gram. For the SD-CCVC isochoric/adiabatic cycle, which requires 3,726 kJ/hr, that would require 31.05 grams of H2/hr being converted to H2O. O has 8× the mass of H2. Therefore, 248.4 grams of O2/hr would be required.

SD-CCVC Isobaric/Adiabatic Cycle:

For the SD-CCVC isobaric/adiabatic cycle, 3,284 kJ/hr, 27.4 grams of H2/hr and 219 grams of O2/hr would be required.

SD-CCVC Isothermal Cycle:

For the SD-CCVC isothermal cycle, 4,498 kJ/hr, 37.5 grams of H2/hr and 300 grams of O2/hr would be required.

Per hour, (38.4/2034=) 1.89% of the total cycled H2 working fluid will be combusted.

The BB Closed Loop Process.

The BB closed loop process essentially utilizes a thermochemical C6H12<=>C6H6+3H2 cycle similar to that disclosed in U.S. Pat. No. 3,225,538, but configures it differently, seeing the H2 generated by dissociation of a cyclical hydrocarbon such as C6H12 as “linked” with a second thermochemical and/or electrochemical cycle that associates and dissociates H2O. FIG. 22 below is based on FIG. 1 in U.S. Pat. No. 3,225,538, with temperatures in degrees Kelvin and pressure in atmospheres measured logarithmically to the base 10 (an insert graphs the base 10 into the righthand bottom of the figure). Per FIG. 22, a cyclical cyclohexane<=>benzene+hydrogen catalytic process at a given temperature and pressure will be either endothermic or exothermic. For example, at a pressure of 4 atm, the temperature for 90% endothermic conversion equals ˜900 K (1,620 R), while the temperature for 90% exothermic conversion equals ˜720 K (1,296 R).

Many forms of thermal energy can be used to dissociate the C6H12 into C6H6+3H2. However, a particularly interesting approach has been proposed whereby the required thermal energy can be generated by oxidizing a quantity of the H2 thus released.

1 kg of H2 has a mass of 2.20 Lb and thus a low heat value of 120,000 kJ/kg (33.3 kWMh, 44.7 HP/hr). Since ˜1,062 kJ are absorbed thermochemically in 1 Lb (0.4536 kg) of C6H12, or 2,341 kJ/kg, that means the combustion of 1 kg of H2 can theoretically convert 51.3 kg of C6H12 into C6H6+3H2. Since C6H6 has a mass of 78.11 g/mol and C6H12 has a mass of 84.16 g/mol, the conversion yields 6.05 g of H2 per mol of C6H12. 51.3 kg equals 609 mols of C6H12. Therefore, the mass of H2 released equals ˜3.7 kg.

That is, 27% of the 3.7 kg H2 released is burned to produce 73% or 2.68 kg of H2. Looking at the information under the heading “Typical SD-CCVC cycle H2 and O2 mass flows” above:

The SD-CCVC isochoric/adiabatic cycle thermal requirement equals 31.05 grams of H2/hr, or 1.12% of the H2 freed from 51.3 kg of C6H12. Therefore, total C6H12 required to fuel the cycle equals 0.59 kg per hour.

The SD-CCVC isobaric/adiabatic cycle thermal requirement equals 27.4 grams of H2/hr, or 1.0% of the H2 freed from 51.3 kg of C6H12. Therefore, total C6H12 required to fuel the cycle equals 0.524 kg per hour.

The SD-CCVC isothermal cycle thermal requirement equals 37.5 grams of H2/hr, or 1.40% of the H2 freed from 51.3 kg of C6H12. Therefore, total C6H12 required to fuel the cycle equals 0.72 kg per hour.

The SD-CCVC displacement expansion cycle thermal requirement will resemble the SD-CCVC isochoric/adiabatic cycle thermal requirement. It will be assumed that total C6H12 required to fuel the cycle will likewise approximate 0.60 kg per hour.

In a BB closed loop process, H2 and O2 will be generated from H2O, as by electrolysis. (Note: H2 and O2 can also be generated from H2O by thermal cracking.) The H2 is then used to convert C6H6 into C6H12, and also generate useful heat, as discussed below. Converting H2 into C6H12 permits storage of the H2 as a liquid at ambient pressure and temperature. The O2 may also be stored, either as a pressurized gas, as liquified oxygen (LOX), or within some form of thermochemical carrier. (One potential thermochemical carrier is hydrogen peroxide (H₂O₂). Note that H₂O₂ can also be stored as a liquid at ambient pressure and temperature.)

The C6H12 and possibly stored O2 are then transported to the power production site. At the site, the C6H12 will be run through an endothermic catalytic reactor, absorbing thermal energy and producing C6H6 and H2.

As noted above, one possible thermal energy source for converting C6H12 into C6H6+H2 is combustion of about a quarter of the H2 thus converted. U.S. patent Ser. No. 18/095,463 (applied for) illustrates several processes that can be made to produce net W-out from a C6H12>C6H6+3H2 reaction. In fact, such work-generating cycles were envisioned by the original B/E Cycle inventors, Reginald Bland and Frederick Ewing, and are disclosed in U.S. Pat. Nos. 3,225,538, 3,067,594, and, posthumously, U.S. Pat. No. 3,871,179.

It is anticipated that an SD-CCVC cycle engine, specifically designed for the main purpose of efficiently generate H2 from the catalytic dissociation of C6H12, and for a secondary purpose of efficiently generating W-out, may be powered by the combustion of 27% of the H2 gas thus produced. Such an SD-CCVC variant would be externally heating, by combustion of H2, an endothermic catalytic reactor. C6H12 would be dissociated into C6H6+3H2, creating half of a classic “Bland/Ewing Thermochemical Cycle” that can take full advantage of the molecular 4:1 expansion therein, first at constant pressure and temperature expansion, and then at adiabatic expansion. The exhaust from what is shown as the 1st stage compressor in FIG. 11 through FIG. 17 would principally be C6H6 and H2 gas (and some remnant C6H12). As will be shown below, hot C6H6+H2 exhaust can be directed into a C6H12 “vaporizer” heat exchanger. A new “charge” of C6H12 would be pumped into the other side of the heat exchanger at a lower pressure than the exhausting C6H6+H2 mix. Higher pressure C6H6 would condense at a higher temperature than that required to vaporize the liquid C6H12, thus utilizing the heat of condensing C6H6 for most of the required thermal input for the C6H12 vaporization process. The vaporized C6H12 would then enter a multistage and carefully intercooled compressor (ensuring that the vaporous content avoided condensation), returning the C6H12 vapor to the desired pressure for receipt by the lower displacer while approaching an isothermal compression. The lower displacer would then cyclically displace the pressurized C6H12 vapor through the regenerator at constant volume. Meanwhile, the condensed C6H6 (and C6H12 remnant) is easily separated from the H2, some of the separated H2 being then potentially used as fuel to “power” the engine, and the remainder of the separated H2 then being available for other uses.

(Note: A possible side benefit of using such a working fluid would be the possibility of using a small amount of liquid C6H12, or possibly C6H6, to help lubricate the engine. That in turn may permit higher operating temperatures. Any lubricant “leakage” would simply be carried through the engine and recaptured in the exhaust.)

One simple system for storing and delivering C6H12 is shown as FIG. 23 in this application. Note that W-in to the hydraulic pump is the total W-in required to feed the system, and that a large degree of the required hydraulic pump W-in is returned as W-out by the hydraulic motor shown receiving the dissociated C6H6 in pressurized liquid form. While the system shown in FIG. 23 appears to be very wasteful of energy, note that the output is not work, but high pressure H2 gas. In effect, the “work out” is H2 that is in effect “pressurized” by the endothermic conversion process, with a given pressure being determined by the temperature at which the endothermic conversion is made to take place. With H2 combustion, very high peak temperatures are easily achieved, thus very high pressures can be achieved as well. In this sense, the catalytic dissociation of C6H12 into high pressure C6H6 and H2 can be seen as a potential means of thermochemically compressing H2 gas by a means that bypasses the need for mechanical compression. This is particularly useful in potentially high pressure heat engines such as the SD-CCVC designs proposed herein. Here are the steps:

• • a. Liquid reactant (C6H12) at ambient temperature and pressure is pumped out of storage, for example in a cylinder possessing a double-acting piston, on one side of which is stored the liquid reactant and on the other side of which is stored the liquid product (C6H6 and remnant C6H12).
  • b. The liquid reactant is pumped to 100 atm and exhausted into a recuperator (or regenerator such as a valved regenerator.
  • c. The recuperator which will preheat, vaporize, and superheat the reactant to approximately the temperature of the endothermic reactor. To accomplish this preheat, it will transfer heat to the reactant from the product exhausting from the endothermic reactor.
  • d. The endothermic reactor will convert the reactant (C6H12 superheated vapor) into the product (superheated C6H6+H2 and remnant C6H12) with heat it receives from a high temperature heat source. That high temperature heat source can be a portion of the H2 produced and/or excess heat generated by an operation such as the combustion of H2.
  • e. The product is flowed to the other side of the recuperator or regenerator.
  • f. The product is cooled under pressure sufficiently to separate the H2 (and other gases, if any) from the C6H6 and remnant C6H12 (and other liquids, if any).
  • g. The liquid products are flowed through a hydraulic expander, reducing them to ambient pressure.
  • h. The liquid products are sent to the other side of the double-acting piston, where they are stored. The stored products will eventually be exchanged for a fresh charge of reactant.
  • i. The high pressure gases are flowed to their destination, for example to a reheater (either constant pressure or constant volume) and then to an H2-burning heat engine.

(It is also possible to envision such a system for converting pressurized H2O2 into H2O plus highly pressurized O2. However, in H2O2, the dissociation reaction is highly exothermic. Under the circumstances, it might be best to simply inject pressurized H2O2 directly into the H2 stream, possibly without any preheating.)

Following endothermic conversion, the C6H6 and H2 product will be cooled to the point where the C6H6 is liquified and separated from the H2 gas. The H2, and possibly previously stored O2. are then converted back into H2O, either in a fuel cell or by combustion, essentially releasing the energy stored when they were generated in the first place.

FIGS. 24A, 24B, and 24C illustrate how a fluid, for example liquid C6H6 [1], can be captured in a tank [2] that is holding, for example, liquid C6H12 [3], as the original C6H12 is “emptied” and converted into C6H6, showing top, middle, and bottom positions of the “charging piston”.

Illustrated in FIGS. 24A, 24B, and 24C is a simple “charging” piston [4] and cylinder [2] arrangement. As will be shown, this charging arrangement permits charging or discharging of fluid [1 and 3] moving into and out of the displacer cylinder. In this instance, a connecting rod [5] is shown running through the charging cylinder, connecting to the double-acting charging piston internal to the charging cylinder (charging cylinder connecting rod seals are not shown). Note: The rod does not have to be double-sided, but doing so illustrates that any pressure acting on one side of the piston is perfectly equal to the pressure acting on the other side (barring friction).) The double acting charging piston is shown with a sealing ring [6] the sealing ring arrangement around its circumference, which can allow a differentiation of the fluids on either side of the piston. The charging piston is used to create the direction of constant pressure fluid flow, being typically moved by some small force in one direction or the other. It thus represents a simple directional pump.

From U.S. patent application Ser. No. 18/197,902, section “Specification—Miscellaneous Descriptions and Operations” entitled “The Benzene Battery Cycle”: “It is obvious that the BB cycle energy storage and delivery process has a potential usefulness beyond the lunar surface. In fact, it can easily be shown to represent a meaningful alternative to the present hydrocarbon-combustion processes that currently underpin much of the human race's energy generation and distribution network.” The application then goes on to describe the following means: “The service station fills a transport's tank with C6H12 while emptying the same tank of C6H6 (a partition keeping the two liquids separate from one another).

It is anticipated that the generated C6H6 (plus any remnant C6H12) [1] would then be returned to the tank [4] holding the C6H12 [2], with the two liquids being kept separate, as by the use of a piston. For a complete system, a vehicle receives a fresh charge of C6H12 at a “service station” in exchange for the liquid C6H6 and remnant C6H12. The received liquid is then shipped back to a H2 production plant for conversion back into C6H12.

C6H6 as a Means for Capturing and Transporting Potential Thermal Energy.

In supplying the thermal energy to release the H2 gas from the C6H12, thereby converting the C6H12 to C6H6, note that the total amount of dissociation thermal energy required is “captured” chemically within the C6H6 (assuming a perfect thermal recuperation, see FIG. 23). That thermal energy will be “released” exothermically when the C6H6 is converted back into C6H12 by the absorption of a new supply of H2. Note that the temperature of the thermal release can be increased or decreased by simply adjusting the pressure of the C6H6+3H2 mix prior to the mix entering the exothermic reactor, as shown in FIG. 22. In the Benzene Battery closed loop system, upon being returned to the thermal source power plant (such as a solar thermal power plant), the thermal energy generated by the recombination of H2 and C6H6 can be made available to the thermal power plant, thereby increasing its overall efficiency.

A simple heat generating system is shown in U.S. patent Ser. No. 18/095,463, FIG. 14, shown as FIG. 25 in this application. This system makes use of a technique termed an Exothermic Reactor Exhaust Compressor (EREC), disclosed in U.S. patent Ser. No. 18/095,463. In essence, the EREC, by raising the relative pressure of the C6H12 and remnant C6H6 exiting the exothermic reactor, permits thermal energy of condensation from that C6H12 and remnant C6H6 to supply much if not all of the required energy of vaporization for the inflowing liquid C6H6. The result is the generation of a net thermal energy output of 90 kJ from a conversion of 0.038 kg of C6H6+H2 to C6H12 which yielded 98.3 kJ of thermal energy. The cost, which is essentially the W-in required to operate the EREC, is a theoretical power input of about 1.2 kJ.

U.S. patent Ser. No. 18/095,463 also proposes several alternative approaches termed “Exothermic production cycles. There are three use cases for this thermochemically produced heat; cogeneration or combined heat and power (CHP), combined cycle (CC), and hybrid CHP/CC. The use cases differ primarily in the relative percentage of power output versus thermal output.

It is also possible to use the released exothermic thermal energy to power, or aid in powering, an SD-CCVC cycle engine. In FIG. 22, it is shown that an exothermic reaction at only ˜50 psi (345 kPa) will generate well over the 555 K (1,000 R) temperature regime explored in the above calculations. Thus, the C6H6+H2 reaction itself can easily power an SD-CCVC cycle engine. A very low temperature SD-CCVC displacement expansion engine, for example, has been shown herein to theoretically generate power at a thermal efficiency in the range of 16%, and much higher peak temperatures and hence efficiencies appear to be possible.

Finally, another alternative that uses an SD-CCVC cycle involves replacing the O2 injected in the engine designs described herein with vaporized C6H6, while leaving the pressurized working fluid as H2. Generally speaking, an exothermic catalytic reactor would be added just before the SD-CCVC expander, replacing the manifold labeled “injector site” shown in FIG. 16 and FIG. 17. Vaporized and pressurized C6H6 would be cyclically injected into a primarily H2 working fluid at approximately the point where working fluid is passing out of the upper displacer cylinder. The working fluid mix would then be passed through the catalytic reactor, endothermically converting the C6H6 to C6H12, thus supplying the thermal energy to drive the engine by internal catalytic heating. The C6H12 would then be liquified and removed from the working fluid at approximately the point where H2O is shown removed in FIG. 15 and FIG. 17.

The SD-CCVC cycle proposed for use with this alternative is the SD-CCVC isobaric/adiabatic expansion cycle. A constant pressure and temperature heat addition needs to be maintained during the injection of the vaporized and pressurized C6H6 as it passes into the exothermic catalytic reactor, although there is a possibility of a small amount of isochoric exothermic chemical heat addition slightly in advance of the expansion. Such a cycle can be termed a benzene-fueled SD-CCVC isobaric/adiabatic cycle.

As in SD-CCVC cycles described above, the H2 gas will be “densified” using multi-staged and intercooled compressors to approach isothermal as closely as practicable. Waste exhaust heat will be added via the STREP process. And a highly pre-pressurized H2 working fluid is also possible, increasing power density. Pre-pressurizing the H2 working fluid will also increase the temperature of the catalytic exothermic recombination of C6H6 vapor and H2.

Ideally, pressurized C6H6 will be vaporized using the EREC process as shown in FIG. 25 and described in U.S. patent Ser. No. 18/095,463. Note that care will need to be taken to exhaust the working fluid mix from the 1st stage compressor before the vaporous C6H12 begins to condense out. The working fluid mix would then exhaust into a C6H6 “vaporizer”, utilizing the heat of condensing C6H12 for most of the required thermal input for the vaporization process. This process is described in detail in U.S. patent Ser. No. 18/095,463. That would permit the C6H6 to be pumped into the system at high pressure, greatly increasing overall net W-out and thermal efficiency.

In designing a prototype SD CCVC engine that best utilizes the existing CCVC prototype, several new applications and improvements have been discovered.

Note: To develop a first order analysis for various proposed cycles, an ideal gas law calculator and the CGL calculator will be used. To prepare a visual estimate of various cycles, these calculations will be used in conjunction with FIG. 18, taken from FIG. 70, “Marks' Mechanical Engineers' Handbook”, 6th edition, 9-148, “Internal-combustion engines”. FIG. 27 shows lines traced onto FIG. 18 and then moved from the tracing for clarity purposes, representing various states of working fluid for various cycles. FIG. 27 also includes letters representing various points for the various states. See “1st Cycle analysis”, paragraph 10, below. The solid vertical lines in FIG. 27 represent zero heat transfer (isentropic/adiabatic) changes, the solid non-vertical curves represent constant volume (isochoric) changes, and the dotted non-vertical lines represent constant pressure (isobaric) changes. In FIG. 18 and FIG. 27, temperatures are in degrees Rankine, pressures are in pounds per square inch (psi), energy and enthalpy are in Btu's, and volumes are in cubic feet. To aid in use of the ideal gas law and CGL calculators, temperatures will also be shown in degrees Kelvin, pressures in kiloPascals, energy in kiloJoules, and volumes in liters. Heat and power will be shown in kiloJoules and Btu's.

Calculations herein concerning endothermic and exothermic dissociation and re-association of a mol of cyclohexane (C6H12) into a mol of benzene (C6H6) and three moles of hydrogen (H2) will utilize FIG. 1 from U.S. Pat. No. 3,225,538, shown as FIG. 22 herein. As shown in FIG. 22, in the presence of a proper catalyst C6H12 will move towards C6H6+3H2 if temperature remains constant and pressure is increased and in the opposite direction if pressure decreases. Likewise, if pressure is constant, C6H12 will move towards C6H6+3H2 if temperature is decreased and in the opposite direction if it is increased. FIG. 22 indicates the percentage of conversion in a given direction, with the line marked, “99” representing a 99% C6H12 content and the line marked 0.01 representing a 99% C6H6+3H2 content. Finally, note that this particular reaction is highly stable and thus highly cyclical. However, it has been established that some degradation will occur over time, with conversion into other forms of hydrocarbon.

Finally, note that all proposed cycles and mechanisms are being shown solely to describe the inherent nature of the broader inventions being herein disclosed. Thus, in FIGS. 11 through 17, the expander cylinder, the synchronizer, the compressor or compressors, and the displacers are shown as separate modules that interact. Alternatively, as will be shown herein, a single crankshaft throw may be used for all modules, with the various drive rods of the various modules physically connected. Also, shown embodiments of the proposed cycles, such as starting temperatures, pressures, volumes, proposed working fluids, etcetera, do not innately describe ideal embodiments but are to be considered first order estimates. Much actual reduction to practice and experimentation will be required to determine that.

Additional Improvements:

• • (1) In some versions, as will be shown, the isochoric exhaust displacer or “synchronizer cylinder” can be easily converted into a secondary expander cylinder, and thus will have a larger volume than the primary expander cylinder. Note that, to permit true isochoric exhaust displacement, a third cylinder, that is, the cylinder that receives the exhaust from such a secondary expander cylinder, will have the same volume as the secondary expander cylinder.
  • (2) A secondary addition of source heat following the primary expansion of the working fluid is also possible, as will be discussed below. Such additional heat may be added prior to exhausting into a secondary expander cylinder. Note that additional source heat can be added at constant volume, pressure, or temperature, or at some combination thereof. Also note that such a secondary heat addition is well known to those skilled in the art to be thermodynamically beneficial.
  • (3) It has been found to be useful to have two or more isochoric thermal input displacement means performed in series. For example, the first displacement can utilize lower temperature exhaust waste heat. A second displacement can then utilize higher temperature H-in, for example from a thermochemical catalytic reaction, such as the exothermic catalytic reformation of benzene (C6H6) plus hydrogen (H2) into cyclohexane (C6H12). In one such means, the cylinder presently shown as the 2nd stage compressor cylinder in FIGS. 11 through 17 can be used instead as a new lower displacer cylinder, the lower displacer cylinder shown in FIGS. 11 through 17 will become the middle displacer, and the upper displacer shown in FIGS. 11 through 17 will remain the upper displacer. Between the lower and middle displacer cylinder will be the lower temperature regenerator, and between the middle and upper displacer cylinder will be the higher temperature regenerator. As will be shown, two different approaches have been considered for thermally charging the higher temperature regenerator with heat produced by an exothermic reactor. The first approach would heat a thermal transfer fluid (such as He or H2) via a heat exchanger used to cool the exothermic reactor, said transfer fluid then being intermittently flowed through, and thus “charging”, the regenerator. The second approach, perhaps more useful for generating useful work, would add excess cooler thermal transfer fluid at the same pressure to the endothermic fluid as it is being created by said exothermic reaction, essentially cooling the reactor internally. This may require multiple injection points along the reactor body to avoid over-cooling.
  • (4) Rather than performing “displacement expansion” or non-adiabatic expansion by exhausting from the final displacer through a heater and directly into the expander cylinder as proposed in FIGS. 11 through 17, the working fluid exhausting from the final displacer cylinder can isobarically exhaust. This permits “decoupling” of the displacer system from the expander system so that the displacer system and the expander system don't have to be phase-locked to one another. It also permits isobaric source heat addition, allowing a large distance to exist between the final isochoric displacer and the expander. That in turn would permit, for example, a thermal solar energy system of large volume, such as a trough-type solar concentrator, to be used as the thermal source for the heat engine. The primary expander can then utilize, for example, a combination isobaric/adiabatic expander, akin to a diesel cycle, whereby:
  • A. During the first portion of expansion, a “charge” of working fluid exhausted from the upper displacer at constant pressure will be taken into the primary expander at constant pressure via the existing intake/transfer valve.
  • B. During the second portion of expansion, the charge of constant pressure working fluid thus captured will then be expanded adiabatically. For example, following a degree of isobaric expansion, the intake/transfer valve can be instantaneously closed, as by the action of a solenoid, thus instantaneously converting the expansion into an adiabatic one.
  • (5) Isobaric heat addition is useful, but heat may also be added isochorically, isothermally, or some combination of the three. Note that, in the graphs indicating working fluid states shown in FIG. 27, multiple possible configurations are shown, including isobaric and isochoric configurations. Isothermal source H-in is also possible.
  • (6) A “valved cell” can be the means for attaching pre-heated and/or pre-pressurized fluids into the expander cylinder, via the concepts disclosed in U.S. Pat. Nos. 4,817,388 and 5,179,839.
  • (7) STREP heat exchangers can be seen as applicable to what can be termed the Bland/Ewing Composite Cycle (B/E-CC) process. As stated in Pending U.S. patent Ser. No. 18/095,463: “The underlying improvement to the foundational B/E Cycle invention disclosed in U.S. Pat. Nos. 3,225,538, 3,067,594, and 3,871,179 takes the form of apparatuses for combining various independently operating or semi-independently operating endothermic and exothermic half-cycles which coupled together form a complete B/E Cycle.
  • (8) In U.S. patent application Ser. No. 18/095,463, use of an Exothermic Reactor Exhaust Compressor (EREC) is proposed to assist in the vaporization of C6H6 by a counter flowing exchange of heat with condensing higher pressure C6H12. It is herein proposed that an Endothermic Reactor Exhaust Compressor (which will be termed an ENREC) be used to permit the condensation of higher pressure C6H6 to supply some or all of the thermal energy required to vaporize a similar mol count of C6H12.
  • (9) FIGS. 11 through 17 proposed increasing the efficiency of a isochoric counterflow recuperator, such as was used in the original Closed Cycle Valved Cell (CCVC) prototype as described therein, by using (1) an expander exhaust receiver cylinder of equal volume to said expander, termed an SD cylinder, (2) valving and connecting manifold means that intermittently and cyclically connected said SD cylinder to (3) a regenerator, and (4) valving and connecting manifold means that intermittently and cyclically connected said regenerator to an equal volume second displacer cylinder. Note that such a means meets the general definition of a STREP. Various additional modified STREP designs are herein proposed.

For example, a modified STREP can be used as means of exchanging heat between a low pressure high volume “receiver” mechanism and a high pressure low volume “displacer” mechanism, as shown in FIGS. 1 through 4. Note that the higher volume flow of the “heating” fluid would be due to the difference in density required for a given amount of heat to be transferred. It can be desirable for a high temperature but low pressure exhaust fluid (for example, exhaust from a heat engine) to be passed at constant pressure through a counterflow regenerator-type heat exchanger, thermally “charging” the regenerator. The regenerator can then be raised in pressure, for example by a semi-adiabatic compression of remnant fluid in the “receiver” mechanism such as would be caused by an early closure of the regenerator's exhaust valve. A separate stream of counter-flowing fluid at the higher pressure but at low temperature can then enter the regenerator through an intake valve and flow isobarically through the regenerator via the displacer mechanism, thus isobarically absorb the thermal energy deposited by the earlier low pressure flow. Finally, the high pressure in the regenerator can be reduced to that of the low pressure stream and once more be used to “charge” the regenerator, as by re-expansion of remnant high pressure fluid in the regenerator back into the low pressure displacer at the beginning of its intake stroke, followed by low pressure fluid intake into said low pressure displacer.

In addition, the high pressure, high temperature fluid thus generated can then be used to isobarically “charge” an isochoric regenerator up to the peak temperature of the low pressure fluid. That is, a modified STREP may be used to thermally charge a “valved regenerator” with an isobaric fluid, then “switch” the regenerator via valving in order to isochorically remove some or all of the thermal charge thus deposited isobarically. Note that this is ideal for capturing high temperature thermal energy from the low pressure exhaust of a heat engine. For example, a high pressure, high temperature fluid can be passed at constant pressure through a counterflow regenerator-type heat exchanger, whereby the exhausting fluid's thermal energy can thus thermally “charge” the regenerator. A separate stream of counter-flowing fluid at the same pressure but at low temperature can then enter the regenerator through an intake valve. If the flow were made to be isochoric, the counter-flowing fluid would absorb the thermal energy deposited by the earlier flow at constant volume, as by displacement from an initial cylinder, through the regenerator, and into an SD cylinder, thus theoretically both raising the temperature of the counter flowing fluid to the peak temperature of the thermal charging fluid, and thermally rather than compressively raising the pressure of the counter-flowing fluid. Following the thermal charging of the higher pressure counter-flowing fluid, the SD cylinder exhaust valve would close and the remnant gas in the regenerator would be expanded back into the displacer cylinder until it dropped to the pressure of the thermal charging fluid, at which point the regenerator intake valve would close, the regenerator exhaust valve would open, and the cycle would begin again.

1st Cycle Analysis

FIGS. 29a through 30c show different cutaway views of a CAD solid model of one possible approach to converting the existing CCVC prototype engine into a prototype engine with the above characteristics. Several assumptions are made:

• • (1) Initially, the existing prototype will be used as much as possible, which “locks in” certain volumes and temperatures. These are defined above and associated with FIGS. 11 through 18.
  • (2) Ideally, the proposed engine will use multiple intercooled compressions. A “sink” or low cooled temperature of approximately the boiling point of water is assumed.
  • (3) The proposed engine will assume use of a single expander, although a second expansion is possible, as proposed above.
  • (4) In the proposed prototype engine, there will be two isochoric displacements through two different heaters via three displacer cylinders. The upper displacer piston, which will operate in a higher peak temperature environment, will be formed with an upper piston extension which will permit the upper piston seal to mount a teflon/spring seal in the upper cylinder wall. Mounting the seal in the cylinder wall will allow cooling of the seal via the cylinder wall area that is in physical contact with the seal, permitting a potentially higher working fluid temperature.
  • (5) In the proposed prototype engine, there will be an isobaric displacement through a third heater into an expander valved cell working fluid injector. There will then be an adiabatic displacement via the valved cell expander injector into the primary expander. That will permit testing of additional cycles, as will be shown.
  • (6) Once in operation, the proposed prototype is expected to be capable of testing at much higher base pressures, much as a classic stirling engine may be operated at very high base pressures. Note that this “pre-pressurization” will not measurably impact the various temperatures within such a pre-pressurized engine, just as in a classic stirling engine. What will be impacted is power density per cycle, which is an extremely important element in achieving workable power outputs and thus potentially approaching theoretical potential. That is because, for any set of given inefficiencies from friction, pumping, or thermal losses, a higher power density decreases these losses relative to the actual net power produced.

In the proposed prototype engine, it will be initially assumed that the volume of each of the three displacer cylinders will be comprised of a 2.5″ diameter cylinder containing a 2″ diameter connecting tube over a displacer piston stroke of 2.75″, or an effective piston area of 1.767 sq in and thus a cylinder volume of 4.86 cu in (0.0796 L).

As previously stated, FIG. 27 includes letters representing various points for the various working fluid states. These will be referenced as this analysis proceeds.

It is estimated that, at 671 R (373 K, 100 deg C., 212 F) and a pressure of 1 atm, the volume is increased to 497 L/min (17.6 cu ft/min) (Point A). Following stage one compression (Point B) and intercooling (Point C), per the CGL calculator, the 497 L/min charge of working fluid will be assumed to be compressed to about 25 psi (172.4 kPa) at 671 R (373 K), decreasing volume to 353.7 L/min (12.5 cu ft/min). Following stage 2 compression (Point D) and inter cooling (Point E), the charge is returned to 373 K (671 R) at a pressure of about 74 psi (510 kPa), dropping the volume per the ideal gas calculator to 103.8 L/min (3.667 cu ft/min, 6,337 cu in/min).

Beginning at BDC, the lower displacer cylinder will receive a constant pressure charge of 74 psi (510 kPa) working fluid at 373 K (671 R). That will equal 1/1,304th of the 17 cu ft/minute flow, indicating that the engine would need to have a rotational speed of 1,304 rpm to pass through 17 cu ft/minute. Per the CGL calculator, at a final temperature of 373 K the final flow volume per minute into the lower displacer equals 3.67 cu ft (103.8 L), or 4.86 cu in/cycle (0.0796 L/cycle) @ 1,304 rpm), internal energy change equals 18.5 kJ, W-in (of exhaust) equals 8.1 kJ (7.6 Btu), and rejected heat equals 27.9 kJ. Per the ideal gas law, the H2 flow rate equals 17.07 moles/min or 0.0131 moles/cycle. Since 16.24 moles pass through every minute, the total charge inducted into the lower displacer cylinder would equal 0.0124 moles. At STP, H2 (gas) has a mass of 2.02 g/mol. 0.0124 moles of H2 thus equals 0.026 g or 26 mg.

At TDC, following intake of 0.0131 moles of working fluid at 373 K and 172 kPa into the lower displacer cylinder, the working fluid will be exhausted completely into the #1 regenerator and the middle displacer cylinder. Since the two displacer cylinders have the same displacement and the same stroke, the displacement occurs isochorically. Since thermal energy is added, the temperature of the working fluid following isochoric displacement via the following move to BDC will ideally reach the temperature of the exhaust fluid charging the #1 regenerator, or 840 R (466 K). Per the CGL calculator, the pressure will increase to about 648 kPa (94 psi) (Point F).

Note that the thermal energy source may be otherwise-waste heat from the engine exhaust.

During the following move to TDC, the middle displacer will then exhaust completely into the #2 regenerator and the upper displacer cylinder. Since the two displacer cylinders have the same displacement and the same stroke, the displacement occurs isochorically. Since thermal energy is added, the temperature of the working fluid following isochoric displacement via the following move to BDC will reach the approximate temperature of the exhaust fluid charging the #1 regenerator, or about 984 R (547 K). Per the CGL calculator, the pressure will increase to about 758 kPa (110 psi) (Point G).

Note that the thermal energy source may be thermochemical heat released by the conversion of C6H6 plus H2 into C6H12.

In the following model, the temperature of the proposed prototype will not exceed 1,140 R (633 K, 360 deg C., 680 deg F.). That temperature will be limited to the primary expander and the primary expander constant pressure injector, as will be shown. To accomplish this, source heat will be added at constant pressure. The heat exchanger exterior to the upper displacer will raise the working fluid temperature to 1,140 R (633 K).

The isobaric H-in will occur as the working fluid stream exhausted at constant pressure from the upper displacer is subsequently heated to a higher temperature. Temperature during constant pressure upper displacer exhaust will occur at about 547 K (984 R). Consequently, the upper displacer exhaust will be through a high temperature exhaust check valve. A non-lubricated externally cooled teflon bearing exhaust check valve guide should be possible if it is sufficiently physically removed from direct contact with the working fluid. In addition, to allow this higher temperature without using a lubricant, the upper displacer piston will be lengthened, such that a teflon/ss spring seal can be mounted in the cylinder wall. With external cooling of the cylinder wall in immediate proximity to the teflon seal and with the standoff distance of the lengthened piston, the upper displacer piston will likely not require lubrication.

The initial receiver for the 633 K, 110 psi working fluid will be a “valved cell”. The physical nature of the cell will be that of a piston and cylinder, with the piston sharing the 2.75″ stroke of the overall engine. It will be positioned on top of the expander cylinder and its charge of working fluid will be connected to the expander cylinder via a “transfer valve” which will operate similarly to that on the existing CCVC prototype, as will be shown. The valved cell will also possess a mechanically operated high temperature inlet valve. Non-lubricated externally cooled teflon bearings for the intake and exhaust check valve guides should be possible if the valve guide seats are sufficiently physically removed from direct contact with the working fluid. Between the upper displacer exhaust check valve and the valved cell intake valve, an isobaric high temperature heat source heat exchanger will be situated. It may be partially or completely composed of the existing CCVC prototype heater and/or partially or completely composed a solar concentrator such as the original Bland/Ewing Cycle parabolic trough solar concentrator. As noted above, volume equals 3.80 cu ft at 1,140 R, closely matching the relevant graphed line in FIG. 27.

In operation, the valved cell would act as a constant pressure expander as it is charged via its intake valve, as the valved cell piston is moving from BDC to TDC. The transfer valve into the expander would then be opened at or around TDC, and the working fluid would exhaust in an adiabatic displacement expansion into the expander piston, resulting in rapidly falling pressure and temperature. Towards the end of the expansion stroke, the transfer valve would close, allowing a small amount of gas to be trapped and pressurized up to 110 psi. On the move back to TDC, the valved cell intake valve would then be opened, allowing the valved cell to be charged once again with H2 working fluid at 110 psi and 1,140 R (633 K) (Point H).

To allow this higher temperature without using a lubricant, both the expander and the valved cell pistons will be lengthened in the same manner the upper displacer piston is lengthened, such that a non-lubricated teflon/ss spring seal can be mounted in the cylinder walls of each. Note that, with external cooling of the two cylinder walls in immediate proximity to the teflon seals and with the standoff distance of the lengthened pistons, the valved cell displacer and the expander will likely not require lubrication at these temperatures.

To compensate for the potentially significant added mass of the much longer pistons, the crankcases of the upper displacer, valved cell, and expander cylinders will be pressurized to a higher pressure than the peak pressure developed within the three cylinders, and be held at constant pressure. Consequently, the piston walls can be made quite thin, and piston mass can be held down appreciably. Note that it is possible to accomplish some cooling of the internal portions of the three pistons by this compressed crankcase gas, thus permitting a higher peak temperature for the non-lubricated engine.

In the primary expander, which has an expanded volume of 13.5 cu in/cycle (0.221 L/cycle), or 288.5 L/min (10.19 cu ft/min), an adiabatic expansion process will be assumed. A charge per cycle, equal in mass to the charge exhausted from the upper displacer, will be injected into the expander via displacement. Dead space at the completion of the expander exhaust will be considered nearly equal to zero. That equates to an expansion ratio of 13.5/4.86 or 2.777 to 1 (Point I).

The expander charge would then essentially completely exhausted at constant pressure and temperature into a SD cylinder, which has a close volumetric match to the expansion cylinder. Since pressure and temperature are maintained, no work either in or out is done.

Following synchronizer piston intake, the charge will be essentially completely exhausted into a displacer/1st stage compressor cylinder with matching piston diameter and stroke. That creates an essentially isochoric displacement process that can be timed to pass through and thus charge the regenerator between the lower and middle “compression” displacer. Note that this can be thought of as a kind of constant volume displacement “decompression”, since the working fluid pressure is lower following displacement due to the removal of heat by the regenerator. The result is the charging of the regenerator with waste heat from the expander and the simultaneous cooling of the isochorically-displaced charge exhausted by the synchronizer, dropping its temperature to approximately 671 R (373 K) at a pressure of about 25 psi (172.4 kPa). Volume remains 10.2 cu ft (Point A).

W-in, W-out, H-in, and theoretical thermal efficiency can now be estimated.

Exhaust from the expander from BDC now proceeds at constant pressure, temperature, and volume by means of the synchronizer cylinder. Since all states remain unchanged, the only negatives are friction, pumping losses, and thermal leakage, which are assumed to equal zero. By TDC, the expander has displaced all but a small amount of clearance gas into the manifold and from thence into the synchronizer cylinder. Closure of the expander exhaust valve slightly early allows the expander clearance gas to be pressurized to equal that of the gas that, in the previous stroke, charged the valved cell injector, opening the transfer valve at approximately TDC.

The synchronizer piston now exhausts to BDC through the valved regenerator and into the 1st stage compressor at constant volume, charging the regenerator with heat and isochorically cooling the gas displaced from the synchronizer and into the 1st stage compressor. Per the CGL calculator, volume equals 9.71 cu ft (275 L) and stays constant, pressure drops from 36.5 psi (252 kPa) to 29.2 psi) (201 kPa), and temperature drops from 466.25 K (840 R) to 671 R (373 K), releasing/storing 34.1 kJ (32.3 Btu) of thermal energy via the regenerator. (A negligible amount of Wout is produced at the end of this process, as was described in FIGS. 11 through 17, and above in “TDC to BDC—SD cylinder to 1st stage compressor via regenerator” through “BDC to TDC—1st stage compressor compression”.)

The 1st stage compressor now ascends towards TDC, adiabatically compressing the captured gas. Using the CGL calculator for an adiabatic compression from 29.2 psi (201 kPa), 373 K, and 9.71 cu ft (275 L), to 47 psi (324 kPa) and 428 K (770 R), final volume equals 6.92 cu ft, internal energy change and W-in equal 20.0 kJ (19.0 Btu).

At 47 psi and part way through the upward stroke, the 1st stage exhaust check valve opens and an isobaric exhaust occurs through a cooler. Per the CGL calculator, at 373 K, the final flow volume equals 6.03 cu ft (170 L)/minute, internal energy change equals 20.1 kJ (19.0 Btu), W-in equals 8.2 kJ (7.7 Btu), and rejected heat equals 28.3 kJ (26.82 Btu). (A negligible amount of W-in is required at the end of this process, as was described above.)

The 2nd stage compressor now descends towards BDC, compressing the captured gas from 47 psi, 373 K, and 6.03 cu ft to 425 K (770 R) and 74 psi (510 kPa). Using the CGL calculator for an adiabatic compression, final volume equals 4.37 cu ft (123.6 L), and internal energy change and W-in equal 19.1 kJ (18.1 Btu).

Per the CGL calculator, at 74 psi, the 2nd stage exhaust check valve opens and an isobaric exhaust occurs through a cooler. At 373 K, the final flow volume per minute equals 3.67 cu ft (103.8 L), or 4.86 cu in/cycle (0.0796 L/cycle) @ 1,304 rpm, internal energy change equals 18.5 kJ, W-in equals 8.1 kJ (7.6 Btu), and rejected heat equals 27.9 kJ. Per the ideal gas law, the H2 flow rate equals 17.07 moles/min or 0.0131 moles/cycle.

An isochoric H-in now takes place, increasing the temperature from 373 K to 466 K via the exhaust gas valved regenerator. Per the CGL calculator, pressure increases to about 94 psi. Internal energy change and H-in equal 33.7 kJ (31.9 Btu).

A second isochoric H-in now takes place, increasing the temperature to about 547 K (984 R) via the exothermic reactor heat exchanger. Per the CGL calculator, pressure increases to about 110 psi (758 kPa). Internal energy change and H-in equal 29.4 kJ (27.9 Btu).

Using the CGL calculator, an isobaric expansion from BDC of H2 at a pressure of 110 psi (758 kPa), and a temperature of 547 K (984 R) to a temperature of 646 K (1,162 R) indicates a flow volume of 4.43 cu ft/minute (125.4 L/minute) into the valved cell injector. W-out equals 14.6 kJ (13.8 Btu) and H-in equals 50.6 kJ (48 Btu).

An adiabatic expansion occurs from TDC as the valved cell expands/displaces the working fluid into the expander. Using the CGL calculator and assuming initial parameters of 4.43 cu ft (125 L), 640.5 K (1,152 R), and 110 psi (758 kPa), a final pressure of 36.5 psi (252 kPa) would equal a final volume of 9.71 cu ft (275 L), a final temperature of 466 K (840 R). Internal energy change and W-out would equal 63.8 kJ (60.5 Btu).

Net W-out can be determined by a sum of these processes as shown on FIG. 27:

• • A. 1st stage compression adiabatic compression W-in equals 20 kJ (19 Btu).
  • B. 1st stage compression isobaric exhaust W-in equals 8.2 kJ (7.7 Btu).
  • C. 2nd stage adiabatic compression W-in equals 19.7 kJ (18.7 Btu).
  • D. 2nd stage isobaric exhaust W-in equals 8.1 kJ (7.6 Btu). (Alternative is to increase adiabatic compression to ˜815 R (453 K). That corresponds to a pressure of about 92.2 psi (635.7 kPa). Volume would equal 105.7 L. Internal energy change and W-in would equal 29.2 kJ, which is close to (19.7+8.1=) 27.9 kJ (95.6%).)
  • E. Isochoric internal energy change and exhaust H-in equals 33.7 kJ (31.9 Btu)
  • F. Isochoric internal energy change and exothermic reactor H-in equals 29.4 kJ (27.9 Btu).
  • G. Expander isobaric expansion W-out equals 14.6 kJ (13.8 Btu): H-in equals 50.6 kJ (48 Btu)
  • H. Expander adiabatic expansion W-out equals 63.8 kJ (60.5 Btu).
  • I. Total W-in equals 56 kJ/min (48 Btu).
  • J. Total W-out equals 78.4 kJ/min (74.3 Btu/min).
  • K. Net W-out equals 22.4 kJ/min (21.2 Btu/min).

Examining the thermal elements of the proposed cycle, there are two isochoric (constant volume) H-inputs and one isobaric H-input. The first isochoric input is, of course, direct “waste heat” from the expander, and thus not included in calculating thermal efficiency. The second isochoric input is more complex, since it is heat received from a thermochemical catalytic reactor, the reactants of which are created in a separate thermochemical cycle. For the purposes of calculating the efficiency of the standalone process, it might be considered source heat. However, when looked at as the Benzene Battery concept's H2 storage process, exothermic reactor heat may also be considered otherwise-waste heat. It will therefore also be calculated both ways below.

Thermal efficiency can be determined by the ratio of net W-out minus H-in:

Total external source H-in (such as solar or geothermal) equals 50.6 kJ/min (48 Btu/min).

Total exothermic H-in equals 29.4 kJ/min (27.9 Btu).

Total external source plus exothermic H-in equals 80 kJ/min (75.8 Btu/min).

Total external source efficiency (exothermic heat as otherwise-waste heat) equals 44.3%

Total source heat plus exothermic source heat efficiency equals 28%.

For such a low temperature heat engine cycle, these are good results, even for a theoretical engine.

Note that, for the cycle above, 29.4 kJ/minute of output are required from the catalytically exothermic combination of C6H6+3H2 at a temperature of about 547 K (984 R). In U.S. Pat. No. 3,225,538, Table I, chemical heat of reaction changes for C6H12<=>C6H6+3H2 are given. In Table I, chemical heat change equals approximately 52.3 kilocalories per mol (219 kJ/mol, 207.5 Btu/mol) of C6H12 for both endothermic and exothermic reactions. The information given is for 1 atm (14.7 psi) constant pressure, but since heat is chemically stored, it would essentially be the same at any pressure or temperature driving the reaction. Thus, per minute, approximately 13.4% of a mol of C6H12 will need to be created. That will require 13.4% of a mol of C6H6 plus 40.2% of a mol of H2 per minute. At STP, H2 (gas) has a mass of 2.02 g/mol, so 0.402 moles will equal an H2 mass of about 0.812 g/minute.

Assuming the low heat of combustion, 1 g of H2 has a combustion value of ˜120 kJ. Electrolysis of H2O into H2+O is about 93% efficient. That means it requires ˜130 kJ to produce 1 gram of H2. For the H2-using process described above, electrolysis would require 105.6 kJ. Thus, the 24% efficient thermal process described above can ideally generate approximately 21% of the H2 required to power the cycle's exothermic heater, and the balance of the H2 required will have to come from some other source.

However, these results can be improved upon relatively easily by increasing the amount of source temperature added per cycle. One way to do so is to raise the engine's peak temperature, as for example by the additional H-in taking place at constant pressure (see the dotted constant pressure H-in lines in FIG. 27). Such an isobaric “superheating” can be accomplished by reworking the cycle thusly; increasing the temperature of the isobarically heated H2 working fluid to the higher temperature, increasing the expansion ratio by increasing the volume of the expander relative to the volume of the valved cell H2 injector, increasing the volume of the SD cylinder to match the new volume of the expander, increasing the volume of the 1st stage compressor to match the increased volume of the synchronizer cylinder, and adding a third stage to the compressor/intercooler.

In one possible prototype, it will be initially assumed that the three displacer cylinders will be comprised of a 2.5″ diameter cylinder containing a 2″ diameter connecting tube over a displacer piston stroke of 2.75″, or an effective piston area of 1.767 sq in and thus a cylinder volume of 4.86 cu in (0.0796 L), or 6,337 cu in/min (3.667 cu ft/min, 103.8 L/min) @ 1,304 rpm.

In this design, there is no requirement for a synchronizer, since the top displacer exhaust is isobaric. Instead, the expander exhaust is isobaric through the exhaust heat regenerator and directly into the 1st stage compressor. Since the 1st stage exhaust is likewise isobaric, the displace process can operate completely independent of the expansion/compression process. Thus, there is no need to synchronize the expander exhaust and the regenerator thermal charging processes with an SD cylinder. From above, the expander piston diameter equals 2.5″ and stroke equals 2.75″, thus piston area equals 4.91 sq in and cylinder volume equals 13.5 cu in (0.221 L). However, in this particular design, the expander has a drive rod on the top of the piston that pierces the expander cylinder head and connects to the 1st stage compressor piston, allowing the compressor piston to be “driven” by the expander piston.

Assuming the area of the connecting rod is adjusted for by increasing the area of the expander and compressor cylinders, the area of the expander/1st stage compressor pistons minus the connecting rod equals 13.5 cu in/cycle (0.221 L/cycle), or 288.5 L/min (10.19 cu ft/min.

An isochoric displacement takes place between the expander and the 1st stage compressor, with theoretically zero W-in or W-out. Per the CGL calculator, an isochoric heat exchange via a regenerator will theoretically reduce the temperature of the H2 working fluid to 373 K and will simultaneously reduce the pressure to 17.94 psi (123.7 kPa) (Point J).

An adiabatic compression then takes place. Per the CGL calculator, an adiabatic compression in the 1st stage compressor to 29.2 psi would require compression to about 11.3 cu ft (321 L) and raise the temperature to about 428 K. It would require W-in of about 20.9 kJ (19.8 Btu) (Point K).

The 1st stage adiabatic compressor now moves towards BDC, exhausting the gas at constant pressure. The isobaric inter-cooling to 373 K and a final volume of about 9.8 cu ft (278 L) would require about 8.3 kJ (7.9 Btu) of W-in (Point A). (A negligible amount of W-in is required at the end of this process, as was described earlier.)

Per the CGL calculator, in the 2nd stage, for an adiabatic compression from 29.2 psi (201 kPa), 373 K, and 9.71 cu ft (275 L), to 47 psi (324 kPa) and 428 K (770 R), final volume equals 6.92 cu ft, internal energy change and W-in equal 20.0 kJ (19.0 Btu) (Point B).

At 47 psi and part way through the upward stroke, the 2nd stage exhaust check valve opens and an isobaric exhaust occurs through a cooler. Per the CGL calculator, at 373 K, the final flow volume at 373 K into the third stage compressor equals 6.03 cu ft (170 L)/minute, internal energy change equals 20.1 kJ (19.0 Btu), W-in equals 8.2 kJ (7.7 Btu), and rejected heat equals 28.3 kJ (26.82 Btu) (Point C). (A negligible amount of W-in is required at the end of this process, as was described earlier.

The 3rd stage compressor now compresses the captured gas from 47 psi, 373 K, and 6.03 cu ft to 428 K (770 R) and 75 psi (510 kPa). Using the CGL calculator for an adiabatic compression, final volume equals 4.3 cu ft, and internal energy change and W-in equal 19.7 kJ (18.7 Btu) (Point D).

At 74 psi (510 kPa), the 3rd stage compressor exhaust check valve opens and an isobaric exhaust occurs through a cooler. Per the CGL calculator, at a final temperature of 373 K the final flow volume per minute into the lower displacer equals 3.67 cu ft (103.8 L), or 4.86 cu in/cycle (0.0796 L/cycle) @1,304 rpm), internal energy change equals 18.5 kJ, W-in (of exhaust) equals 8.1 kJ (7.6 Btu), and rejected heat equals 27.9 kJ. Per the ideal gas law, the H2 flow rate equals 17.07 moles/min or 0.0131 moles/cycle (Point E).

An isochoric H-in now takes place, increasing the temperature from 373 K to 466 K via the exhaust gas valved regenerator. Per the CGL calculator, H2 pressure increases to about 94 psi. Internal energy change and H-in equal 33.7 kJ (31.9 Btu) (Point F).

A second isochoric H-in now takes place, increasing the temperature to about 547 K (984 R) via the exothermic reactor heat exchanger. Per the CGL calculator, H2 pressure increases to about 110 psi (758 kPa). Internal energy change and H-in equal 29.4 kJ (27.9 Btu) (Point G).

Using the CGL calculator, an isobaric expansion into the valved cell displacer/injector at 110 psi to 5.1 cu ft/min (6.758 cu in/cycle @ 1,304 rpm) will require a temperature of about 737.4 K (1,327 R). Note that increasing the temperature at 110 psi to 4.43 cu ft/min required a H-in of 50.6 kJ (48 Btu) and produced 36 kJ of internal energy change and 14.6 kJ (13.8 Btu) of W-out. H-in to increase the temperature at 110 psi to 737.4 K would equal an additional 49.8 kJ (47.2 Btu), internal energy change would equal an additional 35.5 kJ (33.6 Btu), and W-out would equal an additional 14.4 kJ (13.6 Btu). Net H-in would thus equal 100.4 kJ/min (95.16 Btu/min), internal energy change would equal 71.5 kJ/min (67.77 Btu/min) and net W-out would equal 29 kJ/min (27.49 Btu/min) (Point L).

An adiabatic expansion from 110 psi to 22.4 psi (154 kPa) would increase volume to about 16.05 cu ft (454.5 L) (21.25 cu in @ 1,304 rpm) with a final temperature of 466 K. W-out would equal about 100.6 kJ (95.4 Btu) (Point M).

(Note: Clearly, that is a larger expansion than the existing prototype expander is capable of. Therefore, it would require either a secondary expander with an expansion ratio 1.635× larger or it would require reducing the size of the three displacer cylinders and the valved cell H2 displacer/injector to 61.2% of the previously calculated size).

Net W-Out can be Determined by a Sum of these Processes:

• • K. 1st stage compression adiabatic compression W-in equals 20.9 kJ (19.8 Btu).
  • L. 1^st stage compression isobaric exhaust W-in equals 8.3 kJ (7.9 Btu).
  • M. 2nd stage compression adiabatic compression W-in equals 20 kJ (19 Btu).
  • N. 2nd stage compression isobaric exhaust W-in equals 8.2 kJ (7.7 Btu).
  • O. 3rd stage adiabatic compression W-in equals 19.7 kJ (18.7 Btu).
  • P. 3rd stage isobaric exhaust W-in equals 8.1 kJ (7.6 Btu).
  • Q. Isochoric internal energy change and exhaust H-in equals 33.7 kJ (31.9 Btu)
  • R. Isochoric internal energy change and exothermic reactor H-in equals 29.4 kJ (27.9 Btu).
  • S. Total expander isobaric expansion W-out equals 29 kJ (26.4 Btu): Total isobaric H-in equals 100.4 kJ (95.2 Btu).
  • T. Expander adiabatic expansion W-out equals 100.6 kJ (95.4 Btu).
  • U. Total W-in equals 85.2 kJ/min (80.8 Btu/min).
  • V. Total W-out equals 129.6 kJ/min (122.8 Btu/min).
  • W. Net W-out equals 44.4 kJ/min (42.1 Btu/min).

Thermal efficiency can be determined by the ratio of net W-out minus H-in:

Total external source H-in (such as solar or geothermal) equals 100.4 kJ/min (95.2 Btu/min).

Total exothermic H-in equals 29.4 kJ/min (27.9 Btu/min).

Total external source plus exothermic H-in equals 129.8 kJ/min (123.0 Btu/min).

Total external source efficiency (exothermic heat as otherwise-waste heat) equals 44.2%.

Total source heat plus exothermic source heat efficiency equals 34.2%.

Note that in this approach, total ideal W-out can generate 42.0% of the H2 required to power the cycle's exothermic heater.

Note: Staged and intercooled compressors are used to approximate an isothermal compression, which is the ideal approach to compressing a gas or vapor. In the cycle above, it can be calculated that an isothermal compression from 17.94 psi (123.7 kPa), 288.5 L/min (10.19 cu ft/min), and 373 K (671 R), to 510 kPa (74 psi) and 66.67 L/min, W-in and heat out (H-out) will both equal 50.6 kJ/min.

Net W-out equals (129.6-66.67−) 62.93 kJ/min (59.65 Btu/min).

Total external source efficiency (exothermic heat as otherwise-waste heat) equals 62.7%.

Total source heat plus exothermic source heat efficiency equals 48.5%.

However, true isothermal compression is almost impossible to achieve in a real-world heat engine. In FIG. 27 and the calculations above, it is approached by a three stage inter-cooled isobaric-adiabatic compression process. Per the above, the requirement to use multi-staged and inter-cooled compressors created a reduction in total solar plus exothermic source heat efficiency thermal efficiency of about 30%.

A second way to increase the amount of source temperature added per cycle that doesn't require increasing the peak temperature of the engine is to use a reheat followed by a “super-expansion” within a secondary expander. One possible prototype of a secondary reheating process, in this instance assuming isobaric heating, is shown in FIG. 27 as a dotted line traveling from a primary expansion back to a peak temperature of 646 K (1,162 R). In the proposed prototype, the previous amount of adiabatic expansion is reduced, ending at about 60 psi (414 kPa), for example by reducing the diameter of the primary expansion cylinder (Point N). The H2 working fluid is then pumped at constant pressure from the primary expander through a secondary isobaric source heater, with the secondary heater raising the temperature once more to the peak temperature, in this case 646 K (1,162 R). The primary expander exhausts at constant pressure into a second valved cell injector (Point O), which injects the working fluid adiabatically (that is, at a constant rate during expansion) into the #2 expander, creating a second adiabatic expansion (Point P). Finally, the expanded gas is exhausted isochorically from the secondary expander through the regenerator/preheater and into the 1st stage compressor, dropping the temperature to about 373 K (671 R) (Not shown). A 1st stage adiabatic compression takes place, followed by an isobaric exhaust through a heat exchanger, reducing the final temperature back to 373 K (Not shown). A second and third compression-and-inter-cooled process may then be initiated, in this example at similar pressures and compressions to the process shown above. Thus, assuming similar processes as described in steps A thru H directly above:

An adiabatic expansion occurs from TDC as the valved cell expands/displaces the working fluid into the expander (see step J directly below). Initial parameters of 110 psi, 4.43 cu ft (125 L), 640.5 K (1,152 R) are assumed. Using the CGL calculator, a final pressure of 67.5 psi (462 kPa) would equal a final volume of 6.27 cu ft (178 L), a final temperature of 555 K (×R). Internal energy change and W-out would equal 30.78 kJ 29.6 Btu).

Using the CGL calculator, an isobaric expansion to a temperature of 640.5 K (1,152 R) indicates a flow volume of 7.22 cu ft/minute (206 L/minute) into the second valved cell injector. W-out equals 12.5 kJ (12.0 Btu) and H-in equals 43.3 kJ (41.6 Btu).

An adiabatic expansion occurs from TDC as the valved cell expands/displaces the working fluid into the expander. Using the CGL calculator, an adiabatic expansion to 22 psi (versus 22.4 psi (154 kPa) above) would increase volume to about 16.04 cu ft (versus 16.05 cu ft (454.5 L) above) with a final temperature of 464 K (843.5 R) versus 466 K above. W-out would equal about 64.7 kJ (60.9 Btu).

Net W-out can be determined by a sum of these processes:

• • A. 1st stage compression adiabatic compression W-in equals 20.9 kJ (19.8 Btu).
  • B. 1st stage compression isobaric exhaust W-in equals 8.3 kJ (7.9 Btu).
  • C. 2nd stage compression adiabatic compression W-in equals 20 kJ (19 Btu).
  • D. 2nd stage compression isobaric exhaust W-in equals 8.2 kJ (7.7 Btu).
  • E. 3rd stage adiabatic compression W-in equals 19.7 kJ (18.7 Btu).
  • F. 3rd stage isobaric exhaust W-in equals 8.1 kJ (7.6 Btu).
  • G. Isochoric internal energy change and exhaust H-in equals 33.7 kJ (31.9 Btu).
  • H. Isochoric internal energy change and exothermic reactor H-in equals 29.4 kJ (27.9 Btu).
  • I. Expander #1 isobaric expansion W-out equals 14.6 kJ (13.8 Btu): H-in equals 50.6 kJ (48 Btu).
  • J. Expander #1 adiabatic expansion W-out equals 30.78 kJ (29.6 Btu).
  • K. Expander #2 isobaric expansion W-out equals 12.5 kJ (12.0 Btu): H-in equals 43.9 kJ (41.6 Btu).
  • L. Expander #2 adiabatic expansion W-out equals 64.7 kJ (60.9 Btu).
  • M. Total expander isobaric expansion W-out equals 26.1 kJ (25.9 Btu): Total isobaric H-in equals 93.9 kJ (89.6 Btu).
  • N. Total expander adiabatic W-out equals 95.5 kJ (95.5 Btu).
  • O. Total W-in equals 85.2 kJ/min (80.8 Btu/min).
  • P. Total W-out equals 121.6 kJ/min (116.4 Btu/min).
  • Q. Net W-out equals 36.4 kJ/min (35.6 Btu/min).

Examining the thermal elements of the proposed cycle, there are two isochoric (constant volume) heat inputs and one isobaric heat input. The first isochoric input is, of course, direct “waste heat” from the expander, and thus not included in calculating thermal efficiency. The second isochoric input is more complex, since it is heat received from a thermochemical catalytic reactor, the reactants of which are created in a separate thermochemical cycle. For the purposes of calculating the efficiency of the standalone process, it might be considered source heat. However, when looked at as the Benzene Battery concept's H2 storage process, exothermic reactor heat may also be considered otherwise-waste heat. It will therefore also be calculated both ways below.

Thermal Efficiency can be Determined by the Ratio of Net W-Out Divided by H-in:

Total external source H-in (such as solar or geothermal) equals 93.9 kJ/min.

Total exothermic H-in equals 29.4 kJ/min.

Total external source plus exothermic H-in equals 123.3 kJ/min.

Total external source efficiency (exothermic heat as otherwise-waste heat) equals 38.8%

Total solar plus exothermic source heat efficiency equals 29.5%.

In this model, total ideal W-out can generate 34.4% of the H2 required to power the cycle's exothermic heater.

Isochoric Source Heating by Internal Combustion.

As noted above, an alternative to adding constant pressure or isobaric thermal energy is to add constant volume or isochoric thermal energy (see the solid non-vertical constant volume H-in lines in FIG. 27). Use of combustion means to add thermal energy to a heat engine with compressed air as its working fluid is well known. That includes relatively simple means for rapidly injecting a fluid, vapor, or gas into compressed air.

A valved cell can also be used to inject a gas or vapor. One advantage of using such an “injector valved cell” is that the pressure of the gas/vapor held within it only needs to equal the pressure of the gaseous and/or vaporous working fluid into which it is being injected when the cell's “transfer valve” is first opened. A process of “pressure-balancing” can then be used to aid in a rapid injection process may approach constant volume.

FIGS. 30a, 30b, and 30c illustrate one possible gaseous injector valved cell design that can intermittently connected via an exhaust or transfer valve to an expander cylinder. Note that if injection of the quantity of gas/vapor held within the valved cell is made to exactly match the rate of volume expansion in the expander, an isochoric process will occur.

A piston is situated within a valved cell cylinder and a seal separates the two ends of the cylinder.

The valved cell cylinder is connected on one end to an intake valve and an exhaust/transfer valve.

On the other end, a simple manifold connects to the expansion cylinder's expansion chamber, allowing the piston to match the pressure in the expansion cylinder.

A cylindrical rod is attached to the piston head on the intake and exhaust side and is aligned with the axis of the plunger. The rod passes through the gas injecting valved cell chamber, through a seal, and through a small cylindrical linear bearing and a sealing ring in the “head” of the valved cell. Consequently, when both sides of the plunger are subjected to an equal gaseous/vaporous pressure, the difference in the volume displaced by the rod will move the plunger in the direction of the gas-injecting chamber.

The cylindrical rod, having passed through the gas-injecting chamber, seal, and bearing, is connected to a low friction travel-limiting device, in this case a small freely rotating crankshaft. In the present design, the crank throw rotates inside a horizontally-sliding bearing which itself is captured within a guide frame of a vertically-sliding bearing that is itself captured within a immovable guide frame. The immovable guide frame forces the vertically sliding bearing to travel vertically, and since the vertically sliding bearing restricts movement of the horizontally-sliding bearing to horizontal movements, the rotation of the crank throw creates perfectly vertical movement of the cylindrical rod and thus of the valved cell piston.

The travel-limiting device ensures that the plunger will never physically contact the ends of the valved cell cylinder. It may even create an “accelerating injection” system as an aid in maintaining constant volume.

During the processes of injection and refilling, there will be times when the pressure on the gas-injecting side of the plunger will be higher than the pressure on the expansion chamber manifold side of the plunger. Accordingly, the plunger is constructed to be strongest on the sealing ring side, and constructed to be as light as possible on the expansion chamber manifold side.

The piston is also constructed with sufficient length along its axis to keep the piston sealing ring from running on the piston wall exposed intermittently on the expansion chamber manifold side of the cylinder. If necessary, the portion of the valved cell injector cylinder wall in which the piston sealing ring sits is cooled.

The gas intake valve may be a simple poppet-type check valve that seals against higher pressure on the injection chamber side. It is biased to return to closed, for example by a return spring.

The exhaust or transfer valve is constructed similarly to the CCVC prototype transfer valve. It is opened primarily by pressure equalization across the valve head at the end of the expander exhaust stroke. The pressure equalization is created, as in the CCVC prototype, by a slightly early closure of the expander exhaust valve that captures a small amount of remnant gas in the limited space between the expansion piston and the expansion head, thus allowing pressurization of said remnant gas to equalize pressure across the transfer valve. When pressure is thus equalized, some means, such as a spring bias, may be used to easily and quickly open the transfer valve, thus connecting the expander cylinder to the previously-charged gas-injecting chamber.

Upon movement of the transfer valve towards open, the force differential across the valved cell piston (created by the cylindrical rod) will begin to inject the gas into the expander combustion chamber.

If there is a single gaseous fuel injector such as H2 injecting into an oxidizing environment, the fuel and oxidizer instantly begin to mix and may be instantly combusted, instantly driving up the pressure in the combustion chamber, the valved cell injector chamber, and the manifold connecting to the other side of the valved cell piston. That in turn instantly drives the valved cell piston to inject all the contents of the injector chamber into the combustion chamber.

If there are two gaseous injectors, as for example a gaseous O2 injector and a gaseous H2 injector, then both injector exhaust or transfer valves open simultaneously by the same process. As a result, O2 and H2 instantly begin to mix and instantly combust, instantly driving up the pressure in the combustion chamber, the H2 valved cell chamber, the O2 valved cell chamber, the expander cylinder, and the manifold connecting to the O2 valved cell plunger. That in turn instantly drives the O2 valved cell plunger to inject all the O2.

Referring to the above, FIG. 27, and the CGL calculator, at 74 psi and 373 K (Point E), the final flow volume per minute equals 3.67 cu ft (103.8 L), or 4.86 cu in/cycle (0.0796 L/cycle) @ 1,304 rpm, internal energy change equals 18.5 kJ, W-in equals 8.1 kJ (7.6 Btu), and rejected heat equals 27.9 kJ. Per the ideal gas law, the H2 flow rate equals 17.07 moles/min or 0.0131 moles/cycle.

Assuming an isochoric thermal input to 555.6K, final pressure would equal 110.2 psi, and recycled H-in would equal 63.82 kJ (Point G).

Assuming an isochoric thermal input to the peak temperature of 950 K (1,710 R, 677 deg C., 1,250 deg F.) (Point Q) shown in FIG. 27, an H2 flow rate of 17.07 moles/min or 0.0131 moles/cycle, and a volume of 3.67 cu ft/min (103.8 L/min), pressure equals 1300 kPa (188.5 psi), and source H-in would equal 138 kJ. An adiabatic expansion to 555 K and 13.8 cu ft (390.6 L) would equal about 29.3 psi (202 kPa). Total change in internal energy change and W-out equals 138 kJ/min.

The instant the gas/vapor injector valved cell connected to the expander are emptied, an adiabatic expansion then occurs within the main expander. Note that, at expander BDC, continued expansion may follow into a lower pressure, lower temperature, uncooled and non-lubricated secondary expander.

When the pressure in the gas/vapor injector valved cell is equal to the feed pressure of the gas/vapor sources (in this case, at 110 psi, occurring when the adiabatic expansion drops to about 1,580 R or 877.7 K), the gas/vapor injector valved cell exhaust or transfer valve is closed. As soon as said injector exhaust valve is closed, the gas/vapor intake valve is opened.

When pressure in the expander drops below the feed pressure of the gas/vapor sources, pressure is reduced in the pressure-equalizing displacer system, and the gas/vapor injector valved cell cylinders are automatically refilled.

When the gas/vapor injector valved cells are fully charged, the gas/vapor injector valved cell intake valve is closed, completing the cycle and preparing for the next cycle.

If the gas/vapor is derived from a liquid, excess heat from the heat engine can be used to preheat the gas/vapor to a high pressure with very little W-in. Note that there is then no requirement for a gas compressor.

In the instance that an H2+O2 combustion process occurred, following adiabatic expansion and exhaust through the preheater, the working fluid may be cooled sufficiently for H2O to be easily separated from the non-combusted working fluid gas/vapor. The remnant working fluid may then be recycled through the compression system, as described above

Continuing on, the exhaust from the expander (at about 1,000 R (555 K) in this instance) can either be isobaric or isochoric.

For an isobaric exhaust, the usual approach (and the approach used in the original CCVC prototype) would be to “capture” thermal energy by means of a counterflow heat exchanger. As a result:

• • (1) The two counter flowing streams of fluid are required to each have their own “containers”, and heat is only able to transfer by conduction through the walls of those containers (usually made up of many small tubes in direct physical contact with one another). This results in relatively poor heat transfer over time.
  • (2) In order to give the heat transfer process more time to take place, the tubes are generally quite long, creating a large amount of internal volume, thus resulting in low changes in temperature and pressure over a short distance, should that be required by one or the other fluid streams.
  • (3) Because of the length of the heat exchanger and, in a heat engine, the requirement for the receiving fluid to be at a much higher pressure and thus the receiving fluid container to be stronger than the thermal charging fluid, a great deal of mass must be heated/cooled.
  • (4) Since the exhaust process is isobaric, the work of exhaust is relatively high.
  • (5) Assuming an isochoric heat absorption process, an isobaric exhaust will theoretically contain more thermal energy than the isochoric process can use, and thus may represent waste energy.

For an isochoric regeneration, the advantages are:

• • (1) Since the gases pass through a “thermal sponge”, the internal masses required of the heat exchanger are greatly reduced, since heat is given off in one flow and taken in in the opposite flow. This results in a highly efficient heat transfer process.
  • (2) The internal volumes are greatly reduced, resulting in much higher changes in temperature and pressure for a distance traveled by the fluids.
  • (3) No W-in is required in exhausting isochorically from the exhauster displacer volume, and, since thermal energy is removed, the receiver displacer volume's exhaust will be at a much lower pressure and temperature, thus requiring less work overall.
  • (4) Assuming a match in mass displaced in both directions, there is exactly as much thermal energy charging the regenerator (at a lower pressure) as is required by an isochorically-displaced (higher pressure) gas that will be absorbing the thermal energy.

An isochoric expansion to about 474 L/min (16.73 cu ft/min), 373.5 K (670 R), and 112.4 kPa (16.3 psi) would generate 64.0 kJ of heat. In an ideal cycle, an isothermal compression to 373.5 K (670 R) and 74 psi (510 kPa) would reduce volume to 104.4 L/min (3.68 cu ft/min). W-in and H-out would equal 80.5 kJ. Since exactly as much thermal energy would be available in the exhaust at constant volume as in the regeneration into the compressed working fluid at constant volume, then it is easier to calculate the potential ideal thermal efficiency if the exhaust through the regenerator were at constant volume.

An isochoric exhaust process proceeding from 24.1 psi (166.2 kPa) at a temperature of 555 K (1,000 R) and a volume of 474 L/min (16.73 cu ft/min) to a pressure of 112.4 kPa (16.3 psi) and a temperature of 373 K (670 R) would have an internal energy change and a heat rejection of 64.0 kJ

For an isothermal compression, an isothermal compression from 474 L/min (16.73 cu ft/min), a pressure of 112.4 kPa (16.3 psi), and a temperature of 373 K (670 R) to a final pressure of 510 kPa (74 psi) and a final volume of 104.4 L/min (3.69 cu ft/min) would equal H-out and W-in equal to 80.5 kJ. Total work generated equals 57.3 kJ/min. Thermal efficiency would thus equal total W-out divided by total H-in or 41.5%.

However, as stated above, a three stage inter-cooled isobaric-adiabatic compression process will require about 85 kJ/min, reducing overall W-out to 53 kJ/min. Thus, the theoretical thermal efficiency of the above process would equal about 38%.

Since the peak temperature is 950 K and the sink temperature equals 373 K, theoretical thermal efficiency equals (T1−T2)/T1 or 60.7%, the theoretical thermal efficiency equals 62.5% of Carnot.

Isochoric Source Heating+Exothermic Preheating

It is quite possible to use multiple isochoric regeneration as is described above to replace some high grade source heat with medium grade source heat. In FIG. 22, at 1 atm, for a 90% conversion of H2+C6H6 into C6H12, thermal output would equal about 650 K (1,170 R). Note that, in U.S. Pat. No. 3,225,538, Table I, chemical heat change equals approximately 52.3 kilocalories per mol (219 kJ/mol, 207.5 Btu/mol) of C6H12 for both endothermic and exothermic reactions. At a 95% conversion rate, that would equal 197 kJ per mol. Thus, per minute, approximately 16.8% of a mol of C6H12 will need to be created. That will require 16.8% of a mol of C6H6 plus 50% of a mol of H2 per minute. At STP, H2 (gas) has a mass of 2.02 g/mol, so 0.5 moles will equal an H2 mass of about 1 g/minute of H2.

Per the CGL calculator, to isochorically raise the temperature of the mix from 555 K to 650 K would increase the pressure to 888.8 kPa (128.9 psi) and require 33 kJ/min, decreasing the source heat required to 105 kJ. That in turn increases the thermal efficiency to 50.4%, and increases the percentage of Carnot to 83%.

Finally, since W-out equals 53 kJ and electrolysis is 93% efficient, this model can produce 49 kJ worth of H2 per minute. Since, from above, it requires 130 kJ to produce 1 gram of H2, this model can produce about 0.38 grams of H2/minute, in this model, total ideal W-out can generate 38% of the H2 required to drive the exothermic reaction.

Isochoric Source Heating by Internal Combustion in Compressed Air.

One possible prototype would involve injecting pressurized H2 into compressed and preheated air in an “open” cycle process (see one proposed gaseous injector design above under the “Isochoric source heating by internal combustion” heading). Assuming the “used” air is exhausted following the heat content being removed to preheat a new charge of compressed air, then a fuel, such as pressurized H2, can be injected with essentially the same potential efficiencies determined above.

Assumptions:

Use of compressed air as the working fluid at 74 psi (510 kPa) and 670 R (372 K); i.e., open cycle isochoric combustion.

Use of compressed H2 as the fuel.

Use of the existing prototype expansion ratio of 1:2.777; 4.86 cu in/cycle (0.0796 L/cycle) displaced into 13.5 cu in/cycle (0.221 L/cycle); 3.67 cu ft/min (103.8 L/min) displaced into 10.18 cu ft/min (288.2 L/min).

The upper displacer cylinder is used as the valved cell.

Two intercooled compressions totaling approximately 50 Btu/min; an adiabatic expansion totaling approximately 75 Btu/min; total theoretical W-out totaling approximately 25 Btu/min (26.4 kJ) or about 0.6 HP/hr (0.44 kWh).

Peak temperature of about 1,180 R (655.6 K); expander exhaust temperature of about 840 R (466.7 K); synchronizer displacer exhaust temperature of about 670 R (372 K).

Peak pressure would approach 120 psi (827 kPa); expander exhaust pressure would equal about 30 psi; synchronizer displacer exhaust pressure would equal about 25 psi (note that a small turbocharger could use this exhaust energy to “boost” the input stream to the 1st stage compressor), improving thermal efficiency).

H-in equal to about 75 Btu/min or 0.0575 Btu/cycle (60.7 J/cycle); assuming the low heat of combustion of 1 g of H2 or 120 kJ, the mass of the injected H2 would equal approximately half a mg/cycle (30.3 mg/minute, 1.8 g/hour, at STP, H2 (gas) has a mass of 2.02 g/mol, and the molar mass of H2 injected per cycle would equal 0.00025 moles; Pressure at the end of isochoric waste heat regeneration would equal about 80 psi (551.6 kPa) and temperature would equal about 840 R (466.7 K); per the ideal gas calculator, at 0.00025 moles, 466.7 K, and 551.6 kPa, injector volume per cycle would equal 0.00176 L (0.107 cu in). Assuming a 0.75″ diameter injector cylinder, stroke would equal about 0.25″.

Overall theoretical thermal efficiency (assuming no turbocharger) equals 25 Btu W-out divided by 75 Btu H-in or 33.3% (about equal to the maximum efficiency of a typical gasoline-burning engine). Ideal Carnot efficiency or T1/T2)/T1 equals 43.2%. Percentage of ideal Carnot efficiency thus equals 77.1%.

Isochoric Heating with an Isobaric Heat Source

An isobaric valved regenerator or STREP is proposed in FIGS. 1 through 4. Note that, for example, the thermal output of an exothermic reactor can be easily and efficiently absorbed by a coolant such as H2 recirculated at constant pressure by an isobaric STREP.

As mentioned above, a different kind of STREP was proposed in FIGS. 11 through 17, where the SD cylinder received the output from the expander and the “receiver cylinder” served double-duty as the 1st stage compressor. Unlike the STREP in FIGS. 1 through 4, which exchange thermal energy isobarically or at constant pressure, the STREP in FIGS. 11 through 17 exchanged thermal energy isochorically or at constant volume.

It is likewise perfectly feasible to thermally charge a STREP with an isobaric gas flow, then “switch” the regenerator to isochorically remove some or all of the thermal charge thus deposited. FIGS. 32 and 42 illustrate what such a mixed isobaric/isochoric STREP might look like.

In the relatively simple cycle shown in FIG. 33, both solar energy and exothermic reactor energy are shown being absorbed by constant pressure “carrier fluids”, capturing thermal energy at the peak temperature of both processes. In both examples, a mixed isobaric/isochoric STREP is used to then efficiently transfer much of that thermal energy into isochoric processes that are thus “driven” to the peak temperatures of the heat sources. A valved cell displacer/injector and adiabatic expander then generates work, and the exhaust energy is finally captured as well in an isochoric STREP. The result, as shown in FIG. 33, is to create a series of four isochoric thermal inputs: First, an isochoric capture is followed by an isochoric regeneration of waste exhaust heat from the engine by using a STREP similar to the one shown in FIGS. 11 through 17; second, an isobaric capture and an isochoric regeneration of exothermic reactor heat; and finally, an isobaric capture and an isochoric regeneration of high temperature source energy such as solar energy.

FIG. 42 illustrates a process whereby a constant-pressurize gas, having absorbed thermal energy from some process, can, by “matching” pressures during a charging cycle with pressures on the isochoric side of a regenerator, in turn be used as the carrier of said externally applied thermal energy, passing it cyclically through a valved regenerator.

Note that FIG. 42 shows a much larger cylinder on the lower left than on the lower right, and a much smaller cylinder on the upper left. If the larger lower cylinder is at high temperature and low pressure, then its volume will be reduced as it deposits thermal energy into the regenerator, reducing the volume being displaced into the upper left cylinder. FIG. 42 is thus illustrating the impact on an isobaric displacement of removing thermal energy over the course of a given stroke distance. In this case, the higher pressure thermally receiving fluid will be in the cylinder on the lower right, which will receive fluid at constant pressure as the pistons move downward.

In another use of the STREP in FIG. 42, a reactant mix, for example composed of one mol of C6H6 and three moles of H2, is passed through the regenerator at constant pressure and temperature, delivering thermal energy to a regenerator that is also a catalytic reaction chamber. As a result, the number of moles of product is potentially reduced to a quarter of the moles within the reactant thus created. And since neither pressure nor temperature changed, volume is correspondingly reduced to one quarter as well.

As stated earlier, a high temperature but low pressure exhaust fluid at constant pressure can be passed through a counterflow regenerator-type heat exchanger, thermally “charging” the regenerator. The regenerator can then be raised in pressure, in this case by a semi-adiabatic compression of remnant product in the large cylinder by an early closure of the regenerator's exhaust valve. In this instance, a separate stream of counter-flowing fluid at the higher pressure but at low temperature can then enter the regenerator through an intake valve and flow isobarically through the regenerator into the lower right cylinder, thus isobarically absorb the thermal energy deposited by the earlier low pressure flow. Note that it could also be a higher pressure isochoric absorption process. Finally, the high pressure in the regenerator can be reduced to that of the low pressure stream and once more be used to “charge” the regenerator, in this case by early closure of the upper left small cylinder exhaust valve causing a semi-adiabatic compression of remnant product there, followed by a re-expansion of remnant high pressure fluid in the regenerator back into the low pressure displacer at the beginning of its intake stroke, followed by low pressure fluid flow out of the large lower cylinder into into the small upper cylinder.

In other words, it is perfectly feasible to thermally charge a valved regenerator with a gas at isochorically, then “switch” the regenerator to isobarically add or remove some or all of the thermal charge thus deposited, or vice versa.

Work-In Requirements of a C6H6+3H2 Exothermic Heat Generator.

An important requirement is that C6H6+3H2, also called an “exothermic fluid”, be made available in a state that is capable of generating the required temperature for conversion into C6H12, also called an “endothermic fluid”, in this case exothermic heat produced at a temperature 547 K (984 R). Per FIG. 22, that would require a pressure of approximately 2 atmospheres (30 psi, 206 kPa).

When a heat engine both produces W-out and useful thermal energy, it is termed a CHP or Combined Heat and Power process. In U.S. patent application Ser. Nos. 17/746,848 and 18/095,463, it is proposed that conversion of C6H6+3H2 into C6H12 can be done in conjunction with or even within the confines of heat engines that generate useful thermal energy, W-out, or a combination of both useful thermal energy and W-out. For generating W-out only, the process is termed a Bland/Ewing Combined Cycle or B/E-CC process. For generating both W-out and useful thermal energy, the process is termed a Bland/Ewing Combined Heat and Power or B/E-CHP process.

In the B/E-CHP process, when some smaller portion of a larger amount of exothermic fluid needs to be converted to endothermic fluid, the heat thus generated can drive an engine used to supply power for the process of converting the whole of the endothermic fluid. In U.S. patent application Ser. No. 18/095,463, FIG. 30 and FIG. 35, an exothermic “production” B/E-CHP Cycle is also proposed. It is estimated to generate a small amount of excess W-out while also converting on the order of 75% of the exothermic fluid into useful thermal energy. Note that this cycle was not “optimized” to take advantage of the STREP heat exchange process.

One proposed mechanism for optimizing a B/E-CHP process through use of a modified STREP would involve:

• • 1. Physically connecting two equal volume displacers with opposing cyclical intake and exhaust strokes via a regenerator;
  • 2. adding an intake and an exhaust valve to each displacer, where the exhaust of the lower temperature displacer connects to the cold side of the regenerator and the intake of the higher temperature displacer connects to the hot side of the regenerator;
  • 3. Adding an additional intake valve and exhaust valve to the regenerator itself, where the regenerator intake valve connects to the hot side of the regenerator and the regenerator exhaust valve connects to the cold side of the regenerator;
  • 4. The regenerator intake valve would be connected to the adiabatic high temperature, high pressure side of a counterflow recuperator, which would receive thermal input from some heat source, for example the high temperature, low pressure exhaust from a heat engine;
  • 5. The regenerator exhaust valve would be connected to the adiabatic low temperature, high-pressure side of said counter-flow recuperator, which would connect to the low temperature side of said counterflow recuperator.
  • 6. A double-acting intermittent and cyclical pump located near to the regenerator exhaust valve receives cold constant pressure working fluid from the cold side of the regenerator and pumps it into the system entering the cold side of the counterflow recuperator. This cyclical pump is timed to intake fluid into one side of the double-acting piston and simultaneously exhaust fluid from the opposite side when the regenerator intake and exhaust valves are open. On the following cyclical opening of the regenerator intake and exhaust valves, the opposite side takes in a “charge” of cold fluid while exhausting the previous “charge” with the double-acting piston. This double-acting intermittent pump mechanism can, for example, be “timed” to operate through use of a “geneva mechanism”, in a manner that is common practice.

The cyclical process would take place in the following manner:

• • a. As the higher temperature displacer reaches its maximum volume and the lower temperature displacer reaches its minimum volume, the displaced contents will reach its maximum pressure and temperature throughout the connected displacers and the regenerator, and the higher temperature displacer intake valve is closed.
  • b. The higher temperature displacer exhaust valve, which may be the expander transfer valve, is instantaneously opened, allowing the fluid in the higher temperature displacer to exhaust. (Note: The exhaust may be to a second isochoric temperature input system, or may be the exhaust to an isobaric temperature input system. However, for this example, it will be assumed to exhaust directly into an expander.) Simultaneously, the lower temperature displacer begins to move away from minimum volume, and the lower temperature displacer exhaust valve is held open by some force, dropping the pressure in the regenerator and the lower temperature displacer.
  • c. The pressure in the regenerator and lower temperature displacer drops to approximately the pressure in the counter-flow recuperator system. Simultaneously, the exhaust valve for the lower temperature displacer instantaneously closes and the regenerator intake and exhaust valves open.
  • d. A measured quantity of the adiabatic fluid proceeding from the recuperator at high pressure and high temperature is then “pumped” through the regenerator, thermally charging it.
  • e. As the higher temperature displacer approaches its minimum volume and the lower temperature displacer approaches its maximum volume, the higher temperature displacer and expander contents will reach its minimum pressure and temperature throughout the connecting manifold, and the higher temperature displacer exhaust valve is closed.
  • f. Continued travel of the higher temperature displacer will now raise the pressure of any remnant fluid until it reaches approximately the pressure in the regenerator. As the higher temperature displacer reaches its minimum volume and the lower temperature displacer reaches its maximum volume, at which point the higher temperature displacer intake valve will open. Simultaneously, the regenerator intake and exhaust valves are closed, and the lower temperature displacer intake valve is opened. (Note: The exhaust may be to a second isochoric temperature input system, or may be to an isobaric temperature input system. In that case, the higher temperature displacer intake valve is not opened until some re-expansion of remnant fluid within the higher temperature displacer occurs, equalizing pressure between it and the regenerator, at which time the higher temperature displacer intake valve will be opened.)
  • g. A purely isochoric displacement occurs. The isochoric displacement through the regenerator then raises both the temperature and the pressure of the displaced working fluid until the upper displacer has completed its expansion and the lower displacer has completed its exhaust, and the cycle begins again.

A C6H12 Dissociation-Pressurized H2 Gas Generator (See FIG. 38 and FIG. 41)

For 0.4536 kg C6H12 converted 100%, the yield is 0.4210 kg of C6H6 and 0.0326 kg of H2.

The vapor molar heat capacity of C6H12 is 105 J/(mol K), or a vapor molar heat capacity of 1.25 kJ/kg/(K).

The vapor molar heat capacity of C6H6 is 82.4 J/(mol K), or a vapor molar heat capacity of 1.05 kJ/kg/(K).

The molar heat capacity of H2 is 28.84 J/(mol K) (6.89 cal, 0.0273 Btu), or 14.27 J/gram/(K), or 14.28 kJ/kg/(K). (For 1 lb (0.454 kg), molar heat capacity equals 6.48 kJ/(K).)

The total molar heat capacity of one mol K of C6H6 plus 3 moles of H2 equals 168.92 J/(K).

C6H12 boils at 1 atm and 353.9 K (637.0° R). C6H12 has a standard heat of vaporization requirement of 32 kJ/mol/(K), or 380 kJ/kg.

C6H6 boils at 1 atm and 353.2 K (635.8° R). C6H6 has a standard heat of vaporization requirement of 33.9 kJ/mol/(K), or 433 kJ/kg.

Per FIG. 22, at 950 K (1,710 R, 677 deg C., 1,250 deg F.), an endothermic reaction would require a pressure of about 5.25 atmospheres (77.2 psi, 532 kPa). It will be assumed that 1 mol of C6H12 (the product) will be converted per stroke to 1 mol of C6H6 plus 3 mols of H2 (the reactant) with a total molar heat capacity of 168.92 J/degree K. The total molar heat content of 1 mol of the reactant will thus equal 160.474 kJ.

As noted above, U.S. patent application Ser. No. 18/095,463, use of an Exothermic Reactor Exhaust Compressor (EREC) is proposed to assist in the vaporization of C6H6 by a counter flowing exchange of heat with condensing higher pressure C6H12. An Endothermic Reactor Exhaust Compressor (ENREC) may likewise be used to permit the condensation of 1 mol of higher pressure C6H6 to supply all of the thermal energy required to vaporize 1 mol of C6H12. Note that C6H6 and C6H12 boil at approximately the same temperature and have approximately the same standard heat of vaporization requirement. Since the total molar heat capacity of C6H12 at 950 K equals 99.750 kJ, only 62.2% of the total molar heat content of the reactant is required to preheat C6H12 from a temperature just above vaporization, estimated at 423 K, to the temperature required for the endothermic reaction at 950 K and 532 kPa. The difference in temperature being 418 K, total heat available would equal 70,608 kJ, heat transferred would equal 43.919 kJ, and remaining heat would equal 26.689 kJ.

It is therefore possible to separate the exothermic fluid exiting the reactor into a 62% stream and a 38% streams, the 62% stream being used to vaporize the endothermic fluid. The reactant mix will now be exhausted through two different heat exchangers. In one possible use case, the smaller fraction will pass through heat exchanger #2 and preheat H2 returning from the gas/liquid separator, as will be shown. The 62% stream will enter the main endothermic reactor preheater, preheating the vaporous endothermic fluid to the endothermic reactor temperature and simultaneously cooling the reactant mix to just above C6H6 condensation temperature. At which time the two reactant mix streams will be recombined, as will be shown. Note that this may be done at constant pressure or at constant volume.

The combined streams now enter the exothermic mix condenser/endothermic mix vaporizor/cooler, which is located between the liquid C6H12 pump and the (endothermic) ENREC compressor. The combined flow of higher pressure reactant will supply the thermal energy to preheat and vaporize the lower pressure liquid C6H12, after which the C6H12 vapor will be compressed by the ENREC to about the pressure of the reactant, in this case to 5.25 atm. Finally, as flow continues, the reactant will enter the cooler and be cooled completely, thus separating the liquid C6H6 and remnant C6H12 from the H2 gas. The C6H6 and C6H12 can then be further separated, as by use of a centrifuge.

Note, however, that there is a huge volume difference between the reactant and the product at any given temperature and pressure, since what was a single mol is now 4 moles. This is somewhat analogous to the high temperature low pressure exhaust gas from a combustion engine being used to preheat low temperature high pressure working fluid for the engine where, unlike is shown in FIGS. 1 through 4, a STREP exchanging thermal energy between those two adiabatic fluid flows would need to have a much larger low pressure cylinder and a much smaller high pressure cylinder, similar to that shown in FIG. 34. Of course, in the case of the vaporous endothermic fluid and the vaporous exothermic fluid, the pressures would be similar, but the volumes would still be hugely different, especially so considering the possession of both temperature and molecular differences.

The pure H2 gas at 5.24 atm is now free to be used. There are several possibilities:

As suggested in U.S. patent application Ser. No. 18/197,092, FIG. 27, the separated 5.25 atm H2 can be passed back through a H2/reactant isobaric recuperator-style heat exchanger/preheater. Alternatively, the H2 can be passed back through an H2/reactant isobaric STREP preheater, either a constant pressure version as described in FIGS. 1 through 4, a constant volume version as used in FIGS. 11 through 17, or a mixed version as described herein. See FIGS. 35 through 41 illustrating one possible Bland/Ewing composite cycle to which the STREP heat exchange process can be applied.

FIG. 35 shows all the paths in FIGS. 36, 37, and 38 as a tracing over FIG. 18. FIGS. 36 through 38 show the working fluid pathways based on FIG. 18 for a work producing cycle and a refrigeration cycle. FIG. 36 shows the paths for a low pressure and a higher pressure exothermic half-cycle that produces endothermic fluid (in this instance C6H12). FIG. 37 shows the H2 compression path for both an exothermic and endothermic work-producing cycle and the H2 path for the exothermic work-producing cycle. FIG. 38 shows the endothermic half-cycle path for a refrigeration cycle.

FIGS. 36 through 38 are labelled. The labels can generally be applied to FIGS. 39, 40, and 41, as shown. FIG. 39 is copied from FIG. 25 and shows a simple exothermic heat generation system, said produced heat which can then help power or completely power an exothermic half-cycle engine or otherwise produce useful heat.

FIG. 40 is copied from U.S. patent application Ser. No. 18/197,092, FIG. 2, and shows a higher pressure work-producing cycle such as is shown in FIG. 37. FIG. 41 shows FIG. 40 reconfigured as a refrigeration cycle such as is shown in FIG. 38. (Note: FIG. 40 indicates an EREC compressor which has been relabeled an ENREC compressor. To avoid confusion, FIG. 41 also indicates an EREC compressor that is actually an ENREC compressor.)

FIG. 36, in addition to show a paths for a low pressure and a path for a higher pressure exothermic half-cycle that produces endothermic fluid, is also comparing an alternative approach to using an EREC to raise pressure following vaporization. The thin dotted line between A and D or A′ and D′ represents the adiabatic compression of an EREC to the higher pressure at which the catalytic reaction (exothermic in this instance) will take place. The line from D to B or D′ to B′ represents an isobaric heating to the temperature of the reactor, while the line from B to C or B′ to C′ represents an isobaric regeneration of the heat from the exhausting product into the cold reactant.

However, it is also possible to raise the pressure of the reactant by using an isochoric rather than an isobaric process such as via a STREP heat exchange process, and use the exhausting product to supply that thermal energy, although it would not be as thermally efficient. Note that the product would still take the line from B to C or from B′ to C′, since it is desirable that the product be at the higher pressure in order to supply the heat of condensation of the liquid constituent of the product to accomplish the vaporization of the liquid or solid reactant.

In essence, an isochoric STREP process can be seen as accomplishing a kind of “thermal isochoric compression” as opposed to an adiabatic/isentropic compression.

FIG. 36 further shows a thin rising curved line between C and D′. This thin line represents the impact of rising pressure on the point at which a liquid converts to a vapor. As shown, at about 5.25 atm, the temperature required to, in this instance, vaporize C6H12, has increased to about 750 R (416 K, 290 deg F., 144 deg C.). Note that, at that pressure, using an isochoric STREP process can essentially only function above that temperature. However, recall that pressure is supplied to the liquid or solid reactant, avoiding the requirement to compress a vapor. While that might seem to “save” considerable W-in, an isobaric system also takes W-out as thermal energy is added and returns that work as thermal energy is removed, so the net efficiency gain favors the isobaric STREP process.

Recall that ideally mol count would equal 3 mols of H2, for a total molar heat capacity of 86.52 J/degree K. Recall that the remaining heat available equals 26.7 kJ. Therefore, if passed through a purely isobaric reheater, assuming a 100% efficient heat transfer, the 3 mols of H2 at 418 K can be raised 308.5 K, to 726.5 K.

Alternatively, the 5.25 atm H2 can be passed back through an H2/reactant isochoric STREP or mixed isobaric/isochoric STREP (FIG. 32). Per the ideal gas law calculator, 1 mol of H2 at 423K and 532 kPa equals 6.61 L. For 3 mols, volume equals 19.8 L. Per the CGL calculator, for an isochoric thermal input of 26.7 kJ, the final pressure would equal 1,079 kPa and the final temperature would equal 858 K. Thus pressure would be increased by 547 kPa, or essentially double the pressure of the isobaric STREP or the typical isobaric recuperator, and temperature would be increased by 131.5 K to nearly the temperature of endothermic dissociation.

Note that the volume of the H2 cylinder would be much closer to the volume of the C6H6+H2 exothermic fluid mix cylinder at a similar temperature and pressure. No calculations have been attempted on this possible approach.

A second interesting alternative for using the H2 is as a refrigerant. Having cooled the C6H12 and H2 below C6H12 condensation temperature (estimated at 670 R (372 K, 99 deg C., 210 deg F.) and thus separated out the H2, the 5.25 atm and 3 moles of H2 can then be further chilled to ambient temperature and expanded to generate cold. Per the ideal gas calculator, volume prior to expansion would equals 17.44 L. Per the CGL calculator, expanding 56.77 L of H2 at 372 K to from 5.25 atm to 1 atm would decrease the temperature of the H2 to 230.7 K (−42.4 deg C., −44.4 deg F.) and would generate 8.684 kJ or work. Assuming a final isobaric exhaust at 1 atm, no exhaust W-in or W-out is required. Note that the 26.7 kJ of thermal energy at 950 K is still available.

The cooled H2 gas at 5.25 atm can be stored for later use, and any excess latent heat can be used for CHP and/or CC purposes.

The cooled and expanded H2 gas can be fed to a low pressure fuel cell, generating electricity at very high efficiency.

The cooled gas can be injected with no additional compression required into the cooled H2 working fluid following the final compression of an SD CVCC H2+O2 combustion engine. That is, it and can be used as “makeup H2” to replace the combusted H2. Note that a gas compressor, preferably multi-staged and inter-cooled, is still required for the majority of cycling working fluid. Recall that, in order to keep peak combustion temperature down to a sustainable level, there must be a large quantity of non-combusted working fluid relative to the combusted working fluid.

If H2 is used as the working fluid, a simple injection of pre-pressurized O2 into pressurized and preheated H2, for example by cyclically injecting the O2 via a “displacer” valved cell (see proposed gaseous injector, FIGS. 30a, 30b, and 30c), the injected O2 can supply the required heat of combustion to drive the cycle. Note that the original pressure of the gas taken into such a displacer valved cell only needs to equal the pressure of the gaseous or gaseous and/or vaporous working fluid in the H2 valved cell prior to combustion. Finally, note that the H2O thus produced in the resulting working fluid mix can easily be condensed and removed at the end of the cycle. In fact, the removal of the H2 captured in the newly-produced H2O is why “makeup H2” is required.

If air is used as the working fluid, for example by cyclically injecting the H2 via a “displacer” valved cell, the air plus water/steam mix can be “dumped” each cycle. Note that an H2 displacer valved cell would be much larger than an O2 displacer valved cell.

C6H12 as a Regenerator Thermal Charging Fluid in a C6H12 Production System.

FIGS. 1 through 4 illustrate a means for accomplishing an isobaric cyclical regeneration. As suggested above, such an isobaric cyclical regeneration can utilize C6H12 as a “regenerator thermal charging fluid” in the conversion of C6H6+H2 to C6H12. A C6H12 regenerator charging fluid system would comprise the steps of:

• • a. Relatively cold C6H6 vapor at constant pressure (P1) and temperature (T1) would intermittently and cyclically pass through regenerator intake valve (In1), passing a stream of working fluid from the “cold” side of a regenerator (A) to the “hot” side, pass out through a port on the “hot” side of the regenerator, and pass through the intake valve (In2) and into a C6H6 receiver mechanism, such as a piston-and-cylinder arrangement (B). The result is to increase the temperature of the C6H6 at constant pressure, and simultaneously remove some or all of the regenerator's stored thermal energy.
  • b. Simultaneously, a second receiver mechanism, such as a piston-and-cylinder arrangement (C), would isobarically receive, at a similar pressure of P1 and a temperature of T2, a charge through its intake valve (In3) of relatively hot C6H12 vapor exiting an endothermic catalytic reactor.

Note, importantly, that the volume of the two receiver mechanisms may or may not be equal, depending on various factors such as relative temperatures at the end of this stroke or the chemical nature of the the two streams. FIG. 34 graphically indicates a case where, for example, a charge of vaporous C6H6+H2 is used to preheat a charge of vaporous C6H12, in this case in order to preheat C6H12 which is about to enter an endothermic reactor and be converted to a new charge of C6H6+H2.

• • c. Both receiver mechanisms, having filled completely, would now reverse direction. Simultaneously, valves In1, In2, and In3 will be closed, and valves Ex1, Ex2, and Ex3 will open. Valve Ex1 exhausts the heated C6H6, valve Ex2 exhausts the hot C6H12 and acts as a check valve when the C6H6 receiver is on its intake stroke, and valve Ex3 exhausts the cooled C6H12 from the regenerator. Note that the exhausts through Ex1 and Ex3 can be (a) into a slightly higher pressure environment, (b) have added resistance to opening, or (c) may be mechanically actuated, ensuring that neither check valve opens prematurely.
  • d. Both receiver mechanisms, having emptied completely, would now reverse direction, and the cycle would begin again.

Some mixing of remnant C6H12 would occur during the C6H6 thermal regeneration process, and some mixing of remnant C6H6 would occur during the C6H12 thermal de-generation process. Also, some amount of H2 would be “mixed in”, since the C6H12+H2 reaction is not likely to be equal to 100%. However, because of the relatively small internal area of the regenerator, that mixing can be limited. The net effect on the overall process is to reduce the amount of fluid converted per cycle. However, this is expected to only slightly impact the overall thermal efficiency of the process in a negative way.

In the instance described above, note that pressures for both C6H6 and C6H12 are isobaric and equal at all times. However, the heat exchange addition and removal processes may also be partially or wholly isochoric. To create isochoric heating of the C6H6, a displacement will be required from a third mechanism, or“isochoric displacer”. Note that the isochoric displacer would have approximately the same volume and stroke as the mated receiver mechanism, similarly to other displacement processes described and shown elsewhere in this document

In another possible STREP variant, the input (and output) of C6H12 can be isobaric while the input of C6H6 can be isochoric. In that instance, both the pressure and temperature of the C6H6 will increase during the isochoric portion of the displacement process. Note that, in this instance, valve Ex2 functions to disallow a high pressure flow back into the C6H12 receiver mechanism. In addition to Valve Ex2 acting as a check valve, regenerator exhaust valve Ex3 will require active sealing against the building pressure differential, which indicates that it will need to be manually operated. Finally, to “match” pressure across Ex2 following the two intake processes, Intl will be held open momentarily, allowing pressure to drop in the “displacer” (not shown) as it re-expands trapped C6H6 vapor in the regenerator on the following stroke down to the pressure of the C6H12 receiver mechanism. Note that, with the stroke reversal, the C6H12 receiver mechanism will simultaneously begin to increase pressure, thus “helping” match pressure across valve Ex2. Ex3 can also be opened slightly before pressure equalizes between the regenerator and the second receiver mechanism, although that will unavoidably cause a sudden and inefficient pressure drop. A sensor on valve Ex2 could determine when valve Ex2 begins to move towards open due to pressure equalization, signaling a solenoid to close In1.

Having a “mixed” isobaric and isochoric process is especially beneficial for the C6H6+H2 to C6H12 conversion process, since the latent heat requirement for a given mass of fluid is reduced for an isochoric process in relation to an isobaric process. C6H6 is a less dense fluid than C6H12, and thus requires less latent heat per change in degree temperature. Using an isochoric flow for the C6H12 delivers less heat. And since the mol count will be ideally equal for both C6H6 and C6H12, such a mixed isobaric and isochoric process is beneficial by allowing the C6H12 to approach giving up just sufficient latent heat to supply the requirement for preheating the C6H6 up to the temperature required for the exothermic generation of said C6H12.

In one approach to utilizing this process in a heat engine, the C6H6 is initially at 1.5 atm and approximately 670 R, while the C6H12 is initially at 1.5 atm and 1000 R. At the end of the isochoric regeneration process, C6H6 is assumed to be at 1.5 atm and 1000 R and C6H12 is assumed to be at 1.0 atm and 670 R. Note that the latent heat of C6H12 condensation is still available for vaporizing C6H6 liquid.

Of course, the C6H6 vapor thus produced would still require E.R.E.C compression, and the required H2 to complete the reaction to C6H112 would still require gaseous compression. Since the temperature (1,000 R) at which the endothermic reaction takes place is explicitly tied to the pressure (1.5 atm) at which the endothermic reaction takes place, and since ideally the only output from the endothermic reactor is C6H12, the required W-in of pumping, E.R.E.C. compression, and H2 compression make this process a net consumer of work. That can amount to very little work required at close to atmospheric pressure, but those low pressures will also limit the temperature of the exothermic reaction produced.

Consequently, the work required to produce C6H12 will have to be laid against any work produced by the exothermic heat. For example, using exothermic heat to help power an “isochoric source heating+exothermic preheating” engine such as has been described above means overall thermal efficiency will be a function of that cycle's W-in requirement plus the W-in requirement of C6H12 production subtracted from the net W-out of the overall process.

Finally, there's the intriguing possibility of a combined regenerator/exothermic reactor STREP (FIG. 38). It's just barely possible that a catalyst can be integrated with a regenerator. In FIG. 38, the large piston on the left represents preheated vaporous/gaseous exothermic fluid, for example one mol of C6H6 and three moles of H2. As is clear from U.S. Pat. No. 3,225,538, the conversion to one mol of C6H12 will occur at both constant pressure and constant temperature. However, volume will be about a quarter the size following the conversion. Hence the much smaller piston on the left that has an intake when the larger piston on the left has an exhaust. The single piston on the right is representing a constant pressure intake and exhaust, taking in a cold high density but completely non-reactive substance through the catalytic reactor/regenerator and exhausting the hot high density substance on the following stroke.

Because the same pressures are seen for all pistons, the net W-in required is only due to friction, pumping, and thermal leakage losses. Ideally, no W-in would be required, outside of the work of pressurizing, and that would be balanced by both W-out and the generation of, and potential utilization of, thermal energy from the exothermic reactor.

Although specific examples are described herein, the scope of the technology is not limited to those specific examples. Moreover, while different examples and embodiments may be described separately, such embodiments and examples may be combined with one another in implementing the technology described herein. One skilled in the art will recognize other embodiments or improvements that are within the scope and spirit of the present technology. Therefore, the specific examples disclosed are not to be interpreted in a limiting sense. The scope of the technology is defined by the following claims and equivalents thereof.

Claims

1. For efficiently exchanging heat between two streams of fluid at approximately equal pressure while simultaneously reducing the internal volume and general overall mass of the heat exchange means per quantity of heat exchanged over time, a means termed a Synchronized Thermal Regenerator Exchange Pump (STREP) composed of (1) a piston and cylinder means termed a receiver, said receiver having an intake valve means and an exhaust valve means, (2) a second piston and cylinder means termed a synchronizer, said synchronizer having an intake valve means and an exhaust valve means, (3) a hollow housing means containing a metallic sponge or regenerator means termed a regenerator, said regenerator having an intake valve means, an exhaust valve means, a port means connected to the exhaust valve of said receiver means, and a second port means connected to the synchronizer intake valve means, (4) piston movement means such as a crankshaft with a crank throw and a connecting rod between the crankshaft and a piston connecting pin termed a prime mover, said prime mover able to move said receiver piston and said synchronizer piston synchronously such that both pistons will reach Top Dead Center (TDC) and Bottom Dead Center (BDC) simultaneously and cyclically, and (5) a force means for operating said prime mover means, where a charge of fluid at a given temperature is taken in by (a) said receiver means through (b) said receiver intake valve means as (c) said receiver piston means is moved by said (d) prime mover means from TDC to BDC, and where a second charge of fluid at a different temperature is simultaneously and synchronously taken in by (e) said synchronizer piston means past (f) the intake valve means of (g) said regenerator means, (h) through said regenerator means, (i) out said regenerator port means connected to (j) said synchronizer intake valve means, and (k) into said synchronizer means, where in the process of said synchronizer cylinder means taking in said charge of fluid heat is either given up to said regenerator means or removed from said regenerator means, thus raising or lowering the temperature of the fluid entering said synchronizer means, and where, upon reaching BDC, both pistons simultaneously and synchronously (l) reverse direction and begin moving towards TDC by action of said prime mover means, whereby said receiver cylinder, synchronizer cylinder, and regenerator means' intake valve means ideally instantaneously, simultaneously and synchronously (m) close as said receiver cylinder, synchronizer cylinder, and regenerator means' exhaust valve means ideally instantaneously, simultaneously and synchronously (n) open, resulting in the simultaneous (o) expulsion of fluid out of said receiver and synchronizer cylinder means by (p) action of said piston means driven by (q) said prime mover means, in the case of said synchronizer cylinder means its fluid being (r) driven out of its exhaust valve means while in the case of said receiver cylinder means its fluid being (s) driven out of its exhaust valve means, (t) through said regenerator means, and (u) out said regenerator exhaust valve means, where in the process of said receiver means (v) passing fluid through said regenerator means its charge of fluid will either (w) receive heat from or give up heat to said regenerator means, said heat having been earlier either (x) deposited by or removed from said regenerator means but in any case the opposite effect achieved by said synchronizer means, thus resulting in an efficient exchange of heat between the two streams at approximately equal pressure while simultaneously reducing the internal volume per quantity of heat exchanged over time and general overall mass.