阅读说明：本技术 旋转式回热器 (Rotary heat regenerator ) 是由 特里斯特拉姆·布雷西 于 2019-05-28 设计创作，主要内容包括：一种用于在流体流之间进行热交换的热交换器结合有旋转特征以控制所述流的通过。(A heat exchanger for exchanging heat between fluid streams incorporates a rotational feature to control the passage of the streams.)

1. A recuperative heat exchanger module comprising: -

a) A chamber;

b) at least one first inlet and at least one first outlet of the chamber for passing at least a portion of the first fluid stream;

c) at least one second inlet and at least one second outlet of the chamber for passing at least a portion of the second fluid stream;

d) at least one rotatable body within the chamber, the rotatable body mounted for rotation about an axis;

it is characterized in that the preparation method is characterized in that,

the at least one body comprises a plurality of fluid flow passages extending transverse to the axis of rotation and separated by a fluid permeable heat transfer medium, the plurality of fluid flow passages being arranged such that in use

i. In a first rotational position, the plurality of fluid flow passages are aligned with the at least one first inlet and the at least one first outlet such that the at least a portion of the first fluid flow enters at least one of the fluid flow passages through the at least one first inlet, passes through the fluid permeable heat transfer medium into at least one adjacent fluid flow passage, and passes through the at least one adjacent fluid flow passage into the at least one first outlet, and

in a second rotational position, the plurality of fluid flow passages are aligned with the at least one second inlet and the at least one second outlet such that the at least a portion of the second fluid flow enters at least one of the fluid flow passages through the at least one second inlet, passes through the fluid permeable heat transfer medium into at least one adjacent fluid flow passage, and enters the at least one second outlet through at least one adjacent fluid flow passage.

2. The recuperator module of any preceding claim, wherein the at least one first inlet and the at least one first outlet are circumferentially opposite one another, and the at least one second inlet and the at least one second outlet are circumferentially opposite one another.

3. The recuperative heat exchanger module of any preceding claim, wherein the at least one of the fluid flow passages into which the at least a portion of the first fluid flow enters from the at least one first outlet in the first rotational position is also the at least adjacent one of the fluid flow passages into which the at least a portion of the second fluid flow enters the second outlet in the second rotational position, and wherein the at least one adjacent one of the fluid flow passages into which the at least a portion of the first fluid flow enters in the first rotational position is also the at least one of the fluid flow passages into which the at least a portion of the second fluid flows in the second rotational position.

4. A recuperative heat exchanger module comprising a chamber according to any preceding claim, wherein the rotatable body is housed in a drum and the drum is housed in a pressure vessel such that the drum is adapted to rotate with the rotatable body and the pressure vessel remains stationary.

5. The recuperator module of claim 4, further comprising external perforations in said pressure vessel and said drum, said external perforations aligned to seal with said first inlet and said first outlet in said first rotational position, and aligned to seal with said second inlet and said second outlet in said second rotational position, and wherein said external perforations in said pressure vessel and said drum are also aligned with at least one of said plurality of fluid flow passages in said fluid permeable heat transfer medium.

6. The recuperator module of any of claims 4-5, including a circular sliding interface between said stationary pressure vessel and said drum, effecting a sealing tight engagement for preventing the passage of the second fluid flow in said first rotational position and preventing the passage of the first fluid flow in said second rotational position.

7. The recuperator module of any preceding claim, wherein the fluid permeable heat transfer media is arranged in a honeycomb matrix structure or mesh arrangement, or as a honeycomb block stacked together.

8. The recuperator module of any preceding claim, wherein the at least one first inlet, the at least one first outlet, the at least one second inlet, the at least one second outlet, the fluid permeable heat transfer medium, and the plurality of fluid flow passages are coplanar with respect to a plane perpendicular to the axis of rotation.

9. The recuperative heat exchanger module of any of claims 1-7,

i) the at least one first inlet and the at least one first outlet are non-coplanar with one another about a plane perpendicular to the axis of rotation;

ii) the at least one second inlet and the at least one second outlet are not coplanar with each other about a plane perpendicular to the axis of rotation;

iii) the plurality of fluid flow passages are distributed along the axis of rotation such that at least one fluid flow passage is coplanar with each of the at least one first inlet, the at least one first outlet, the at least one second inlet, and the at least one second outlet about a plane perpendicular to the axis of rotation;

iv) the fluid permeable heat transfer medium is distributed along the axis of rotation such that in a first rotational position, the at least a portion of the first fluid flow passes through the fluid permeable heat transfer medium from a fluid flow passage coplanar with the at least one first inlet with respect to a plane perpendicular to the axis of rotation to at least one adjacent fluid flow passage coplanar with the at least one first outlet with respect to a plane perpendicular to the axis of rotation, and in a second rotational position, the at least a portion of the second fluid flow passes through the fluid permeable heat transfer medium from a fluid flow path coplanar with the at least one second inlet with respect to a plane perpendicular to the axis of rotation to at least one adjacent fluid flow path coplanar with the at least one second outlet with respect to a plane perpendicular to the axis of rotation.

10. The recuperator of claim 9, wherein the at least one first inlet is coplanar with the at least one second outlet about a plane perpendicular to the axis of rotation, and the at least one second inlet is coplanar with the at least one first outlet about a plane perpendicular to the axis of rotation.

11. The recuperative heat exchanger of any preceding claim, wherein the first fluid flow is at a higher pressure than the second fluid flow, and wherein the second fluid flow enters the heat exchanger at a higher temperature than the first fluid flow, and wherein the first fluid flow exits the heat exchanger at a higher temperature than when the first fluid flow enters, and the second fluid flow exits the heat exchanger at a lower temperature than when the second fluid flow enters.

12. An assembly of at least two recuperative heat exchangers according to any of the preceding claims coupled to allow the first and second fluid flows to pass between the heat exchangers.

13. A system for effecting a turbine cycle comprising a turbine, a combustor and at least one recuperator according to any one of claims 1 to 11 or an assembly of at least two recuperators according to claim 12.

14. The system of claim 13, wherein the turbine is an outer rotor turbine made primarily of ceramic material.

15. The system of claim 13 or 14, wherein the burner is a flameless burner system comprising:

a longitudinally extending combustion chamber through which oxidant flows in a longitudinal direction from an inlet to an outlet, an

At least one fuel line;

wherein at least one fuel injection line in fluid communication with the fuel line extends through a wall of the combustion chamber, wherein each fuel injection line comprises at least one fuel injector for injecting fuel into the combustion chamber.

16. An automobile or an engine for an automobile comprising at least one recuperative heat exchanger according to any of claims 1 to 11, an assembly of at least two recuperative heat exchangers according to claim 12, or a system according to any of claims 13 to 15.

17. A static electricity generator comprising at least one recuperative heat exchanger according to any of claims 1-11, an assembly of at least two recuperative heat exchangers according to claim 12, or a system according to any of claims 13-15.

Technical Field

The feature is a rotary regenerator or recuperator used as a heat exchanger. A gas turbine cycle in which the rotary regenerator may be used is also described. Such regenerators and/or turbine cycles utilizing them may be used in a range of applications including, but not limited to, automotive applications and static power generation.

Background

A recuperative heat exchanger (or regenerator) is a heat exchanger designed to exchange heat between two streams of different temperatures. A flow of hot fluid (typically a gas) is brought into contact with a solid thermal mass or medium, thereby raising the temperature of the medium over a period of time. The cold fluid stream then comes into contact with the now hot medium for a period of time due to a certain switching in the flow paths, thereby transferring heat away through the same surface. The cold and hot fluids may pass in opposite directions (counter-flow) or in the same direction (co-flow), but the former is generally more efficient.

This is in contrast to typical heat exchangers where the cold and hot streams are permanently sealed off and heat must constantly flow completely through the solid separating medium without the fluids being able to flow in reverse.

"permanent magnet" type rotary regenerators are regenerators in which a cylindrical heat storage medium formed by a mesh or honeycomb is rotated about its axis. In its simplest form, the front face of the thermal storage medium is divided by a seal into two sections that are nominally gas-impermeable, and hot and cold gases flow through either section simultaneously in a direction parallel to the axis of rotation and perpendicular to the front face of the thermal storage medium. As the thermal medium rotates, the thermal medium absorbs heat from the hot stream and then transfers this heat to the cold stream.

The "Rothemuhle" type regenerator uses the same principle, but uses a stationary heat accumulation wheel, with the fluid flow being carried through rotating shrouds that direct the flow sequentially to different portions of the media, while sliding sealing surfaces on the heat accumulation wheel to prevent the high pressure flow from turning to the low pressure flow.

US 138198, US 391727 and US 276515 show rotary regenerators in which the inlet and outlet ducts are oriented such that the inflow and outflow of the device is parallel to the axis of rotation of the thermal media wheel. Thus, a seal between the hot and cold streams, typically at different pressures, is formed on the surface of the thermal media wheel. Such regenerators present significant sealing problems because the materials ideally used as high temperature thermal storage media do not lend themselves to large pressure differentials or to wear of sliding interfaces. For example, the walls of ceramic honeycomb structures are thin and the constant application of a pushed sliding interface to the front face of such media can cause them to chip, creating gaps that break the seal.

The prior art, such as US 950707, utilizes a novel sealing mechanism for a "perpetual stirling" or "lottimylor" type regenerator. Such seals use complex geometries and cannot be used with materials required for high temperature applications.

Other solutions ("The design of high efficiency turbomachinery and gas turbines" by d.g.wilson and t.korakianitiss, 2 nd edition, massachusetts institute of technology, 2014 ") also use indexed rotation, in which a seal on The hot media wheel is temporarily formed by a mechanically deployable seal (perhaps hydraulic or perhaps pneumatic) that engages when The thermal storage wheel is stationary, allowing The wheel to rotate and switch The passage of hot and cold flows in and out. The advantage is that the sealing surface does not need to slide on the hot medium wheel but can engage gently down onto the hot wheel surface. A disadvantage is that when the seals are not engaged, leakage from the high pressure stream to the low pressure stream occurs, requiring a fast rotating wheel (or sealing mechanism) to minimize this. Thus, due to the reciprocating and faster motion, such regenerators are mechanically subjected to greater forces than continuous rotary regenerators, thereby reducing the life of any bearings, and the seals need active control, while more moving parts and subsystems are operating at high temperatures, thereby reducing reliability.

Conventional regenerators suffer from a phenomenon known as "carry-over loss" when exchanging heat between differential pressure streams. Carryover losses occur because a certain amount of high pressure gas is periodically sealed in the air space inside the thermal medium and then released into the low pressure stream. The work required to pressurize this amount of gas will be delivered to the system outlet in a burst. Instead, the amount of low pressure gas carried to the high pressure stream causes a burst of flow into the regenerator media to equalize the pressure. In the case of a gas turbine cycle, the effect of carryover losses is to reduce the available high pressure flow to the turbine and increase the wasted compression work by discharging the high pressure fluid, resulting in reduced system efficiency. A typical design requirement may be to ensure that the entrained gas mass flow is small, possibly below 1%, compared to the nominal gas mass flow through the system.

Regenerators are typically used as part of larger systems, such as systems implementing gas turbine cycles. A common gas turbine cycle is the regenerative open brayton cycle. The working fluid (usually air) is extracted from the atmosphere, compressed, raised in temperature from the turbine exhaust using a regenerator, and further raised in temperature using heat input (usually using fuel in the combustor). The turbine extracts work and reduces the temperature and pressure of the fluid. The turbine exhaust is used to preheat the incoming compressed working fluid. The turbine may typically drive a compressor directly attached via a common shaft, and may additionally drive a generator to generate electricity. Matching components to complement each other within a cycle may maximize the efficiency of the cycle.

The thermal efficiency of a gas turbine cycle is defined as the ratio of electricity generation to heat energy input (typically by fuel combustion), typically increasing with increasing size (Bejan, a., Lorente, s., Yilbas, b., & ahin, a. (2011.) the effect of size on efficiency: power plant and maintenance design. international journal of heat and mass transfer, 54(7), 1475-. Single wheel cycles producing <50MW net power output typically do not achieve efficiencies greater than about 35%. This is because larger turbines lend themselves to greater complexity (e.g., through bucket cooling techniques, or steam injection during the combustion phase). The increased complexity, use of dissimilar metals, and/or the requirement to purify the water for steam injection results in increased costs, which are only acceptable at larger scales. The net power output of the cycle may be increased by using a combined cycle system, where the gas turbine cycle is combined with, for example, a steam cycle that extracts more energy from the turbine exhaust. Combined cycles such as a brayton cycle combined with a steam cycle add complexity and are generally not feasible on a small scale due to cost, difficulty in forming steam, and availability of small scale steam turbine technology.

Nitrogen Oxides (NO) will also typically be present in an open Brayton cycle_x) The formation of contaminants. Large gas turbine systems may utilize an exhaust treatment system to remove most, but not all, of the pollutants, however, such treatment increases the complexity, cost, and space usage of small gas turbine systems.

The small scale brayton cycle requires a turbine that can be reasonably efficient. At high temperatures and low power output, there is a need to find an efficient heat recovery device that can extract enough heat from the exhaust of the turbine such that a combined cycle is not required to increase efficiency (such as the rotary regenerator described above), and a low emission combustor that can burn the required fuel without the need for cooling fluids that would reduce efficiency.

In the prior art there is a high efficiency of (>40% power generation heat input), however these systems typically need to be implemented on a large scale to achieve these efficiencies at a reasonable cost. For example, U.S. patent 6622470B 2 shows a semi-closed Brayton cycle that utilizes pure oxygen as the oxidant to change the working fluid from air to a mixture of water, carbon dioxide, and oxygen to eliminate NO_xThe possibility of discharge. Due to the need for air separation, the system is necessarily more complex and only feasible on a large scale. Similar semi-closed systems exist that use a working fluid instead of air (e.g., an Allam cycle using supercritical carbon dioxide), however, both of these systems present greater complexity than open Brayton cycles. (before any other consideration is given, there is the fact that although air is freely available from the atmosphere, these alternative working fluids must be provided specifically.

Other prior art uses a brayton cycle with intermediate heat input to eliminate the associated problems with emissions, system integration, and combustor design that result from combusting hydrocarbons to provide thermal energy. Us patent 5873250 a describes an open brayton cycle with an efficiency of 48-50% powered by stored thermal energy for use in electric vehicles. This patent describes the possibility of burning fuel to regenerate the stored thermal energy, however this action affects the low pollution nature of the invention, resulting in high emission cycles.

Other open brayton micro-turbine cycles exist, but are not generally used for efficient power generation. An example of a micro-turbine system using a brayton cycle is shown in the prior art, such as us patent 6170251B 1. This prior art describes a micro-turbine system that provides a secondary output of compressed air. Since most of the turbine's work is used to compress the air, the power generation efficiency of this system will be low.

By increasing the complexity of the cycle, a high efficiency micro-turbine using a brayton cycle can be achieved. For example, us patent 20120324903 a1 describes a high efficiency multi-rotor brayton cycle with intercooling between compressors on separate shafts. Or us patent 9279364B 2 describes a gas turbine cycle with multiple combustors. This type of complexity adds an inherent complication to the control of the cycle and the overall design of the system, and therefore would not be cost effective if built at the micro-turbine level.

There remains a need for a highly efficient, low complexity micro gas turbine cycle that is feasible on a small scale and capable of utilizing multiple fuels while continuing to maintain low emissions.

Disclosure of Invention

In a first aspect, the present invention provides a recuperative heat exchanger module comprising:

a) a chamber;

b) at least one first inlet and at least one first outlet of the chamber for passing at least a portion of the first fluid stream;

at least one second inlet and at least one second outlet of the chamber for passing at least a portion of the second fluid stream; and

d) at least one rotatable body within the chamber, the rotatable body being mounted for rotation about an axis;

wherein at least one body comprises a plurality of fluid flow passages extending transverse to the axis of rotation and separated by a fluid permeable heat transfer medium, the plurality of fluid flow passages being arranged such that, in use:

a) in the first rotational position, the plurality of fluid flow passages are aligned with the at least one first inlet and the at least one first outlet such that at least a portion of the first fluid flow enters the at least one fluid flow passage through the at least one first inlet, enters the at least one adjacent fluid flow passage through the fluid permeable heat transfer medium, and enters the at least one first outlet through the at least one adjacent fluid flow passage, and

b) in the second rotational position, the plurality of fluid flow passages are aligned with the at least one second inlet and the at least one second outlet such that at least a portion of the second fluid flow enters the at least one fluid flow passage through the at least one second inlet, enters the at least one adjacent fluid flow passage through the fluid permeable heat transfer medium, and enters the at least one second outlet through the at least one adjacent fluid flow passage.

For the purposes of the present invention, the term "first inlet" is understood as "first flow inlet" in the sense that the function of the at least one first inlet is to allow a first fluid flow into the chamber. Likewise, the term "first outlet" may be understood as "first outflow outlet" in the sense that the function of the at least one first outlet is to allow a first fluid flow out of the chamber. Likewise, the term "second inlet" may be understood as "second flow inlet" in the sense that the function of the at least one second inlet is to allow a second fluid flow into the chamber. Likewise, the term "second outlet" may be understood as "second outlet" in the sense that the function of the at least one first outlet is to allow a second fluid flow out of the chamber.

Optionally, the at least one first inlet and the at least one first outlet are circumferentially opposite to each other and the at least one second inlet and the at least one second outlet are circumferentially opposite to each other.

Optionally, at least one of the fluid flow passages into which at least a portion of the first fluid stream enters from the at least one first outlet in the first rotational position is also at least an adjacent flow passage into which at least a portion of the second fluid stream enters the second outlet in the second rotational position, and wherein the at least one adjacent flow passage into which at least a portion of the first fluid stream flows into the at least one first outlet in the first rotational position is also at least one of the fluid flow passages into which at least a portion of the second fluid flows in the second rotational position.

Optionally, the rotatable body is housed in a drum, and the drum is housed in a pressure vessel, such that the drum is adapted to rotate with the rotatable body, and the pressure vessel remains stationary. In this case, the heat exchanger may optionally further comprise external perforations in the pressure vessel and the drum, the external perforations being aligned to seal with the first inlet and the first outlet in the first rotational position and the second inlet and the second outlet in the second rotational position, and said external perforations in the fixed pressure vessel and the drum being further aligned with at least one of the plurality of fluid flow passages in the fluid permeable heat transfer medium.

Optionally, where a pressure vessel and a rotating drum are used, there is a circular sliding interface between the stationary pressure vessel and the rotating drum, thereby achieving a sealing tight engagement for preventing the passage of the second fluid flow in the first rotational position and preventing the passage of the first fluid flow in the second rotational position.

Alternatively, the fluid permeable heat transfer media is arranged in a honeycomb matrix structure or mesh, or arranged as honeycomb blocks stacked together.

Optionally, the at least one first inlet, the at least one first outlet, the at least one second inlet, the at least one second outlet, the fluid permeable heat transfer medium, and the plurality of fluid flow passages are coplanar with respect to a plane perpendicular to the axis of rotation.

Alternatively, a more complex arrangement may be used, in which:

i) the at least one first inlet and the at least one first outlet are non-coplanar with each other about a plane perpendicular to the axis of rotation;

ii) the at least one second inlet and the at least one second outlet are not coplanar with each other about a plane perpendicular to the axis of rotation;

iii) the plurality of fluid flow passages are distributed along the axis of rotation such that the at least one fluid flow passage is coplanar with each of the at least one first inlet, the at least one first outlet, the at least one second inlet, and the at least one second outlet about a plane perpendicular to the axis of rotation;

iv) the fluid permeable heat transfer medium is distributed along the axis of rotation such that in a first rotational position at least a portion of the first fluid flow passes through the fluid permeable heat transfer medium from a fluid flow passage coplanar with the at least one first inlet with respect to a plane perpendicular to the axis of rotation into at least one adjacent fluid flow passage coplanar with the at least one first outlet with respect to a plane perpendicular to the axis of rotation, and in a second rotational position at least a portion of the second fluid flow passes through the fluid permeable heat transfer medium from a fluid flow passage coplanar with the at least one second inlet with respect to a plane perpendicular to the axis of rotation into at least one adjacent fluid flow passage coplanar with the at least one second outlet with respect to a plane perpendicular to the axis of rotation.

Where a more complex arrangement is used, optionally, the at least one first inlet is co-planar with the at least one second outlet about a plane perpendicular to the axis of rotation, and the at least one second inlet is co-planar with the at least one first outlet about a plane perpendicular to the axis of rotation.

Optionally, the first fluid stream is at a higher pressure than the second fluid stream, and wherein the second fluid stream enters the heat exchanger at a higher temperature than the first fluid stream, and wherein the first fluid stream exits the heat exchanger at a higher temperature than when the first fluid stream entered, and the second fluid stream exits the heat exchanger at a lower temperature than when the second fluid stream entered.

In another aspect, the present invention provides an assembly of at least two recuperative heat exchangers as described above coupled to allow a first fluid flow and a second fluid flow to pass between the heat exchangers.

In another aspect, the present invention provides a system for validating a turbine cycle comprising a turbine, a combustor and at least one recuperative heat exchanger or an assembly of a plurality of heat exchangers as described above. The turbine may be an external turning wheel turbine made primarily of ceramic material. The combustor may be a flameless combustor system comprising a longitudinally extending combustion chamber through which oxidant flows in a longitudinal direction from an inlet to an outlet, and at least one fuel line; wherein at least one fuel injection line in fluid communication with the fuel line extends through a wall of the combustion chamber, wherein each fuel injection line comprises at least one fuel injector for injecting fuel into the combustion chamber.

In another aspect, the present invention provides an automobile or an engine for an automobile comprising at least one recuperative heat exchanger according to any of claims 1 to 11, an assembly of at least two recuperative heat exchangers according to claim 12, or a system according to any of claims 13 to 15.

In another aspect, the invention provides a static electric generator comprising at least one recuperator according to any one of claims 1-11, an assembly of at least two recuperators according to claim 12, or a system according to any one of claims 13 to 15.

Drawings

Fig. 1 depicts a module of a modular regenerator according to the present invention.

Fig. 2 depicts a cross-sectional view of the module of fig. 1 at plane a-a.

Fig. 3 depicts a cross-sectional view of the module of fig. 2 at plane B-B.

Fig. 4 depicts a detailed view of block section C of fig. 3.

Fig. 5 depicts a drum of a modular regenerator in accordance with the present invention.

Fig. 6 depicts a cross-sectional view of the drum of fig. 5 at plane F-F.

Fig. 7 depicts in detail the module of fig. 1 and the connections between the same modules.

Fig. 8 consistently depicts a series of five modules according to fig. 1.

Fig. 9 is a graph of diagonal phase versus drum rotation angle for a series of five modular regenerators, as in fig. 1, connected in series.

Fig. 10 depicts fluid flow through a modular regenerator, such as in fig. 1, in the cooling and heating portions of a cycle.

FIG. 11 depicts a greatly simplified version of the invention in which almost all of the components are coplanar with one another about a plane perpendicular to the axis of rotation.

FIG. 12: the system is proceeding with the process of incorporating a rotary recuperator according to the present invention into a turbine cycle.

FIG. 13: as in fig. 12, but also includes an alternative fuel heat exchanger for heat recovering the exhaust gas.

FIG. 14: as in fig. 12, but includes an optional fuel compressor.

FIG. 15: as in fig. 12, but including an optional high pressure burner for start-up purposes

FIG. 16: as in fig. 12, but with the compressor replaced by a multi-stage intercooled shaft driven compressor.

FIG. 17: as in fig. 12, but includes an alternative fuel vaporizer for use with liquid fuel.

FIG. 18: as in fig. 12, but with the compressor replaced by a shaftless driven compressor.

FIG. 19: as in fig. 12, but includes an optional additional motor for starting purposes.

FIG. 20: as in fig. 12, but includes an optional electric air heater for startup purposes.

FIG. 21: as in fig. 12, but includes an alternative cascade burner system for start-up purposes.

FIG. 22: as in fig. 12, but includes an optional auxiliary burner for start-up purposes.

FIG. 23: as in fig. 12, but includes an optional auxiliary compressor for starting purposes.

FIG. 24: as in fig. 12, but including an optional exhaust gas heat exchanger for further heat recovery.

FIG. 25: based on an example of a potential static power generation process, many additions are incorporated into the general system of the system of FIG. 12.

FIG. 26: many additions are incorporated into the general system of the system of fig. 12 based on a potential automotive range extender application example.

FIG. 27 is a schematic view showing: a general start-up procedure of a static power generation system as depicted in fig. 25 is shown.

Detailed Description

The invention is a heat exchanger with rotating components. The present invention may be implemented in a simple geometry in which the inlet and outlet ports and the fluid permeable heat transfer medium are coplanar about a plane perpendicular to the axis of rotation, and as the rotatable body 160 rotates, the fluid permeable heat transfer medium is exposed first to the first inlet port 110 and the first outlet port 116, and then to the second inlet port 112 and the second outlet port 114, such that the fluid flows alternately from the first inlet port to the first outlet port, and then from the second inlet port to the second outlet port through the fluid permeable heat transfer medium 122.

A more complex arrangement is depicted in fig. 1. Fig. 1 depicts a high temperature rotary regenerative heat exchanger comprising at least one thermal wheel cell made of a suitable high temperature material or thermal medium, such as a ceramic honeycomb contained within a suitable pressure vessel or module.

By using radial inflow and outflow of hot and cold fluids to the chamber (containing the hot wheel), the classic problem of sealing the front face of a honeycomb heat transfer media wheel is avoided in the present invention. The fluid flow enters and exits the chamber in a direction perpendicular to the axis of rotation of the hot wheel. This is different from the prior art, where the flow enters parallel to the axis of rotation, perpendicular to the open face of the thermal medium. The invention allows an efficient heat exchange between two streams (hot and cold) that may have a pressure difference.

Fig. 1 depicts a single module, at least one of which is constructed, externally of which is a fixed pressure vessel, such as a closed cylinder. Within each stationary pressure vessel is an inner rotating drum, preferably stainless steel, which can rotate and house at least one hot wheel. Multiple modules may also be connected in series as depicted below.

Preferably, the flow paths through the thermal medium (i.e., along the channels of the honeycomb substrate, or through the mesh) are aligned with the axis of rotation of the drum. The drum may be slowly rotated, typically no more than about 10rpm, by a suitable drive system.

In the case of modules connected in series, the rotational speed of the shaft may be specific to each module or consistent between all modules within the rotary regenerator. The increase of the rotation speed will increase the heat exchange efficiency (eta)_eff) Heat exchange performance is defined as the ratio of actual heat transfer to the maximum possible heat transfer. However, an increase in rotational speed also results in an increase in carry-over losses, which translates into an increase in the parasitic load required to compress the high-pressure cold stream. Thus, for each application, depending on different process variables, these variables include, but are not limited to, flow rate, acceptable carryover loss, desired efficiency (. eta.),_eff) The number of modules used, the module size and the pressure difference, there is an optimum rotational speed. The optimum speed depends on the operator's desired output relative to the application-specific process variable for which the performance (η) is desired_eff) A compromise is made with the carry-over loss.

For example, in a typical gas turbine cycle, assuming all other variables remain constant, then the losses and efficiencies (η) are carried over_eff) The compromise between itself will be shown in the thermal efficiency versus rotational speed curve of the cycle. The shape of the curve depends on other components in the cycle, and the carryover loss and efficiency (η) of the regenerator_eff) The effect on a specific cycle. In the event that the operator wants to maximize the thermal efficiency of the gas turbine cycle, the optimum speed of the regenerator will be at the peak of the curve. In some gas turbine cycles or other applications, it may be desirable to maximize the temperature output of the cold input stream, or to minimize carryover losses (at the cost of efficiency). Therefore, the optimum rotational speed may not correspond to the maximum thermal efficiency of the cycle.

Hot and cold streams are carried into and out of each module through external perforations in the vessel. Perforations or exposure openings are made in the drums, preferably made of stainless steel and optionally lined with insulating material, so that at certain angular positions the fixed outer perforations are aligned with the rotating inner perforations, thereby allowing a flow to flow radially in by the delivery of a fluid flow, and its corresponding flow radially out of each drum. The conduit may be fabricated or cast or manufactured in any manner known to those skilled in the art. At any given time, the module will either allow flow between the inlet and outlet of the hot stream, or allow flow between the inlet and outlet of the cold stream, or neither. Preferably, the exposure ports in the vessel and drum may be spaced apart to align with the passages formed on either side of or between the ceramic honeycomb blocks. Preferably, the exposure apertures are diametrically opposed such that each of the hot or cold streams has an open flow path for up to 180 degrees of rotation. The circumferential positioning of the exposure opening will determine the amount of rotation per cycle, which may vary within the scope of the invention.

A circular sliding interface exists between the inside of the outer fixed pressure vessel and the outside of the bowl, whereby this interface forms a seal that blocks the flow of one stream while allowing the other stream to enter.

Preferably, if honeycomb media is used, the cross-section of the channels will be square, which allows the walls to have a constant thickness. Other shaped holes, such as circular holes, will also work within the scope of the present invention, but this is less preferred because the walls typically have a variable thickness, thereby reducing heat penetration and creating thermal stress. Typically, the width of the honeycomb cells is about 0.5mm to 2 mm. The size and cross-sectional shape of the honeycomb depends on the application and is not limited to a square cross-section or a particular width or height.

The thermal medium may be composed of any material having suitable heat capacity, thermal conductivity, thermal expansion, thermal shock characteristics, and heat resistance to the temperatures and pressures of the hot and cold streams. Suitable materials may include, but are not limited to, the family of materials for cordierite, mullite, alumina, calcium silicate, silicon carbide, and silicon nitride.

Preferably, when the drum is half way between alignment with the two streams to avoid opening both streams at once, there are small diagonal corners, none of which are open, and there is sufficient angular separation to form a seal. Thus, the reveal opening can be sized such that for drum rotation angles less than 180 degrees, the flow is open.

Preferably, multiple modules are used to ensure that the hot and cold streams always have a flow path through the regenerator, and as such will never be shut off. Each module may be phased in sequence so that both the hot and cold streams always have at least one module with an open flow path, but each module forms a tight seal when positioned between the exposed ports of either stream. Sequential phasing may be by exposure port positioning and sizing in the drum and drum exterior perforations or by rotational position of the drum in each module. For example, the shaft of a drum connected to one thermal wheel unit may also be connected to other thermal wheel units in a separate module, allowing all thermal wheel units within the rotary regenerator to rotate simultaneously.

If the regenerator is made up of a plurality of modules, each module will be connected by suitable piping known to those skilled in the art, so that modularity facilitates module replacement in the event of a failure. Preferably, bellows or other expansion joints are used to allow for thermal expansion between the pipes connected between the different modules.

Preferably, stub shafts may be attached to either end of each unit to transmit drive torque along a series of modules.

Preferably, the stub shafts may sit on suitable bearings to transfer the weight of the module back to a suitable frame. Otherwise, the unit may be supported on an outer container, or a combination of both, optionally using a height adjustment mechanism to change the load path.

Alternatively, the invention may be construed as operating with hot and cold flow that are neither continuous nor simultaneous, e.g. batch applications where a single module may suffice. (a single module is essentially a batch regenerator, but if multiple modules are connected in series, continuous operation can be achieved as a result). If the flows are continuous or simultaneous in a particular application, a single module can be used, provided that the back pressure and the regulated flow downstream of the regenerator can be accommodated. In most applications where the flows are continuous or simultaneous (e.g. typical gas turbines)Machine cycle), it is advantageous to provide the regenerator with multiple modules to avoid any back pressure or regulated flow. The number of modules and phasing of the reveal ports will be selected based on the appropriate characteristics of the application, including but not limited to flow rate and pressure, volume of the modules, geometry of the modules, manufacturing cost, efficiency of the regenerator (η [ ])_eff) Specific flow paths in the module, and the ability to adjust throttle and backpressure. Multiple modules also have the advantage of allowing for integration of potential redundancy, wherein failure of one module does not result in shutdown for servicing the entire regenerator.

Alternatively, a multi-module regenerator may be integrated into a single module cell with an internal dividing wall.

Alternatively, instead of the outer pressure vessel sliding against the drum, an inner wear plate may be attached inside the outer pressure vessel, which may be cheaper to replace. This requires that the wear plate has an exposure opening that matches the exposure opening of the outer pressure vessel so that flow is not impeded and can remain stationary. Optionally, additionally or alternatively, an additional wear plate may be added which is attached to rotate with the drum for the same reason.

Optionally, the sliding interface of the wear plate may be coated with a suitable lubricant. An example of a high temperature solid lubricant is hexagonal boron nitride. Liquid lubricants such as oil-based lubricants may burn off at temperatures that the ceramic thermal media can achieve, but it is possible to mitigate the consumption of such oils in view of the benefits of the system or use with other liquids. Thus, neither the use of a solid lubricant nor a liquid lubricant at the sliding interface departs from the present invention.

Alternatively, the fixed wear plate may be separated from the outer vessel by a layer of soft insulation. This serves to both protect the container from high temperatures and also helps to evenly distribute the clamping load over the wear plates, resulting in a more uniform seal.

Alternatively, the heat medium blocks may be cube shaped that can be stacked together into a single wheel layer such that the open flow face on each block does not contact the face of an adjacent block.

Alternatively, the thermal wheel may use any other three-dimensional shape that meets the above criteria, in addition to the cube block. For example, a ring segment shape is within the scope of the invention.

The closed flat outer faces of the block layers may be encapsulated with insulating layers that protect the drum from peak temperatures. The type and number of insulation layers within the drum will be selected based on the shape of the rotating drum and container, the given temperature range necessary for the material selection of the metal drum and outer container, and any structural requirements for a particular geometry. Materials used for insulation may be, but are not limited to, ceramic micro-porous, castable refractory, structural calcium silicate, insulation felt, and ceramic fiber products.

Optionally, any insulating layer adjacent to the flow path may be protected by a flow liner to prevent corrosion by the high velocity gases. Preferably, these liners will use stainless steel and be formed only from thin folded sheets. These flow liners may be sacrificial. Alternatively, the insulating layer may be selected to resist gas attack and may be used without a liner or with a suitable coating.

Preferably, the outer container is formed of an upper half and a lower half which are clamped together at longitudinal flanges on either side. This creates a preload, thereby forming a seal between the rotating interface and the stationary interface.

As described herein, having radial inflow and outflow allows for a more compact design than the prior art. This is because the thermal medium can be stacked into multiple interior walls within the thermal wheel, each separated by an offset that allows flow into the open passages between the wheels, turn 90 degrees, pass through the thermal medium, turn 90 degrees, and exit through the open passages. The prior art hot wheel requires that all flows enter the unit in the same direction parallel to the axis of rotation (i.e. perpendicular to the open face of the thermal medium), requiring that all thermal media be arranged in a single plane rather than in several parallel planes. The present invention greatly reduces the cross-sectional size of the inlet and outlet conduits for the same amount of thermal medium, making the system more compact, reducing heat loss, and stronger and less expensive to construct.

With the invention, the amount of fluid carried (with respect to the carrying losses) also includes the open space between the thermal media inside the rotor, which space is formed by the above-mentioned offset up to the sealing surfaces. Thus, in principle, the carry over losses are slightly higher than with conventional hot wheel designs. However, losses are mitigated by improvements such as scalability, compactness of design, sealing and reliability.

Alternatively, adjacent modules may be connected such that, to reduce carryover losses, a module that has completed its high pressure cycle (and is, for example, in a cryogenic state) may discharge its batch of pressurized air into a corresponding low pressure unit (e.g., at a high temperature prior to its introduction into the cryogenic high pressure stream) until the pressures equalize, thereby recirculating a portion of the internal pressure. The system can be expanded to recirculate pressurized air to a series of units or individual storage vessels to exchange and recirculate a greater portion of the pressure, albeit with ever increasing complexity.

Alternatively, the carryover losses may instead be recirculated by venting the high pressure stream through a turbocharger or a wave rotor that pressurizes the low pressure module with fresh air at ambient pressure using absorbed power.

The principle of flow entering the module radially (i.e. perpendicular to the axis of rotation) can be implemented in a number of ways without departing from the scope of the invention, for example, by aligning the sliding interface parallel to the axis of rotation before turning 90 degrees inside (rather than outside) the module to transfer radially towards the thermal medium.

The invention described herein also reduces the cost, complexity and efficiency (η) of high temperature regenerative heating by removing the high temperature material honeycomb seal_eff). In contrast to the heat medium, neither the outer stationary container nor the drum need to be made of a high temperature resistant material, but different higher strength engineering materials, such as steel grades, can be used. Thus, problems associated with seals that may be cellular high temperature thermal media are avoided. For example, the problem of thin walls on the honeycomb face breaking due to sliding seals, or the problem of complex deployable sealing interfacesReplaced by simpler problems of designing the sliding interface, for example in stronger and strong flat steel surfaces.

Examples of the invention are depicted in the accompanying drawings. The particular example illustrated is intended for use as part of a gas turbine cycle. The invention allows fluid flow perpendicular to the axis of rotation to enable heat exchange between two pressure differential flows at high temperature and/or high temperature differential. Under preferred conditions, high and low pressure fluids flow continuously throughout the regenerator with minimal carryover losses due to the direction of fluid flow. Although any combination of temperature and pressure, including equivalent pressure streams, may be achieved in the present invention, the description herein is based on heat exchange between a low pressure, high temperature feed and a high pressure, low temperature feed.

In the present invention, the rotary regenerator comprises at least one module 42 as depicted in fig. 1. The rotating portion of the thermal wheel unit 42 rotates relative to the axis by the stub shafts 38 that are attached to both sides of the metal drum 24 (as shown in fig. 3). Within the metal drum 24 is a rectangular parallelepiped ceramic honeycomb 20 surrounded by insulation of various geometries 18, 28, 30, 34 and 52 and duct flow liners 36 (as shown in fig. 2, 3 and 4). The rectangular parallelepiped ceramic is arranged by stacking smaller honeycomb segments 22 having open passages 44 (perpendicular to the direction of the honeycomb channels), as depicted in fig. 6. The passages 44 are aligned with openings or exposure ports 46 in the metal drum 24 that allow fluid communication between the rectangular parallelepiped ceramic honeycomb 20 and the inlet and outlet ducts 10, 14, 12, 16.

The conduit flow liner 36 protects the insulation layer 18 from the adverse effects of gas velocity. The insulation and conduit flow liner 36 extends to the hot inlet conduit 14 and outlet conduit 16. Wear plates 26 surround the entire circumference of the rotating metal drum 24 and are held in place by a housing or conduit clamp 32. An insulation layer 28 surrounds the wear plate 26 to reduce the degree of heat transfer to the conduit clamp 32. A suitable high temperature lubricant in the present invention, hexagonal boron nitride coated metal drum 24 contacts the face of wear plate 26.

Parallel to the axis of rotation are a connection line 2a for a high-pressure input fluid, a connection line 2b for a high-pressure output fluid, a connection line 2c for a low-pressure input fluid and a connection line 2d for a low-pressure output fluid. The inlet conduits 10, 14 and the outlet conduits 12, 16 perpendicular to the rotation axis allow fluid communication between the connecting ducts 2a, 2b, 2c, 2b through the inlet conduits 10, 14 and the outlet conduits 12, 16 to the metal drum 24, so that the inlet conduits 10, 14 and the outlet conduits 12, 16 are opposite the circumference of the drum 24 from the circumference of the connecting ducts 2a, 2b, 2c, 2 d.

In this embodiment, the metal drum 24 has exposure ports 46 sized as a portion of a circumference spaced longitudinally along the axis of rotation to allow fluid communication between the ceramic honeycomb 20 and the inlet and outlet conduits 10, 14, 12, 16. The exposure ports 46 are spaced diametrically opposite one another and are longitudinally staggered along the metal drum 24 for proper alignment with the associated conduits and open passages 44 within the rectangular parallelepiped honeycomb 20. The exposure port 46 is provided with a circumferential dimension across one side of the rectangular parallelepiped honeycomb 20 and a longitudinal dimension based on the flow velocity of the fluid.

As the regenerative unit rotates, the regenerative unit cycles during heating and cooling. During the heating period, the hot fluid exposure ports 46 allow fluid communication between the ceramic honeycomb 20 and the low pressure hot inlet duct 14. The hot, low pressure fluid flows through passages 44 into the honeycomb channels parallel to the axis of rotation and flows through passages 44 to the other side of the smaller stacked rectangular parallelepiped honeycomb 22 and then out of the metal drum 24.

Thus, the low-pressure fluid fills the channels in honeycomb stack 22, thereby transferring thermal energy to honeycomb stack 22. At the same time, the cold fluid exposure port allows fluid communication between the low pressure outlet conduit 12 and the ceramic cuboid honeycomb 20. Thus, the reveal ports 46 are positioned such that the heating period allows fluid communication between the low pressure inlet duct 14, the cuboid honeycomb 20 and the low pressure outlet duct 12. Hot low pressure fluid flows from the inlet duct 14 through the ceramic honeycomb channels, transferring heat energy and lowering the temperature, to the low pressure outlet duct 12 where it flows out as cold fluid.

During the cool down period, the exposure ports 46 are aligned with the high pressure fluid inlet conduit 10 and the high pressure outlet conduit 16, allowing fluid communication between the high pressure inlet conduit 10, the ceramic honeycomb 20, and the high pressure outlet conduit 16. In contrast to thermal cycling, the high pressure fluid flows in a counter-current direction from the high pressure inlet conduit 10 through the show openings 46 to the ceramic honeycomb 20 through the honeycomb channels. By extracting thermal energy from the ceramic honeycomb stack 22, the fluid temperature rises and exits through the exposure port 46 to the high pressure outlet conduit 16.

The unit is continuously cycled repeatedly to heat the ceramic honeycomb stack 22 with hot fluid and cool with cold fluid to periodically transfer energy from the hot fluid to the cold fluid. During the transition from the hot cycle to the cold cycle, the rotational speeds are balanced (-1-10 rpm) to maximize heat exchange while minimizing pressure losses or carryover losses in the system. After the heat cycle, any low pressure fluid remaining in the honeycomb channels will reduce the total pressure of the high pressure cold fluid entering the honeycomb channels during the cold cycle.

The individual modules 42 of the present invention may be used alone or may be advantageously connected with additional modules 42. For example, as depicted in FIG. 8, a preferred embodiment of the present invention includes five rotating wheel units 42. Fig. 7 depicts in detail the connection point between two modules of the type depicted in fig. 1 (and the motor 54 providing the rotational force). Each unit is connected via a connecting pipe 2 by a bellows 40, and the stub shafts 38 are connected by a coupling 50. With this arrangement, the rotation of all units can be driven by a single motor. Timing the hot and cold cycle cycles of each cell such that the low and high pressure streams always have a path through at least one module of the regenerator and the cross-sectional area of the rotary reveal ports 46 open to the low and high pressure streams is constant or nearly constant; this advantageously means that the flow rate is constant or nearly constant, thereby eliminating or greatly reducing the backpressure that would otherwise be enhanced. Fig. 9 illustrates the cycle period of the present invention. On the y-axis, "1" indicates the largest diagonal of the low pressure conduit 12, 14 to the reveal port 46, while "-1" indicates the largest diagonal of the high pressure conduit 10, 16 to the reveal port 46, and vice versa. The X-axis represents the angle of rotation (in degrees).

The exposure openings in the metal drum and the exposure openings 46 are sized according to the volume flow of that particular unit, as are the inlet and outlet conduits 10, 12, 14, 16 and the connecting duct 2.

Those skilled in the art will appreciate that while in the present case the optimum number of modules required to ensure a constant or near constant flow rate is five, in other embodiments of the invention more or fewer modules may be required depending on features such as the geometry of the reveal ports.

The rotary regenerators described herein are capable of transferring temperature from two high pressure fluids or two low pressure fluids, and not just from a low pressure fluid to a high pressure fluid (or vice versa). The invention enables the transfer of thermal energy between two streams of equal pressure without requiring significant changes, other than operations that take into account volume changes. As such, the rotary regenerator may be preceded by a wave rotor, a turbo compressor, or any type of compressor known to those skilled in the art to increase the low pressure inflow to a higher pressure. This has the advantage of reducing the carry over losses.

Turbine cycle system incorporating rotary heat regenerator

One use of the rotary regenerator according to the present invention may be as part of a regenerative turbine cycle.

For example, an extremely simple turbine cycle system of the open Brayton type may incorporate a rotary regenerator, combustor and turbine as described above. The regenerator may be arranged to receive the high temperature fluid exiting the turbine and exchange heat between this fluid and the low temperature high pressure fluid also input to the regenerator. The resulting high pressure, high temperature fluid may then enter the combustor at an elevated temperature, preferably above the auto-ignition temperature of the fuel provided to the combustor, causing the combustor to output fluid at a high pressure or even higher temperature. This fluid can then enter the turbine, which will extract work from the fluid, and the turbine's exhaust can reenter the cycle at the regenerator.

Alternative arrangements may include a compressor for intake air, a rotary regenerator as described herein for transferring heat from the turbine exhaust to the compressed air, a combustor for raising the temperature of the compressed air to the desired turbine intake conditions, and a turbine for extracting work from the high temperature compressed air. The rotary regenerators described herein can receive the high temperature fluid exiting the turbine and exchange this high temperature fluid with the low temperature fluid exiting the compressor. After leaving the regenerator, the cryogenic fluid, now at a high temperature, may enter a high temperature combustor, which is at a temperature above the auto-ignition temperature of the fuel. Thus, the burner would need to be a low emission burner capable of handling high temperature intake air. The turbine will receive high temperature air from the combustor and therefore must be made of suitable materials to handle the high temperature intake air, which may require relatively low rotational speeds based on material specifications. The compressor for the intake air may be a high efficiency compressor that can be driven by the turbine shaft, thus compressing the air at a low rotational speed. Such a cycle including the rotary regenerator described herein can have high efficiency compared to other micro-turbine cycles.

In a particularly preferred embodiment, the present invention discloses an open Brayton gas turbine cycle system in which a high temperature heat recovery device, such as a rotary regenerator as described above, is used to compress and heat the working fluid. The temperature of the compressed working fluid is suitably high above the autoignition temperature of the selected fuel. The high temperature, high pressure working fluid is used as an ultra-low NO working fluid capable of receiving a high temperature oxidant and operating on a plurality of fuels_xDischarge (<3ppm) oxidant in a flameless burner. The combustion working fluid, further raised in temperature by fuel combustion, is then delivered to a high temperature turbine which extracts work from the low pressure exhaust gas before it is used in a rotary recuperator to heat the incoming high pressure working fluid. Work may be converted to electricity through a gearbox and generator. Because the high temperature regenerator recovers heat from the expander exhaust, the air-to-fuel ratio of the system is relatively high (typically greater than 1:20 mass ratio). This is because the heat recovery is very high and therefore a smaller amount of heat input to the process is required during the combustion stage to raise it to the turbine inlet temperature.

In a preferred embodiment along these lines, using the rotary regenerator according to the invention, a flameless burner of the preferred type and a high-temperature turbine of the preferred type allow to mutually align those components in the following way:

the preferred type of high temperature turbine operates at a sufficiently high temperature and low pressure ratio such that the exhaust gas is greater than the auto-ignition temperature of the input fuel. When operating in conjunction with a rotary regenerator, the turbine exhaust gas can be used as a high temperature, low pressure input fluid, and the regenerator can be used to transfer heat from the exhaust gas to the high pressure, low temperature working fluid input to raise the temperature of the working fluid sufficiently above the auto-ignition temperature of the fuel.

High potency (. eta.) according to the invention_eff) The rotary regenerator ensures that the thermal efficiency of the system is high, the optimum pressure ratio of the system is low, and the input temperature of the combustion chamber is above the auto-ignition temperature of the fuel. (the optimal pressure ratio for the system is the regenerator efficiency (. eta.))_eff) Thus a high efficiency regenerator produces a lower optimum pressure ratio; in this case, the ratio may be about 3 to 4. )

The rotary regenerator according to the invention is able to handle high temperatures and to effectively seal between two flows at different pressures.

The preferred type of flameless combustor is capable of handling high temperature incoming oxidant without cooling fluid and is capable of combusting fuel without forming thermal NO_xCO or particulates.

Both the flameless burner and the rotary regenerator of the preferred type according to the invention have a sufficiently low pressure drop so as not to affect the overall thermal efficiency of the system.

Preferred turbine types are: most simply, the preferred turbine type is an external turning wheel turbine that is primarily constructed of ceramic materials. A broad example of such turbines suitable for use in the brayton cycle system described above is provided in GB patent application nos. GB 1804912.2, GB 1709339.4, GB 1702648.5 and GB 1616239.8 and PCT application No. PCT/GB 2017/052850. The entire contents of these applications are incorporated herein by reference, and the claims hereof describe particularly suitable examples of preferred types of turbines.

For example, a preferred type of turbine may be a turbine assembly comprising an axial flow turbine comprising an axially arranged series of rotor segments, wherein each rotor segment comprises an outer ring and rotor blades, and the outer rings of the rotor segments are connected to form a rotating casing, wherein the rotor segments are made of Reaction Bonded Silicon Nitride (RBSN), and the turbine assembly comprises an outer jacket providing structural support for the axial flow turbine, wherein the jacket is made of Dense Silicon Nitride (DSN), wherein the rotor segments are mounted to the inside of the outer jacket, and the jacket and the rotating casing are arranged to rotate together.

In this case, the axial flow turbine may further comprise a series of axially arranged stator segments, wherein each stator segment comprises an inner hub and a plurality of stator blades, and the inner hubs of the stator segments are connected to form a stationary shaft, wherein the stator segments are made of reaction sintered silicon nitride. (in this case, the axial turbine may further comprise at least one shroud contained within the inner hub.) the axial turbine may further comprise at least one shroud included within the outer ring. The jacket may be made of DSN having a density greater than 96% of theoretical density. The jacket may be made of silicon nitride 282 or gas pressure sintered silicon nitride.

Alternatively, a preferred type of turbine may be an axial flow external turning wheel turbine comprising a series of rotor segments arranged axially, wherein each rotor segment comprises an outer ring and rotor blades, and the outer rings of the rotor segments are connected to form a rotating casing, wherein the rotor segments are made of RBSN and have radial dimensions selected such that at the maximum shaft speed of the turbine the radial segments are subjected to stresses of not more than 100 MPa.

Alternatively, a preferred type of turbine may be an axial flow external turning wheel turbine comprising an axially arranged series of rotor segments, wherein each rotor segment comprises an outer ring and rotor blades, and the outer rings of the rotor segments are connected to form a rotating casing, wherein the rotor segments are made of silicon nitride other than RBSN (e.g., gas pressure sintered silicon nitride, or recrystallized silicon nitride).

Alternatively, a preferred type of turbine may be an axial flow external turning wheel turbine comprising a series of rotor segments arranged axially, wherein each rotor segment comprises an outer ring and rotor blades, and the outer rings of the rotor segments are connected to form a rotating casing, wherein the rotor segments are made of RBSN, sintered RBSN, or reflection bonded silicon carbide, and further comprising a reinforcing insert. (inserts may comprise dense ceramic or carbon fibers.)

Alternatively, a preferred type of turbine may be an axial flow external rotor turbine comprising an axially arranged series of rotor segments, wherein each rotor segment comprises an outer ring and rotor blades, and the outer rings of the rotor segments are connected to form a rotating casing, wherein the rotor portions are made of silicon carbide (e.g., sintered silicon carbide, nitride bonded silicon carbide, recrystallized silicon carbide, silicon carbide, or reaction bonded silicon carbide).

Alternatively, a preferred type of turbine may be an axial flow external turning wheel turbine comprising a series of rotor segments arranged axially, wherein each rotor segment comprises an outer ring and rotor blades, and the outer rings of the rotor segments are connected to form a rotating casing, wherein the rotor segments are made of RBSN and the rotor segments are fitted to the inside of an outer jacket, the jacket and the rotating casing are arranged to rotate together, and the jacket is made of silicon carbide or silicon nitride other than dense silicon nitride (e.g. the jacket may be made of gas pressure sintered silicon nitride, recrystallized silicon nitride, RBSN or sintered RBSN).

Preferably, the turbine will be a micro-turbine (i.e., a turbine with a net power output below about 1 MW). In order to operate at high temperatures, both conventional inner and outer runner turbines using metal components need to be made of expensive dissimilar metal alloys and have limited operating temperatures (below about 1100 ℃, which limits the efficiency of power generation) or require complex capillary channels to blow the cooled working fluid over the blades to maintain their mechanical strength. The cooling channels add significant complexity to the design and manufacture of the turbine blades, increase compressor parasitic loads to power the working fluid, and do not easily fit in the very thin small blades typical in micro-turbines (i.e., gas turbines with net power output below about 1 MW). Thus, preferably, the turbine will be constructed from a suitable ceramic or ceramic composite material. Preferably, the turbine will be constructed from a ceramic material, such as Reaction Bonded Silicon Nitride (RBSN), that can handle higher temperatures and that can be machined into complex geometries, such as turbine blades. Alternatively, the ceramic turbine will have a denser ceramic jacket to form a turbine assembly that can withstand greater mechanical strain than a turbine without a dense jacket. Alternatively, the ceramic material used for the micro-turbine may be silicon carbide or another material suitable for mechanical strain and temperature to achieve successful operation of the turbine. Preferably, the micro-turbine will be able to operate at relatively low rotational speeds (< 20,000rpm, as opposed to 100,000 rpm).

Preferred burner types: at its simplest, the preferred burner type is a flameless burner as described in GB patent application No. GB 1805687.9, which is incorporated herein by reference in its entirety, the claims of which describe specific examples of preferred types of burners.

The preferred burner type may then be a flameless burner system comprising a longitudinally extending combustion chamber through which oxidant flows in a longitudinal direction from an inlet to an outlet, and at least one fuel line, wherein at least one fuel injection line in fluid communication with the fuel line extends through a wall of the combustion chamber, wherein each fuel injection line comprises at least one fuel injector for injecting fuel into the combustion chamber.

In some examples, the system includes at least two fuel lines, such that at least one of the at least one fuel injection line extends through a wall of the combustion chamber across an interior of the combustion chamber to an opposing wall of the combustion chamber and through the opposing wall, such that the fuel injection line is in fluid communication with both the first and second fuel lines. In this case, it is possible that each of the at least one fuel lines comprises a plurality of fuel inlet lines extending therefrom along its length, and each fuel inlet line comprises a plurality of fuel injection lines extending therefrom along its length.

In some examples, each of the at least one fuel lines has at least one fuel inlet line extending therefrom in fluid communication with the fuel line, and at least one of the at least one fuel injection lines extends from the fuel inlet line such that the fuel injection line is in fluid communication with the fuel inlet line.

In some examples, the at least one fuel line extends substantially parallel to the combustion chamber in the longitudinal direction. (in this case, it may be that each of the at least one fuel lines extends therefrom at least one fuel inlet line in fluid communication with the fuel line, and at least one of the at least one fuel injection lines extends from at least one of the at least one fuel inlet line such that the fuel injection line is in fluid communication with the fuel inlet line, and wherein the fuel inlet line extends in a direction perpendicular to the direction of the at least one fuel line and the at least one fuel injection line.) in this case, it may be that each of the at least one fuel lines includes a plurality of fuel inlet lines extending therefrom along its length, and each fuel inlet line includes a plurality of fuel injection lines extending therefrom along its length.).

In some alternatives, the fuel injector includes an orifice or a nozzle. In this case, it is possible that the size of the orifice or nozzle is selected to determine the velocity of the fuel injected from the orifice or nozzle under operating conditions.

In some alternatives, the fuel injector is oriented toward the outlet of the combustion chamber.

In some alternatives, the diameter of the at least one fuel line, the diameter of the at least one fuel injection line and, if present, the diameter of the at least one fuel inlet line are selected such that, under operating conditions, the fuel pressure delivered to each injection point is the same.

In certain alternatives, the fuel injectors are evenly distributed within the combustion chamber.

Preferably, the flameless burner will be a multi-fuel burner capable of burning different fuels with only minor changes to the combustion system. (e.g., a combustor with a suitable fuel injection system that enables instant mixing by injecting fuel into multiple locations in the combustor.)

Preferably, the flameless burner will be designed such that the temperature within the burner does not reach higher than the typically required thermal NO_xTemperature is formed, and thus NO is formed in comparison with conventional flame burners_xVery few. For example, the burner injects fuel at a sufficient velocity and co-current with the high velocity oxidant stream to ensure a distributed reaction zone and no flame formation.

Optional functions: alternatively, the fuel is passed through a heat exchanger, well known to those skilled in the art, and then injected into a combustor, which extracts more energy from the low pressure, low temperature output of the rotary regenerator, and then discharged to the atmosphere. Due to the large difference in fuel and exhaust mass flow rates, the heat exchanger need not have a novel design to have high efficiency (η)_eff). This will help to improve the overall thermal efficiency of the cycle.

Alternatively, depending on the fuel source, the fuel will pass through a small fuel compressor and then be injected into the combustor or fuel heat exchanger.

Alternatively, the combustor front in the gas turbine cycle would be a typical gas burner for a start-up system, suitably sized to provide an amount of heat to ensure that the input of the flameless combustor is above auto-ignition before the flameless combustor is turned on. This burner will be able to handle the high pressure inlet oxidant and will bypass the burner during normal operation. Alternatively, the burner would be supplied by an external oxidant and an external fuel source rather than the fuel and oxidant already used in the process. Alternatively, for starting purposes, an auxiliary burner can be used, the exhaust of which is led into the rotary regenerator together with the turbine exhaust, so that starting can be carried out without the need for a high-pressure burner. Typically, such burners will require an induced draft fan on the exhaust side to provide the necessary pressure distribution through the rotary regenerator.

Optionally, the start-up system will include an electric preheater, for example using silicon carbide heating elements. This would be based on the compressed fluid outlet of the rotary regenerator and would increase the temperature of the working fluid prior to entering the flameless combustor to ensure that the temperature is higher than auto-ignition before opening the flameless combustor.

Optionally, the start-up combustion system will be comprised of a cascade of at least one smaller flameless burner and at least one small burner such that the input temperature of the first and smallest burner in the cascade is ensured to be higher than the auto-ignition temperature of the fuel by only the size of the burner, thus reducing the size and complexity of the start-up burner. Alternatively, the start-up burner may be an electric heater for heating air.

Alternatively, when liquid fuel is used, the combustor in the gas turbine cycle will be preceded by a vaporizer. This vaporizes the liquid fuel without the need to mount an atomizing nozzle to the injection system. Preferably, a liquid pump will be in front of the fuel vaporizer to raise the liquid to the necessary pressure.

The use of the rotary regenerator of the present invention ensures a compact overall footprint (as opposed to a recuperator) with high efficiency (η)_eff) Preferably the regenerator will be a rotary regenerator with a geometry that allows for a compact design to achieve high efficiency (η)_eff) Such as the present invention.

Preferably, the rotary regenerator will have radial flow with respect to the axis of rotation, as opposed to flow paths parallel to the axis of rotation, to further increase the compact design to achieve high efficiency (η |)_eff)。

Preferably, the rotary regenerator will utilize a ceramic material as the thermal medium so it can receive a high temperature stream. Preferably, the rotary regenerator will be designed such that the sealing faces are metallic, as opposed to ceramic, so that problems associated with ceramic seals are avoided, such as the present invention.

Preferably, the rotary regenerator will be modular, such that it can be easily increased in size for different flow rates and process variables.

Alternatively, the compressed working fluid may be delivered to the rotary regenerator by a shaft driven compressor. Although any compression device is suitable, a shaft driven compressor will eliminate the inherent cycle efficiency loss (i.e., the efficiency lost in generating electricity and then powering the compressor) caused by an electrically driven compressor.

Alternatively, a shaft driven compressor would be able to produce the required pressure ratio at low rotational speeds.

In embodiments where the compressor is shaft driven, the process will have a self-sustaining speed, where the rotational speed of the turbine and compressor is sufficient so that only fuel is required to rotate the turbine. The rotation of the shaft causes the compressor to produce a particular pressure ratio and have a particular mass flow rate through the compressor based on the resistance of other components in the system. This mass flow and pressure ratio is sufficiently large that the work extracted from the turbine is greater than the work required from the compressor when fuel is added to the combustor.

Alternatively, the shaft driven compressor will be a multi-stage intercooled compressor, such that the compressed air is cooled between the stages to reduce the work done by the compressor. In this case, the compressor will have more than one stage and will be driven by the same shaft. The intercooler may be specific to a compressor or a heat exchanger as known to those skilled in the art.

Alternatively, the shaft driven compressor and turbine would be connected to an electrical generator such that during start-up, the electrical generator could rotate the compressor to provide compressed air into the system before switching at the appropriate self-sustaining speed. Alternatively, the shaft driven compressor will be connected with an auxiliary motor for starting purposes, wherein during normal operation the shaft of the auxiliary motor can be disconnected from the main turbine shaft. Alternatively, when a shaft compressor is used, start-up may be achieved by connecting to an auxiliary compressor in the system, such that, at self-sustaining speed, the associated valve closure may ensure that the auxiliary compressor may be shut down and the shaft compressor may be cycled.

The cycle described herein is very efficient (> 40%), ultra low NOx emissions (<3ppm, no exhaust treatment), low power generation (<1MW), low cost, modular, multi-fuel capable, scalable (i.e. can be designed for any power generation output of 30kW to 1MW, low complexity, therefore easy to control, and has the potential to be compact (e.g. for the automotive industry).

Fig. 12-27 depict abstract schematic diagrams of systems incorporating a rotary regenerator as part of a brayton-type turbine cycle or other turbine cycle. In each case, the system is adapted to a core process, including a compressor for delivering compressed air to three core components, a rotary regenerator, a turbine, and a combustor.

Fig. 12 depicts a system adapted to implement a process in which a cool, low pressure working fluid, in this case ambient air 210, is drawn into the compression stage 202. Compressed air 212, which is typically above ambient temperature but below approximately 150 ℃, exits the compression stage and enters high temperature rotary regenerator 204 where it is heated by turbine exhaust stream 218 and exits as high temperature, high pressure stream 214. This stream 214 must be above the auto-ignition temperature of the fuel 222. High temperature high pressure air 214 for ultra-low NO_x(<3ppm) oxidant in the flameless burner 206. The fuel 222 is combusted, thereby raising the temperature of the combustion air 216 to the necessary turbine inlet temperature of approximately 1250 ℃. Expander 208 as an all ceramic micro-turbine (<1MW) extracts work from the stream, which may be converted to electrical power by a gearbox 226 and generator/motor 228. Where a high speed generator 228 may be used, the gearbox 226 is not required. The turbine also drives the compressor 202 via a shaft 224. Turbine exhaust 218 provides high temperature to rotary regenerator 204, which transfers heat to stream 212 and exhausts to the atmosphere as stream 220. Alternatively, the amount of fuel 222 input into the system may be reduced by providing another source of heat input. The input may be direct (i.e. injection of hot fluid) or indirect (i.e. by heat exchange). The location of the input depends on the type and method of input and may be after compression and before regeneration 212, after regeneration 214, or after entering turbine exhaust 218. This includes, but is not limited to, solar energy, waste heat energy (e.g., steam, hot oil, or exhaust from another process), electrical energy, or non-combustion reactionsThe energy that should be generated.

FIG. 13 depicts a system adapted to implement the process as described in FIG. 12, wherein the exhaust stream 220 is redirected through a heat exchanger 230 that uses the heat of the exhaust stream 220 to preheat the fuel 222. The lower temperature exhaust gas 232 is discharged to the atmosphere and the elevated temperature fuel 234 enters the combustor. This is typically wasted, thus helping to improve the overall thermal efficiency of the process.

FIG. 14 depicts a system adapted to implement the process described in FIG. 12, wherein fuel 222 enters an optional fuel compressor 236 to raise the pressure of fuel 238 prior to injection (if required).

FIG. 15 depicts a system adapted to implement the process as described in FIG. 12, but including an optional fuel burner 242 for start-up purposes. In some embodiments, particularly when the combustor 206 is a flameless combustor, the combustor 206 requires a high temperature air input 214 to ensure auto-ignition of the fuel 222. In this embodiment, it is necessary to raise the temperature of stream 214 during startup before rotary regenerator 204 can absorb sufficient heat from expander exhaust 218. One way to do this is to use a high pressure burner 242. The fuel burner 242 may use the external fuel line 240, the redirected line 222b of the primary fuel line, or both. Depending on the application, the fuel burner will also utilize external oxidant 246 or redirection of the high pressure, high temperature compressed air 214 b. Such a fuel burner would be able to handle high pressures and, in the case of using compressed air 214b as the oxidant, would handle high temperatures entering the oxidant. During start-up, high temperature combustion air 244 may be injected into compressed air stream 214, or directly into combustor 206. Preferably, the burner will be closed during normal operation.

FIG. 16 depicts a system adapted to implement the process as described in FIG. 12, wherein the shaft driven compressors have been replaced with multi-stage intercooled shaft driven compressors 202a, 202b to reduce the work done by the compression stages. Fig. 16 shows a two-stage compressor, a low pressure stage 202a and a high pressure stage 202b, where both compressors are shaft driven. Cold ambient air 210 is drawn into first stage compressor 202a and compresses this stream to a specified pressure. For a two-stage compression system, nominally this is the square root of the final pressure ratio (pressure of compressed air 212 versus ambient air 210). The semi-compressed air 248 enters a heat exchanger 252 which extracts heat through a cooling fluid 254, and the semi-compressed air, now at a lower temperature 250, enters the second compressor stage 202 b. The cooling fluid 254 exits the heat exchanger at a higher temperature 256. Second stage compressor 202b compresses semi-compressed air 250 to a desired working fluid pressure 212. Alternatively, the multi-stage compressor may be driven by an external shaft or motor, multi-rotor, and may include more than two stages.

FIG. 17 depicts a system adapted to implement the process described in FIG. 12, including an optional fuel vaporizer 260 that converts liquid fuel 258b to gaseous fuel 258c, which is then injected into the combustor. The liquid fuel 258a will be raised in pressure by a pump or liquid compressor 259. This approach would be used if the process required the use of liquid fuel without changing the multi-fuel burner 206 to use liquid fuel.

Fig. 18 depicts a system adapted to implement the process as described in fig. 12, but may be used with a non-shaft driven compressor 262. Typically, a compressor like this will be electrically driven or driven by a diesel generator. This will allow all of the turbine power to be used to generate electricity via generator 228.

Fig. 19 depicts a system adapted to implement the process as described in fig. 12, including an optional auxiliary motor 264 and shaft 266 for rotating compressor 202 for start-up purposes only. During normal operation, the start shaft 266 may be disengaged from the main shaft 224. The motor 264 may contain its own generator or may be attached to the mains electricity. This allows the compressor 202 and turbine 208 to be decoupled during start-up.

Fig. 20 depicts a system adapted to implement the process as described in fig. 12, but with an electrical air preheater 268 for heating the high pressure air 214c during startup. This would raise the temperature of the compressed air 214 by external electrical input, for example, using a silicon carbide heating element embedded in the fluid conduit, so that the temperature of the heater outlet 270 is a temperature high enough for start-up purposes. The electric air preheater 268 may preheat the redirected air of the high pressure, high temperature compressed air 214c or may be located within the duct of the main flow 214. This provides a way to raise the temperature of the compressed air 214 above the autoignition temperature required for use with a flameless burner.

Fig. 21 depicts a system designed to implement a process as described in fig. 12 using at least one cascade start-up burner 272a and optionally a cascade flameless burner 272 b. In one embodiment, during start-up, the compressed air stream 214 is split into a larger fluid stream 214e and a correspondingly smaller fluid stream 214 d. The smaller fluid stream 214c enters the burner or burner 272a, wherein the fuel 240b is combusted, whereupon the temperature of the fluid rises (270) and mixes with the larger fluid stream 214e before entering the main burner 206. This causes a further increase in the temperature of the compressed air 214 through heat recovery into the rotary regenerator 204. The fluid temperature is partially increased until a portion of the fluid 214d is sufficiently increased such that flameless combustion may occur in the secondary start-up burner 272 b. Preferably, the secondary start-up burner will also have another input fluid 214f such that the total mass flow of the elevated temperature combustion air 270 is greater than when the secondary burner is not used. Such a cascade system will continue until the temperature of the combusted fluid 270 mixed with the working fluid 214g is sufficiently high, i.e., above the necessary auto-ignition temperature, such that flameless combustion can occur in the main combustor 206. Optionally, the system will be bypassed 214 during normal operation. In such a system, the number and size of the cascade burners and flameless burners 272a and 272b depends on the application, as does the division of the streams 214d, 214e, 214f, 214 g.

FIG. 22 depicts a system adapted to implement the process described in FIG. 12 with an optional fuel burner 274 for startup purposes. The secondary burner 214 preferably combusts the fuel source 240 at ambient pressure or uses a retro-fitted existing fuel source 222b with external oxidant 246. The secondary burner exhaust 276 enters the rotary regenerator and provides the necessary heat for heat exchange to raise the temperature of the compressed air 214 before entering the combustor 206. This may continue until the mass flow and temperature of the turbine exhaust 218 are sufficient to provide the necessary heat to the compressed air 214. This provides a way to raise the temperature of the compressed air 214 above the autoignition temperature required for use with a flameless burner. Such a system may require placing an induced draft fan on the process exhaust 220.

Fig. 23 depicts a system adapted to implement the process as described in fig. 12 with an auxiliary compressor for start-up purposes during use of the shaft driven compressor 202. In some embodiments, when using a shaft driven compressor 202, the generator will not be suitable for starting the compression process. In these embodiments, boost would need to be provided from the auxiliary compressor 278. During start-up, the auxiliary compressor 278 compresses the cold working fluid 282 into the high pressure fluid 280. This fluid is then passed through a rotary regenerator 204 to raise the temperature. The auxiliary compressor will continue to be utilized until the shaft driven compressor 202 rotates fast enough to provide the necessary mass flow of compressed fluid 212 so that the process is at a self-sustaining rotational speed. Alternatively, the outlet of the shaft compressor may be vented to atmosphere 286 until a switch of compressors occurs.

Fig. 24 depicts a system adapted to implement the process as described in fig. 12 with the addition of an optional exhaust heat recovery device 284. The device may be a heat exchanger of a district heating system, or may be a refrigeration system or other device that extracts energy from an exhaust stream.

FIG. 25 depicts one embodiment of a system designed to implement the process described in FIG. 12, wherein the core process is adapted for static power generation. This includes multi-stage shaft driven compressors 202a, 202b with inter-stage cooling 252. Fuel 222 is compressed by fuel compressor 236 before being heated by fuel heat exchanger 230 using process exhaust 220, which is then directed to heat exchanger 284 for use in applications such as zone heating. The auxiliary compressor 278 and the auxiliary burner 274 are used for start-up purposes. This is just one embodiment for a static power generation application, and there are other embodiments of various auxiliary systems and core processes for the same application.

FIG. 26 depicts one embodiment designed to implement the process described in FIG. 12, where the core process is adapted for automobile range extension. This includes an electric heater 268 for heating the compressed air 214c during start-up to bring the temperature of the heated compressed air 270 high enough to start the flameless multi-fuel burner 206. The liquid fuel 258a is pumped by a liquid pump 259 to a vaporizer 260, which vaporizes the pressurized liquid 258b for injection as the gaseous fuel 258b into the combustor 206. In this embodiment, the high speed generator 228 is used to convert turbine work to potential energy, thus eliminating the need for a gearbox. Additionally, the generator 228 will be used to drive the shaft for starting purposes. This is just one embodiment for automotive range-extending applications, and there are other embodiments of various auxiliary systems and core processes for the same application.

FIG. 27 depicts a startup process for the system depicted in FIG. 25. Initially, the auxiliary compressor 278 is started to provide compressed air to the system. The auxiliary burner 274 is then activated to heat the rotary regenerator 204 to the stream 214. The turbine 208 and the shaft compressors 202a, 202b will begin to spin. As the shaft compressors 202a, 202b spin faster, the mass flow through them increases and the compressed air from the shaft compressors is discharged to the atmosphere. As the rotary regenerator heats up, the temperature of the compressed air stream 214 will increase. When the flameless burner inlet 214 is high enough to perform flameless combustion of the fuel 234, the secondary burner 274 is turned off and the flameless burner 206 is turned on. The more fuel input, the greater the energy input into the system and the higher the temperature in the turbine, so that when the fuel input into the system increases, the speed of the turbine and thus the speed of the shaft will increase. The fuel input 222 is increased until the speed of the shaft 224 is sufficiently high that the mass flow through the compressors 202a, 202b is sufficient for the system to be self-sustaining without additional compressed air input from the auxiliary compressor 278. At this point, compressed air from the shaft compressors 202a, 202b directed to the rotary regenerator 204 and the auxiliary compressor 278 is turned off, which may be accomplished by, for example, a three-way valve. The fuel input 222 and load from the generator 228 are manipulated until the speed of the shaft 224 reaches the design point.

Regenerator and use of a turbine cycle incorporating a regenerator

Regenerators and turbine cycles incorporating regenerators can be used in a range of different applications, particularly where particular advantages are required. For example, an automobile may incorporate such a regenerator and/or turbine cycle; such a turbine cycle can also be used for static power generation.