Patent Grant
6866092
The general technical field of the present invention pertains to systems that include one or more fluid circuits for transferring heat from one or more heat sources to one or more heat sinks with a heat-transfer fluid circulating around the one or more fluid circuits; a heat sink—to which heat is released by the heat-transfer fluid—having, at an instant in time, a maximum temperature below the maximum temperature of the heat source from which the released heat is absorbed at that instant in time. Such heat-transfer systems—which by the foregoing description exclude heat pumps—can be grouped into two general categories:
The specific technical field of the present invention pertains to two-phase heat-transfer systems. Such systems include, in addition to a heat-transfer fluid, hereinafter named a refrigerant, an evaporator and a condenser. The evaporator has one or more refrigerant passages in which the refrigerant absorbs heat from a heat source, at least in part, by changing from its liquid to its vapor phase. The condenser has one or more refrigerant passages in which the refrigerant releases heat to a heat sink, at least in part, by changing back from its vapor phase to its liquid phase at pressures which, at an instant in time, do not exceed the lowest pressure at which the refrigerant changes phase in the one or more evaporator refrigerant passages at that instant in time. Two-phase heat-transfer systems also include means for transferring refrigerant vapor from the evaporator refrigerant passages to the condenser refrigerant passages, and means for transferring liquid refrigerant from the condenser refrigerant passages to the evaporator refrigerant passages. The two just-cited means, and the evaporator and condenser refrigerant passages, form a circuit around which the refrigerant circulates while the refrigerant alternates between its liquid and its vapor phases. I shall refer to such a circuit as a ‘refrigerant principal circuit’.
Two-phase heat-transfer systems may have one or more refrigerant principal circuits with the same or different kinds of refrigerant, and each of these refrigerant principal circuits may have associated with it one or more refrigerant auxiliary circuits in the sense that they share a refrigerant-circuit segment with each refrigerant principal circuit. Refrigerant auxiliary circuits differ from refrigerant principal circuits in that
The invention disclosed in the present document pertains exclusively to airtight two-phase heat-transfer systems, namely to two-phase heat-transfer systems which, in the absence of a failure, do not ingest air while they are active or while they are inactive.
Many potentially important applications exist for two-phase heat-transfer systems whose refrigerant has, while they are not operating, saturated-vapor pressures substantially below ambient atmospheric pressure. However, prior-art embodiments of such two-phase heat-transfer systems have often been unable to compete successfully with single-phase heat-transfer systems. This is in particular true in the case of internal-combustion-engine prior-art two-phase cooling systems which have so far never been mass-produced, and have been used only in a few concept-demonstration vehicles and in a few ground installations.
I assert that a principal reason for the fact recited in the immediately preceding sentence is that most prior-art internal-combustion-engine two-phase cooling systems ingest air each time they are deactivated and their refrigerant approaches ambient air temperatures. I also assert that the prior-art describes no generally useful techniques for eliminating air ingestion from internal-combustion-engine cooling systems without
The handicaps of prior-art internal-combustion-engine airtight two-phase cooling systems recited above under (a) and (b) apply also to many other airtight two-phase heat-transfer systems, whose refrigerant has, while they are not operating, saturated-vapor pressures substantially below ambient atmospheric pressure. Nevertheless, the prior art discloses no techniques for maintaining the internal pressure of inactive airtight two-phase heat-transfer systems above their refrigerant saturated-vapor pressure without imposing at least one of the constraints recited above under (a) and (b).
In addition to the handicaps recited above under (a) and (b), prior-art airtight two-phase heat-transfer systems in general, and internal-combustion-engine airtight two-phase cooling a systems in particular, have several additional major handicaps which must be eliminated before airtight two-phase heat-transfer systems can realize their full potential. The nature of those additional handicaps will become apparent whilst reading this DESCRIPTION.
Non-airtight two-phase heat-transfer systems do not have some of the handicaps of prior-art airtight two-phase heat-transfer systems. However, the air ingested by non-airtight systems has often been a sufficient handicap for them to be unable to compete successfully with single-phase heat-transfer systems. A prominent example where this has happened are steam building-heating systems which have been superseded by hot-water building-heating systems primarily because of the unacceptable rate of corrosion caused by air ingestion.
1. General Remarks
Terms between single quotation marks are defined in this DESCRIPTION. Some of those terms are defined in section III,A,2 under the heading PRELIMINARY DEFINITIONS, and others are defined elsewhere in this DESCRIPTION.
2. Preliminary Definitions
Certain terms used in describing and claiming the invention disclosed in the present document shall have the following meaning:
1. The term ‘refrigerant’ is used to denote a fluid employed—under at least some operating conditions—to absorb heat, at least in part by changing from a liquid to a vapor and to release the absorbed heat at least in part by changing from a vapor back to a liquid. A refrigerant is said to ‘absorb latent heat’ when the refrigerant changes from a liquid to a vapor and to ‘release latent heat’ when the refrigerant changes from a vapor to a liquid; and a refrigerant is said to ‘absorb sensible heat’ when the refrigerant's (sensible) temperature rises while the refrigerant remains in one of the refrigerant's two phases (namely while the refrigerant remains in either its liquid phase or in its vapor phase) and to ‘release sensible heat’ when the refrigerant's (sensible) temperature falls while the refrigerant remains in one of the refrigerant's two phases. I intend the last four terms in quotation marks to apply to refrigerants which are a non-azeotropic mixture of single-component fluids as well as to refrigerants which are single-component fluids or an azeotropic mixture of single-component fluids. I shall often herein refer for brevity to fluids which are a non-azeotropic mixture of single-component fluids as ‘non-azeotropic fluids’. I shall also often refer herein to single-component fluids, and to fluids which are an azeotropic mixture of single-component fluids, collectively as ‘azeotropic-like fluids’, where the word ‘like’ indicates that, in contrast to non-azeotropic fluids, both single-component and azeotropic fluids boil at only one temperature while subjected to a given constant pressure. It follows from my definition of the term ‘refrigerant’ that the term ‘refrigerant’ is used herein to denote the function of a heat-transfer fluid and not the nature of a heat-transfer fluid; and is not used herein to restrict the kinds of heat-transfer fluid employed in the systems of the present invention to a particular class of fluids such as fluids more volatile than H2O, and especially not to exclude water as for example in U.S. Pat. No. 4,120,289 (Bottum), 17 Oct. 1978, and U.S. Pat. No. 4,220,138 (Bottum), 02 Sep. 1980. Liquid refrigerant is said to ‘evaporate’ when it is changing from a liquid to a vapor, and refrigerant vapor is said to ‘condense’ when it is changing from a vapor to a liquid. And refrigerant is said to absorb heat by evaporation when refrigerant absorbs heat while changing from a liquid to a vapor, and to release eat by condensation when refrigerant releases heat while changing from a vapor to a liquid.
2. The term ‘evaporator’ denotes means for transmitting heat from a heat source to a refrigerant and for evaporating liquid refrigerant; the evaporator having one or more surfaces which are the bounds of one or more enclosed spaces, named by me refrigerant passages, where refrigerant absorbs heat from the heat source at least in part by changing from a liquid to a vapor.
3. The term ‘preheater’ denotes means for transmitting heat from a heat source to a refrigerant and for heating, namely increasing the (sensible) temperature of, liquid refrigerant; the preheater having one or more surfaces which are the bounds of one or more enclosed spaces, named by me refrigerant passages, where refrigerant absorbs heat from the heat source solely while the refrigerant is in the refrigerant's liquid phase.
4. The term ‘superheater’ denotes means for transmitting heat from a heat source to a refrigerant and for heating, namely increasing the (sensible) temperature of, refrigerant vapor; the superheater having one or more surfaces which are the bounds of one or more enclosed spaces, named by me refrigerant passages, where refrigerant absorbs heat from the heat source solely while the refrigerant is in the refrigerant's vapor phase.
5. The term ‘condenser’ denotes means for transmitting heat from a refrigerant to a heat sink and for condensing refrigerant vapor; the condenser having one or more surfaces which are the bounds of one or more enclosed spaces, named by me refrigerant passages, where refrigerant releases heat to the heat sink at least in part by changing from a vapor to a liquid.
6. The term ‘subcooler’ denotes means for transmitting heat from a refrigerant to a heat sink and for cooling, namely decreasing the (sensible) temperature of, liquid refrigerant; the subcooler having one or more surfaces which are the bounds of one or more enclosed spaces, named by me refrigerant passages, where refrigerant releases heat to the heat sink solely while the refrigerant is in the refrigerant's liquid phase.
7. The term ‘desuperheater’ denotes means for transmitting heat from a refrigerant to a heat sink and for cooling, namely decreasing the (sensible) temperature of, refrigerant vapor; the desuperheater having one or more surfaces which are the bounds of one or more enclosed spaces, named by me refrigerant passages, where refrigerant releases heat to the heat sink solely while the refrigerant is in the refrigerant's vapor phase.
8. The term ‘hot heat exchanger’ denotes a member of the family consisting of all evaporators, preheaters, and superheaters.
9. The term ‘cold heat exchanger’ denotes a member of the family consisting of all condensers, subcoolers, and desuperheaters.
10. The term ‘heat exchanger’ denotes any heat exchanger; including any member of the family consisting of all hot heat exchangers, and all cold heat exchangers, as defined in definitions (8) and (9). I note that no restriction is imposed on the nature of the heat source of the hot heat exchangers defined under (2), (3), (4), and (8) in this section III, A, or on the nature of the heat sink of the cold heat exchangers, defined under (5), (6), (7), and (9), in this selfsame section; and it therefore follows—in contrast to the definition of the term ‘heat exchanger’ found in the art—that the heat exchangers cited hereinafter in this DESCRIPTION may—except where otherwise stated—include heat exchangers for transmitting heat from a solid to a fluid, and from a fluid to a solid, and are not restricted to heat exchangers for transmitting heat from a fluid to another fluid. A heat exchanger has a fluid inlet, and in particular a refrigerant inlet, consisting of a set of one or more inlet ports and a fluid outlet, and in particular a refrigerant outlet, consisting of a set of one or more outlet ports.
11. The term ‘principal heat exchanger’ denotes a heat exchanger whose purpose is to transfer heat from a heat source of a two-phase heat-transfer system to one of the system's one or more refrigerants, or to transfer heat from a refrigerant of a two-phase heat-transfer system to one of the system's one or more heat sinks. A principal heat exchanger may be a hot heat exchanger, and in particular an evaporator, a preheater, or a superheater; or it may be a cold heat exchanger, and in particular a condenser, a subcooler, or a desuperheater. In this DESCRIPTION and in the CLAIMS, the terms ‘evaporator’, ‘preheater’, ‘superheater’, ‘condenser’, ‘subcooler’, and ‘desuperheater’, refer, for brevity, to principal heat exchangers, except where the qualifier ‘accessory’ is explicitly stated or obviously implied.
12. The term ‘accessory heat exchanger’ in general, and the terms ‘accessory evaporator’, ‘accessory condenser’, ‘accessory subcooler’, etc. in particular, denote heat exchangers used for accessory functions. Examples of such accessory heat exchangers are the accessory condensers used to assist in removing refrigerant vapor from a refrigerant-vapor and non-condensable gas mixture, and which, to this end, transfer heat from the mixture to a heat sink, and accessory heat exchangers used to transfer heat from an inert gas to a heat sink and from a heat source to an inert gas.
13. The term ‘separating surfaces’ denotes any set of surfaces (including surfaces forming a centrifugal separator) for separating the liquid and vapor phases of wet refrigerant vapor flowing over the set of surfaces. Separating surfaces may be an integral part of the refrigerant passages of an evaporator.
14. The term ‘separator’ denotes means for separating the liquid and vapor phases of wet 2 refrigerant vapor; the separator having a vessel, named ‘separator vessel’, for storing, whenever appropriate, liquid refrigerant. A separator may include separating surfaces (often referred to as baffles) to help separate the liquid and the vapor phases of wet refrigerant vapor in the separator.
15. The term ‘2-port separator’ denotes a separator having a first set of one or more ports through which usually wet refrigerant vapor enters the separator and liquid refrigerant exits the separator; and a separate second set of one or more ports through which refrigerant vapor exits the separator, the refrigerant vapor exiting the separator usually being drier than the refrigerant vapor entering the separator.
16. The term ‘3-port separator’ denotes a separator having a first set of one or more ports through which usually wet refrigerant vapor enters the separator; a separate second set of one or more ports through which refrigerant vapor exits the separator, the refrigerant vapor exiting the separator usually being drier than the refrigerant vapor entering the separator; and a separate third set of one or more ports through which liquid refrigerant usually exits the separator but may also enter the separator.
17. The term ‘3*-port separator’ denotes a separator having a first set of one or more ports through which usually wet refrigerant vapor enters the separator and through which liquid refrigerant exits the separator; a separate second set of one or more ports through which refrigerant vapor exits the separator, the refrigerant vapor exiting the separator usually being drier than the refrigerant vapor entering the separator; and a separate third set of one or more ports through which liquid refrigerant enters the separator.
18. The term ‘4-port separator’ denotes a separator having a first set of one or more ports through which usually wet refrigerant vapor enters the separator; a separate second set of one or more ports through which refrigerant vapor exits the separator, the refrigerant vapor exiting the separator usually being drier than the refrigerant vapor entering the separator; a separate third set of one or more ports through which liquid refrigerant exits the separator; and a separate fourth set of one or more ports through which liquid refrigerant enters the separator.
19. The term ‘separating assembly’ denotes means for separating the liquid and vapor phases of wet refrigerant vapor that does not include a vessel for storing liquid refrigerant. A separating assembly may be an integral part of a separator.
20. The term ‘2-port separating assembly’ denotes a separating assembly having a first set of one or more ports through which usually wet refrigerant vapor enters the assembly and liquid refrigerant exits the assembly, and a separate second set of one or more ports through which refrigerant vapor exits the assembly, the refrigerant vapor exiting the assembly usually being drier than the refrigerant vapor entering the assembly. A 2-port separating assembly almost always includes separating surfaces.
21. The term ‘3-port separating assembly’ denotes a separating assembly having a first set of one or more ports through which usually wet refrigerant vapor enters the assembly; a separate second set of one or more ports through which refrigerant vapor exits the assembly, the refrigerant vapor exiting the assembly usually being drier than the refrigerant vapor entering the assembly; and a separate third set of one or more ports through which liquid refrigerant exits the assembly. A 3-port separating assembly may include no separating surfaces other than the internal surfaces of the assembly's refrigerant passages, and may, for example, merely be a shallow V-tube having the first set of one or more ports essentially at the top of one of the two arms of the vee, the second set of one or more ports essentially at the top of the other arm of the vee, and the third set of one or more a ports essentially at the bottom of the vee.
22. The term ‘separating device’ in this DESCRIPTION, and synonymously the term ‘separating means’ in the CLAIMS, denotes means for separating the liquid and vapor phases of wet refrigerant vapor. A separating device or means may be (1) a separator which includes a distinguishable separating assembly, (2) a separator which has no distinguishable separating assembly, or (3) a separating assembly.
23. The term ‘refrigerant-circuit’ denotes a fluid circuit around which, whenever appropriate, a refrigerant circulates.
24. The term ‘refrigerant line’ denotes a conduit for transferring refrigerant between components such as heat exchangers, separators, separating assemblies, refrigerant valves, refrigerant pumps, and receivers (see definition 41).
25. The term ‘refrigerant-circuit segment’ denotes a part of a refrigerant circuit. A refrigerant-circuit segment may include several refrigerant lines connected in parallel, or the refrigerant passages of several similar, or several dissimilar, components, connected in parallel. These components include refrigerant valves (see definition 29), heat exchangers, separators, refrigerant pumps (see definition 33), and receivers (see definition 41).
26. The term ‘refrigerant space’ denotes an enclosed space containing essentially only refrigerant. The term ‘refrigerant space’ subsumes the space inside a refrigerant line, and the space inside a refrigerant passage of a heat exchanger, a refrigerant pump, or a refrigerant valve.
27. The term ‘refrigerant enclosure’ denotes a structure delineating the bounds of a set of one or more fluidly-connected refrigerant spaces containing in essence only refrigerant.
28. The term ‘valve’ denotes a device by which the flow of a fluid, in its liquid or in its vapor phase, can be started, stopped, or regulated, by any known means capable of exerting a force on the particular fluid in the valve's one or more fluid passages. Examples of such a force include a mechanical, a magneto-hydrodynamic, an electro-dynamic, an electro-osmotic, and a capillary, force. Where the force is a mechanical force, the flow of the fluid through the valve's one or more fluid passages is started, stopped, or regulated, by a movable mechanical part which respectively opens, shuts, or partially obstructs, the valve's one or more fluid passages. The term ‘valve’, where the force is a mechanical force, includes an actuator for controlling the position of the movable mechanical part.
29. The term ‘refrigerant valve’ denotes a valve where the fluid whose flow is controlled by the valve is a refrigerant in its liquid or in its vapor phase, and where the one or more fluid passages are refrigerant passages.
30. The term ‘pump’ denotes a device for generating an increase in fluid pressure causing a fluid to flow in a desired direction. A pump has one or more fluid passages through which the fluid flows while the pump is active. A pump may be driven by any known means capable of exerting a force on the particular fluid in the pump's one or more fluid passages. Examples of such a force include a mechanical, a pneumatic, an hydraulic, a magneto-hydrodynamic, an electro-dynamic, an electro-osmotic, and a capillary, force. Where (1) means used to drive a pump is used exclusively to drive the pump and the pump is not driven by any other means, the term ‘pump’ includes the pump-driving means; and where (2) means used to drive a pump is also used for another purpose, or is merely an alternative means for driving a pump, the term ‘pump’ excludes the one or more pump-driving means. An example of the case recited under (1) in the present definition is an electric motor used to drive a pump where the electric motor is used exclusively to drive the pump; an example of the former of the two cases recited under (2) in the present definition is an engine used to drive a vehicle which is also used to drive a pump; and an example of the latter of the two cases recited under (2) in the present definition is a pump driven by an engine used to drive a vehicle and alternatively by an electric motor.
31. The term ‘inherent capacity’, where the subject is a pump, denotes the fluid mass-flow rate induced by the pump, through the pump's one or more fluid passages under the action of the device or means driving the pump, for a given fluid pressure at the point where a fluid enters the pump's one or more fluid passages and for a given fluid-pressure rise in the pump's one or more fluid passages. The inherent capacity of a pump may, for a given fluid density, be essentially constant, or the inherent capacity of a pump may, for a given fluid density, be varied by the device driving the pump. In the particular case where the pump exerts a mechanical force on the fluid flowing through its one or more fluid passages, the pump's inherent capacity can be varied, for example, by one or more of the three techniques known as pump-speed control, pump-vane control, and on-off control. The fluid mass-flow rate delivered, under the earlier-cited fluid-pressure conditions in this definition, at a given point by a pump with a constant inherent capacity, or with a variable inherent capacity, may be modified by using a flow-control valve in series with the pump, or a flow-control valve in parallel with the pump. I shall refer to the latter valve as a ‘pump-recirculation valve’. (Pump-recirculation valves may be an integral part of a pump.)
32. The term ‘effective capacity’ where the subject is a pump, denotes the fluid mass-flow rate delivered by a pump at a given fluid-circuit segment cross-section after the inherent capacity of the pump has been modified by the pump's recirculation valve or by a flow-control valve upstream from the given segment. The flow-control valve is, depending on the type of pump, located upstream from or downstream from the pump.
33. The term ‘refrigerant pump’ denotes a pump causing liquid refrigerant to flow through a refrigerant-circuit segment in a desired direction. A refrigerant pump has one or more refrigerant passages through which liquid refrigerant flows while the refrigerant pump is active.
34. The term ‘refrigerant principal circuit’ denotes a refrigerant circuit which includes the one or more refrigerant passages of an evaporator, and the one or more refrigerant passages of a condenser, (where the evaporator and the condenser are principal heat exchangers).
35. The term ‘refrigerant auxiliary circuit’ denotes a refrigerant circuit, other than a refrigerant principal circuit. A refrigerant auxiliary circuit may include the one or more refrigerant passages of an evaporator and no condenser refrigerant passages; or the one or more refrigerant passages of a condenser and no evaporator refrigerant passages; or no evaporator or condenser refrigerant passages. Refrigerant circulating around an auxiliary refrigerant circuit remains in the same fluid phase during a circulation cycle; whereas refrigerant circulating around a refrigerant principal circuit changes—during each circulation cycle—at least in part, under most operating conditions, from the refrigerant's liquid phase to the refrigerant's vapor phase and from the refrigerant's vapor phase back to the refrigerant's liquid phase.
36. The term ‘forced refrigerant-circulation principal circuit’, or more briefly, ‘FRC principal circuit’, denotes a refrigerant principal circuit around which a refrigerant circulates continuously or intermittently, primarily under the forced action of a refrigerant pump, while the refrigerant is transferring heat from a heat source to a heat sink.
37. The term ‘natural refrigerant-circulation principal circuit’, or more briefly, ‘NRC principal circuit’, denotes a refrigerant auxiliary circuit around which a refrigerant circulates usually continuously, solely under the combined action of gravity and of the heat supplied by a heat source, while the refrigerant is transferring heat from the heat source to a heat sink.
38. The term ‘forced refrigerant-circulation auxiliary circuit’, or more briefly, ‘FRC auxiliary circuit’, denotes a refrigerant circuit around which a refrigerant circulates continuously or intermittently, primarily under the forced action of a pump, while the refrigerant is transferring heat from a heat source to a heat sink.
39. The term ‘natural refrigerant-circulation auxiliary circuit’, or more briefly, ‘NRC auxiliary circuit’, denotes a refrigerant auxiliary circuit around which a refrigerant circulates usually continuously, solely under the combined action of gravity and of heat supplied by a heat source, while the refrigerant is transferring heat from the heat source to a heat sink.
40. The term ‘refrigerant principal configuration’, or more briefly ‘principal configuration’, denotes a material structure for transferring heat from one or more heat sources to one or more heat sinks; the configuration comprising
41. The term ‘liquid-refrigerant receiver’, or more briefly ‘receiver’, denotes a vessel for storing, whenever appropriate, liquid refrigerant, provided the vessel is not a part of a separator.
42. The term ‘1-port receiver’, or equivalently ‘surge-type receiver’, denotes a receiver having a single set of one or more ports through which liquid refrigerant enters and exits the receiver.
43. The term ‘2-port receiver’, or equivalently ‘feed-through receiver’, denotes a receiver having a first set of one or more ports through which refrigerant condensate enters the receiver, and a second set of one or more ports through which liquid refrigerant, stored in the receiver, exits the receiver.
44. The term ‘refrigerant-vapor transfer means’ denotes means, including one or more distinguishable refrigerant spaces, for transferring refrigerant vapor exiting a principal configuration's one or more evaporator refrigerant passages to the principal configuration's one or more condenser refrigerant passages. In particular, the term ‘refrigerant-vapor transfer means’ may, for example, (1) merely consist of a single refrigerant line, not excluding an essentially zero-length refrigerant line such as a port; or (2) may include space inside a separating device occupied by refrigerant vapor, one or more refrigerant lines for transferring refrigerant vapor exiting the one or more evaporator refrigerant passages to the separating device, and one or more refrigerant lines for transferring refrigerant vapor from the separating device to the one or more condenser refrigerant passages; the one or more refrigerant lines not excluding refrigerant lines forming a manifold.
45. The term ‘liquid-refrigerant principal transfer means’ denotes means, including one or more distinguishable refrigerant spaces, for transferring liquid refrigerant exiting a principal configuration's one or more condenser refrigerant passages to the principal configuration's one or more evaporator refrigerant spaces. In particular, the term ‘liquid-refrigerant principal transfer means’ may, for example, (1) merely consist of a single refrigerant line; (2) may include a refrigerant line and the one or more refrigerant passages of a refrigerant pump and/or the one or more refrigerant passages of a refrigerant valve; or (3) may include a receiver not excluding a 1-port receiver, the one or more refrigerant passages of a refrigerant pump, a refrigerant line for transferring liquid refrigerant from the receiver to the one or more refrigerant-pump refrigerant passages, one or more refrigerant lines for transferring liquid refrigerant exiting one or more condenser refrigerant passages to the receiver, and one or more refrigerant passages for transferring liquid refrigerant from the one or more refrigerant-pump refrigerant passages to the one or more evaporator refrigerant passages; the last-cited one or more refrigerant lines not excluding & refrigerant lines forming a manifold.
46. The term ‘liquid-refrigerant auxiliary transfer means’ denotes means for transferring liquid refrigerant, the means including one or more distinguishable refrigerant spaces which (1) are a part of a refrigerant principal configuration, but which (2) are not a part of a liquid-refrigerant principal transfer means. An important example of a liquid-refrigerant auxiliary transfer means is means for transferring liquid refrigerant from the separating device of a principal configuration to one or more points of the configuration's refrigerant principal circuit. Such a liquid-refrigerant auxiliary transfer means may, for instance, consist of (1) merely a single refrigerant line; (2) several refrigerant lines forming a manifold; or (3) the one or more refrigerant passages of an evaporator-overfeed pump, a refrigerant line for transferring liquid refrigerant from the separating device to the one or more refrigerant passages of the evaporator-overfeed pump, and one or more refrigerant lines for transferring liquid refrigerant from the one or more refrigerant passages of the evaporator-overfeed pump to one or more evaporator refrigerant passages, the one or more refrigerant lines not excluding refrigerant lines forming a manifold.
47. The term ‘type 1 evaporator refrigerant auxiliary circuit’ denotes, in a principal configuration having several refrigerant circuits, a refrigerant auxiliary circuit which includes the one or more refrigerant passages of the configuration's evaporator; and which excludes
48. The term ‘type 2 evaporator refrigerant auxiliary circuit’ denotes, in a principal configuration with several refrigerant circuits, a refrigerant auxiliary circuit which includes the one or more refrigerant passages of the configuration's evaporator and the one or more refrigerant-pump refrigerant passages which are a part of the configuration's refrigerant principal circuit; and which excludes the one or more refrigerant passages of the configuration's condenser.
49. The term ‘evaporator refrigerant auxiliary circuit’ denotes a member of the family of all refrigerant auxiliary circuits consisting of type 1 evaporator refrigerant auxiliary circuits and type 2 evaporator refrigerant auxiliary circuits.
50. The term ‘type 1 separator’ denotes all 3-port and 4-port separators having two sets of ports which are a part of a type 1 evaporator refrigerant auxiliary circuit.
51. The term ‘type 2 separator’ denotes all 3-port and 4-port separators having two sets of ports which are a part of a type 2 evaporator refrigerant auxiliary circuit.
52. The term ‘type 1′ separator’ denotes all 2-port and 3*-port separators having no set of ports which is a part of an evaporator refrigerant auxiliary circuit.
53. The term ‘type 1 separating assembly’ denotes a 3-port separating assembly having two sets of ports which are a part of a type 1 evaporator refrigerant auxiliary circuit.
54. The term ‘type 2 separating assembly’ denotes a 3-port separating assembly having two CD sets of ports which are a part of a type 2 evaporator refrigerant auxiliary circuit.
55. The term ‘type 1′ separating assembly’ denotes a 2-port separating assembly having no set of ports which is a part of an evaporator refrigerant auxiliary circuit.
56. The term ‘type 1 separating device or means’ denotes a type 1 separator or a type 1 separating assembly.
57. The term ‘type 2 separating device or means’ denotes a type 2 separator or a type 2 separating assembly.
58. The term ‘type 1′ separating device or means’ denotes a type 1′ separator or a type 1′ separating assembly.
59. The term ‘subcooler refrigerant auxiliary circuit’ denotes a refrigerant auxiliary circuit which includes (1) the one or more refrigerant passages of a subcooler of a principal configuration, and (2) the one or more refrigerant passages of a refrigerant pump of the configuration; and which excludes (1) the one or more refrigerant passages of the configuration's evaporator, and (2) the one or more refrigerant passages of the configuration's condenser.
60. The term ‘condensate-return pump’, or more briefly ‘CR pump’, denotes a refrigerant pump having one or more refrigerant passages which are a part of a refrigerant principal circuit and of no other refrigerant circuit.
61. The term ‘evaporator-overfeed pump’, or more briefly ‘EO pump’, denotes a refrigerant pump having one or more refrigerant passages which are a part of a type 1 evaporator refrigerant auxiliary circuit and of no other refrigerant circuit.
62. The term ‘dual-return pump’, or more briefly ‘DR pump’, denotes a refrigerant pump having one or more refrigerant passages which are a part of a refrigerant principal circuit and of a type 2 evaporator refrigerant auxiliary circuit belonging to the same principal configuration as the refrigerant principal circuit, and which are a part of no other refrigerant circuit.
63. The term ‘subcooler-circulation pump’, or more briefly ‘SC pump’, denotes a refrigerant pump having one or more refrigerant passages which are a part of a subcooler refrigerant auxiliary circuit and of no other refrigerant circuit.
64. The term ‘hybrid-flow pump’, or more briefly ‘HF pump’, denotes a refrigerant pump having one or more refrigerant passages which are a part of a refrigerant principal circuit and of a subcooler refrigerant auxiliary circuit belonging to the same principal configuration as the refrigerant principal circuit, and which are a part of no other refrigerant circuit.
65. The term ‘principal-circulation pump’, or more briefly ‘PC pump’, denotes a refrigerant pump having one or more refrigerant passages which are a part of a refrigerant principal circuit. The one or more refrigerant passages of a principal-circulation pump may, for example, be (1) a part of no other refrigerant circuit, as in the case of a condensate-return pump; (2) also a part of a type 2 evaporator refrigerant auxiliary circuit of the same principal configuration, as in the case of a dual-return pump; or (3) also a part of a certain type of subcooler refrigerant auxiliary circuit of the same principal configuration, as in the case of a hybrid-flow pump.
66. The term ‘liquid-refrigerant reservoir’, or more briefly ‘LR reservoir’, denotes a vessel for storing liquid refrigerant, the vessel not being a part of a principal configuration.
67. The term ‘liquid-refrigerant ancillary transfer means’, or more briefly ‘ancillary transfer means’, denotes means for transferring liquid refrigerant from an LR reservoir to a principal configuration and for transferring liquid refrigerant from the principal configuration to the LR reservoir. An ancillary transfer means usually includes one or more refrigerant lines, and may also include the one or more refrigerant passages of one or more refrigerant pumps, and/or the one or more refrigerant passages of a refrigerant valve. However, an ancillary transfer means may sometimes merely be a port through which liquid refrigerant, in the LR reservoir, flows into the principal configuration's one or more refrigerant circuits, and through which liquid refrigerant, in the principal configuration's one or more refrigerant circuits, flows into the LR reservoir.
68. The term ‘liquid-transfer pump’, or more briefly ‘LT pump’, denotes a refrigerant pump having one or more refrigerant passages which are a part of an ancillary transfer means and of no other liquid-refrigerant transfer means.
69. The term ‘refrigerant ancillary configuration’, or more briefly ‘ancillary configuration’, denotes a material structure for storing liquid refrigerant and for transferring liquid refrigerant between the ancillary configuration's LR reservoir and a principal configuration; the ancillary configuration comprising the LR reservoir and an ancillary transfer means, and no principal heat exchanger.
70. The term ‘refrigerant configuration’ denotes a material structure consisting in essence of a single principal configuration and one or more ancillary configurations, and having only one refrigerant enclosure.
71. The term ‘airtight refrigerant configuration’ denotes a refrigerant configuration having a refrigerant enclosure
72. The term ‘inert gas’ denotes a gas which does not react chemically in a significantly adverse manner with the refrigerant employed, or with the internal surfaces of the walls of an airtight enclosed space within which the refrigerant and the inert gas are contained, during the operating life of the equipment having the airtight enclosed space. Consequently, the term ‘inert gas’, used in this DESCRIPTION and in the CLAIMS, not only denotes gases usually referred to as inert (such as the noble gases); but also denotes gases such as hydrogen and CO2, or gases such as multi-element gases containing hydrogen and CO2, where they do not react chemically in a significantly adverse manner with the refrigerant, or with the internal surfaces of the walls of an airtight enclosed space within which the refrigerant and the inert gas are contained, during the operating life of the equipment having the airtight enclosed space. In particular, the term ‘inert gas’ includes a gas containing a significant amount of oxygen at the time the gas is inserted in an enclosed space—made immediately thereafter airtight—even where the walls of the enclosed space include one or more metals; provided (1) the refrigerant's heat-transfer properties are essentially unaffected, and provided (2) the one or more metals have essentially not been corroded, by the time essentially all the inserted oxygen has been absorbed by the one or more metals. Thus air may—depending on the refrigerant employed, and on the surfaces with which the refrigerant is in direct contact—be an inert gas. The term ‘inert gas’ also denotes a gas which does not condense over the entire range of operating and environmental conditions experienced by airtight configurations (see definition 86) containing an inert gas.
73. The term ‘inert-gas reservoir’, or more briefly ‘IG reservoir’, denotes a vessel for storing inert gas; but may contain refrigerant vapor mixed primarily with the inert gas, and may even contain liquid refrigerant.
74. The term ‘gas-transfer valve’, or more briefly ‘GT valve’, denotes a valve where the fluid whose flow is controlled by the valve is an inert gas, and where the one or more fluid passages are inert-gas passages interconnecting two spaces containing inert gas.
75. The term ‘gas-transfer pump’, or more briefly ‘GT pump’, denotes a pump for causing inert gas to flow in a desired direction. A GT pump has one or more inert-gas passages through which inert gas flows while the GT pump is active.
76. The term ‘condensate-type refrigerant-vapor trap’ denotes means for removing refrigerant vapor from a fluid which is a mixture of inert gas and refrigerant vapor, the means including means for condensing at least a portion of the refrigerant vapor mixed with the inert gas. A condensate-type refrigerant-vapor trap has a first set of one or more ports through which the inert-gas and refrigerant-vapor enters the trap, and a separate second set of one or more ports through which inert gas, or inert gas and refrigerant vapor, exit the trap. Where inert gas and refrigerant vapor exit a condensate-type refrigerant-vapor trap the mass-flow rate at which refrigerant vapor exits the trap is, under most operating conditions, lower than the mass-flow rate at which refrigerant vapor enters the trap. A condensate-type refrigerant-vapor trap may also have a separate third set of one or more ports through which liquid refrigerant exits the trap. In condensate-type refrigerant-vapor traps having no third set of ports, liquid refrigerant, generated in the traps, exit the traps through their first set of one or more ports.
77. The term ‘inert-gas line’ denotes a conduit for transferring inert gas, or a mixture of inert gas and refrigerant vapor, between the components of an airtight configuration. An inert-gas line may at times also contain a small amount of liquid refrigerant.
78. The term ‘inert-gas transfer means’, or more briefly ‘IG transfer means’, denotes means for transferring inert gas from an IG reservoir to a principal configuration's one or more refrigerant circuits. An IG transfer means usually includes one or more inert-gas lines; and may also (1) include the one or more inert-gas passages of one or more GT pumps, and/or the one or more inert-gas passages of one or more gas-transfer valves; and/or (2) the one or more inert-gas passages of a condensate-type refrigerant-vapor trap.
79. The term ‘inert-gas configuration’, or more briefly ‘IG configuration’, denotes a material structure for storing inert gas, and for controlling the transfer of inert gas between the IG configuration and the one or more refrigerant circuits of a principal configuration. An IG configuration includes an IG reservoir, and active means for causing said inert-gas transfer. The inert gas may, in at least a part of an IG configuration, be mixed with refrigerant vapor.
80. The term ‘inert-gas passive configuration’, or more briefly ‘IGP configuration’, denotes a material structure for storing inert gas and for transferring inert gas between the IGP configuration and the one or more refrigerant circuits of a principal configuration, the IGP configuration including no active means for causing said inert-gas transfer. Consequently, the IG transfer means of an IGP configuration includes no GT-pump inert-gas passages and no GT-valve inert-gas passages. However, an IGP configuration may include one or more valves which perform a different function from that of a GT valve. Examples of non-GT valves are charging, purging, and pressure-relief valves.
81. The term ‘refrigerant & inert-gas space’ or more briefly ‘R&IG space’, denotes an enclosed space containing essentially only refrigerant and inert gas.
82. The term ‘refrigerant & inert-gas enclosure’, or more briefly ‘R&IG enclosure’, denotes a structure determining the bounds of a set of fluidly-connected R&IG spaces containing collectively in essence only refrigerant and inert gas.
83. The term ‘refrigerant & inert-gas configuration’, or more briefly ‘R&IG configuration’, denotes a material structure consisting in essence of
84. The term ‘refrigerant and inert-gas passive configuration’, or more briefly ‘R&IGP configuration’ denotes a material structure consisting in essence of
85. The modifier ‘airtight’ (1) in the term ‘airtight refrigerant & inert-gas configuration’, or ore briefly ‘airtight R&IG configuration’, or (2) in the term ‘airtight refrigerant and inert-gas passive configuration’, or more briefly ‘airtight R&IGP configuration’, denotes respectively an R&IG configuration, or an R&IGP configuration, having an R&IG enclosure
86. The term ‘airtight configuration’ denotes an airtight refrigerant configuration, an airtight R&IG configuration, or an airtight R&IGP configuration.
87. The term ‘supplementary-configuration means’ in the CLAIMS denotes a refrigerant ancillary configuration, an IG configuration, or an IGP configuration.
88. The term ‘inside’, where the subject is an airtight refrigerant configuration, is an abbreviation for the phrase ‘inside the refrigerant enclosure of the airtight refrigerant configuration’. The term ‘inside’, where the subject is an airtight R&IG configuration, is an abbreviation for the phrase ‘inside the R&IG enclosure of the airtight R&IG configuration’. The term ‘inside’, where the subject is an airtight R&IGP configuration, is an abbreviation for the phrase ‘inside the R&IGP enclosure of the airtight R&IGP configuration’. Lastly, the term ‘inside’, where the subject is an airtight configuration, is an abbreviation for, as applicable, the phrases ‘inside the airtight refrigerant configuration’, ‘inside the airtight R&IG configuration’, or ‘inside the airtight R&IGP configuration’.
89. The term ‘total pressure’, where the subject is an airtight configuration, a principal configuration, a refrigerant ancillary configuration, an IG configuration, or an IGP configuration, denotes the sum of the partial refrigerant pressure and the partial inert-gas pressure inside one of the five last-cited configurations.
90. The term ‘airtight two-phase heat-transfer system’ denotes a system which includes an airtight configuration.
91. The term ‘supercharger’ denotes any device employed to increase the pressure, and hence the density, of the combustion or intake air supplied to an internal combustion engine. In particular, the term ‘supercharger’ includes a mechanically-driven supercharger, and an exhaust-gas-driven supercharger, usually referred to as a ‘turbocharger’.
92. The term ‘hot fluid’ denotes a heat source of an airtight configuration, or more specifically a heat source of an airtight configuration's principal configuration. A hot fluid may be a liquid, a gas, or a fluid which changes from its vapor to its liquid phase while it releases heat. In the last of the just-cited three cases the hot fluid may, in particular, be the refrigerant of another airtight configuration. A hot fluid of an airtight configuration transmits heat to the airtight configuration's refrigerant through one or more of the three modes of heat transfer known in the art as conduction heat transfer, convection heat transfer, and radiation heat transfer.
93. The term ‘cold fluid’ denotes a heat sink of an airtight configuration, or more specifically of a heat sink of the airtight configuration's principal configuration. A cold fluid may be a liquid, a gas, or a fluid which changes from its liquid to its vapor phase while it absorbs heat. In the last of the just-cited three cases the cold fluid may, in particular, be the refrigerant of another airtight configuration. The refrigerant of an airtight configuration transmits heat to a cold fluid of the airtight configuration through one or more of the three modes of heat transfer known in the art as conduction heat transfer, convection heat transfer, and radiation heat transfer.
94. The terms ‘hot-fluid valve’ and ‘cold-fluid valve’ denote a valve where the fluid whose flow is controlled by the valve is respectively a hot fluid and a cold fluid, in either their liquid or their vapor phase, and where the one or more fluid passages are respectively hot-fluid passages and cold-fluid passages.
95. The terms ‘hot-fluid pump’ and ‘cold-fluid pump’ denote a pump for causing respectively a hot fluid and a cold fluid—in either their liquid or their vapor phase—to flow in a desired direction. The device has one or more fluid passages through which the hot or cold fluid flows while the device is active.
96. The term ‘motor’ denotes any means for generating mechanical power irrespectively of the source of energy transformed by the motor into mechanical power. Thus, for example, the term ‘motor’ subsumes an internal-combustion engine and an electric motor.
97. The term ‘signal’ denotes any means—including electrical, pneumatic, and hydraulic means—for transmitting information about a thing, and in particular information relating to the current value of a parameter characterizing the state of the thing; or for transmitting information about a required action to be performed by an active device—and in particular about the action to be performed by a refrigerant pump or by a refrigerant valve.
98. The term ‘transducer’ denotes any means for transforming a parameter characterizing state of a thing—and in particular of a refrigerant—into a signal representing the current value the thing's characterizing parameter.
99. The term ‘control unit’ denotes a unit which receives signals from transducers and, on the basis of instructions stored in the unit, generates signals controlling the activities of one or more controllable elements such as pumps and valves. A control unit is usually a microcontroller, with a self-checking capability, having a microprocessor, a read-only memory for storing preselected instructions, a random-access memory for storing signals received by the control unit, and analog and/or digital input-output units for receiving signals from transducers and for supplying signals to one or more controllable elements and to system-status indicators. I distinguish between (1) a principal control unit, referred to in this DESCRIPTION as a ‘central control unit’, or more briefly as a ‘CCU’, because it corresponds to the central control units of the systems disclosed in my co-pending U.S. patent application Ser. No. 400,738, filed 30 Aug. 1989, and (2) a ‘minimum-pressure-maintenance control unit’, or more briefly an ‘MPMCU’, used only to control a system of the invention while the system's principal configuration is inactive.
100. The term ‘active’, where used to indicate the state of a principal configuration, denotes that refrigerant is circulating at a significant rate around at least one of the principal configuration's refrigerant circuits.
101. The term ‘inactive’, where used to indicate the state of a principal configuration, denotes that refrigerant is circulating at a significant rate around none of the principal configuration's one or more refrigerant circuits.
102. The term ‘void fraction’, where the subject is a point along and inside a refrigerant line or a refrigerant passage, denotes the proportion of space occupied by refrigerant vapor at said point, the void fraction being zero where no refrigerant vapor is present and unity where no liquid refrigerant is present.
103. The term ‘flooded’, where the subject is a point on the one or more refrigerant-side heat-transfer surfaces of the condenser of a principal configuration, denotes,
104. The term ‘pre-prescribed’, where used to qualify the way in which something occurs, denotes that way has been specified during the design of a system of the invention. And the term ‘certain pre-prescribed’, where used to qualify operating conditions of a system of the invention, denotes the operating conditions have been specified during the design of the system.
105. The term ‘characterizing parameter’ denotes a parameter providing information about the state of a thing; and in particular the state of (1) an airtight configuration; (2) a heat source of an airtight configuration; (3) the equipment in which the heat source is located; (4) a heat sink of an airtight configuration; (5) the equipment in which the heat sink is located; or (6) the environment of an airtight configuration, where the term ‘environment’ is defined in definition (112). Where an airtight configuration is a refrigerant configuration, the state of an airtight configuration includes the state of the airtight configuration's structure and the state of the airtight configuration's refrigerant; and where an airtight configuration is an R&IG configuration, the state of the airtight configuration includes the state of the airtight configuration's refrigerant, and the state of the airtight configuration's inert gas. (A characterizing parameter may merely be the position of a manually-operated on-off switch.)
106. The term ‘preselected’ where used to qualify the value of a parameter characterizing the state of a thing, or to specify an operating condition, or a range of operating conditions, denotes that the value of the parameter, the operating condition, or the range of operating conditions, respectively, has been specified during the design of a system of the invention. The preselected value of a characterizing parameter—where not otherwise stated or obvious from the context—may be (1) a single value, (2) a value below a preselected upper limit, (3) a value above a preselected lower limit, or (4) a value between a preselected upper limit and a preselected lower limit. A preselected single value, a preselected upper limit, or a preselected lower limit, may (1) be fixed, (2) have a range of manually selectable fixed values, or (3) change with time in a pre-prescribed way as a function of one or more preselected characterizing parameters.
107. The term ‘preselected range of operating conditions’, and the term ‘preselected range of environmental conditions’, where the subject is an airtight configuration, denote respectively the entire range of operating conditions under which the airtight configuration is designed to function and the entire range of environmental conditions under which the airtight configuration has a specified property; the preselected range of operating conditions and environmental conditions being specified, during the airtight configuration's design, in terms of preselected ranges for the values of one or more preselected characterizing parameters.
108. The term ‘steady-state conditions’, where the subject is an airtight configuration, denotes operating conditions under which all characterizing parameters affecting refrigerant flow, and where applicable inert-gas flow, in the airtight configuration, change at a negligible rate compared to the slowest response rate of the airtight configuration's one or more refrigerant circuits, and where applicable inert-gas circuits.
109. The term ‘transient conditions’, or more briefly ‘transient’, where the subject is an airtight configuration, denotes operating conditions under which at least one characterizing parameter affecting refrigerant flow, and where applicable inert-gas flow, changes at a faster rate than the slowest response rate of the airtight configuration's one or more refrigerant circuits, and where applicable inert-gas circuits.
110. Each of the two terms ‘upstream’ and ‘downstream’ denotes the relative location of two points, or of two components, with respect to the direction of flow of, as applicable, a refrigerant, an inert gas, a hot fluid, or a cold fluid. The last-cited two terms apply to the case where, as applicable, the refrigerant, the hot fluid, or the cold fluid, flows in only one direction under steady-state conditions, and refer to the direction of flow of respectively the refrigerant, the hot fluid, or the cold fluid, under those conditions.
111. The term ‘amount of liquid’ denotes the volume occupied by a liquid.
112. The term ‘heating load’ denotes the rate at which heat is transmitted from a heat source to a refrigerant. (A heat source may be a refrigerant.)
113. The term ‘cooling load’ denotes the rate at which heat is transmitted from a refrigerant. (A heat sink may be a refrigerant.)
114. The term ‘environment’, where the subject is an airtight configuration, denotes the one or more contiguous and/or remote material substances which surround an airtight configuration, and which collectively determine the temperature to which the airtight configuration's refrigerant tends while the refrigerant's circulation is zero around all of the one or more refrigerant circuits of the airtight configuration's principal configuration. For example, in most applications where an airtight configuration is located inside a building, the airtight configuration's environment is the air inside that building in direct contact with the airtight configuration; and the walls, ceiling, and floor, with which the airtight configuration exchanges heat. And, in the case where the airtight configuration is located in an open space, the airtight configuration's environment is the air in direct contact with the airtight configuration; and the bodies, including celestial bodies, outside the airtight configuration with which the airtight configuration exchanges heat.
115. The term ‘controllable element’ in this DESCRIPTION, and synonymously the term ‘controllable means’ in the CLAIMS, denotes an active device which can be controlled by a signal. Examples of controllable elements or means are refrigerant pumps and valves, hot-fluid pumps and valves, cold-fluid pumps and valves, controllers of electric motors or of the burners of a boiler, and electrical switches for starting and stopping internal-combustion engines. A controllable element or means may be a part of a system of the invention, or of another system with which a system of the invention interacts. In either of the two cases cited in the immediately-preceding sentence, a controllable element or means (1) may be controlled exclusively by a system of the invention or only in part by a system of the invention, or (2) may not be controlled by a system of the invention even though it is a part of a system of the invention. The signal cited in this definition 113 includes a signal generated by a transducer which is an integral part of the controllable element or means, as for example in the case where the controllable element is a thermostat.
116. The term ‘system-controllable element’ in the DESCRIPTION, and synonymously the term ‘system-controllable means’ in the CLAIMS, denotes a controllable element or means which is controlled at least in part by the system. A system-controllable element or means may be a part of a system of the invention, or of another system with which the system of the invention interacts. Where in this DESCRIPTION it is obvious that a controllable element is a system-controllable element I shall simply refer to a system-controllable element as a ‘controllable element’. Examples of cases where a controllable element is obviously a system-controllable element include the cases where a controllable element is described as being controlled, or is shown in the FIGURES as being controlled, by a signal supplied by a central control unit, or by a minimum-pressure-maintenance control unit, of the system.
117. In the context of a system of the invention, (1) the term ‘control mode’ denotes a set of one or more preselected rules for controlling one or more system-controllable elements or means, the set of one or more preselected rules including a single rule for controlling each system-controllable element or means in a pre-prescribed way as a function of one or more preselected characterizing parameters; (2) the expression ‘has several control modes’ denotes the system includes means for executing each of the several control modes; and (3) the expression ‘is in a control mode’ denotes the system is executing a control mode. ‘A preselected rule’ may be an instruction stored in a control unit; or may be, as for example in the case of a thermostat, a rule inherent in the design of a system-controllable element or means. In the context of a system of the invention having several control modes and in the context of a recited system action, the expression ‘in at least one of several control modes’ denotes the system executes the recited action in at least one of the system's one or more control modes. The term ‘control mode’ may include a set of one or more rules requiring none of the one or more system-controllable elements or means to be controlled by the system.
118. In the context of a system of the invention (1) the term ‘transition rule’ denotes a set of one or more preselected rules for changing from one of the system's several control modes to another of the system's several control modes; and (2) the expression ‘has several transition rules’ denotes the system includes means for executing each of the several transition rules. The term ‘transition rule’ may include a set of one or more preselected rules for changing (1) from a control mode where none of the the one or more system-controllable elements or means is controlled by the system to another control mode where at least one of the one or more system-controllable elements or means is controlled by the system; and (2) from a control mode where at least one of the one or more system-controllable elements or means is controlled by the system to a control mode where none of the system-controllable elements or means is controlled by the system.
119. The term ‘system-control means’, in the CLAIMS, denotes the devices employed to control the system-controllable elements or means of a system of the invention, and subsumes, where applicable, a central control unit, a minimum-pressure-maintenance control unit, one or more transducers, and the means used to control the one or more system-controllable elements or means of the system. Where a system-controllable element or means of a system of the invention is controlled by a transducer and an actuator which are an integral part of the system-controllable element or means, the transducer and the actuator of the system-controllable means are a part of the system-control means of the system to which the system-controllable means belongs. An example of a system-controllable element or means having its own transducer and actuator is a thermostatically-controlled valve.
120. The term ‘and/or’ denotes, as applicable, that two or more material things referred to may be, or may not be, located in the selfsame structure; or that two or more events, or two or more actions, referred to may occur, or may not occur, simultaneously.
121. The term ‘major paragraph’ denotes text in this DESCRIPTION between a heading and a horizontal line consisting of dashes, or text between two horizontal lines consisting of dashes.
122. The term ‘minor paragraph’ denotes in this DESCRIPTION a subparagraph within a major paragraph.
123. The term ‘one or more airtight refrigerant circuits’ denotes a set of one or more refrigerant circuits which ingest essentially no ambient air after they have been charged with refrigerant.
124. The phrase ‘an airtight configuration having an enclosure’ denotes that an airtight configuration has a refrigerant enclosure (see definition 27) or an R&IG enclosure (see definition 82).
125. The phrases ‘inside an airtight configuration’ and ‘inside the airtight configuration’ are abbreviations for respectively the phrases ‘inside the enclosure of an airtight configuration’ and ‘inside the enclosure of the airtight configuration’.
126. The term ‘refrigerant-circuit configuration’ is in essence synonymous with the term ‘refrigerant principal configuration’. The only difference between the last two terms is that the former term is used where a system of the invention has no supplementary-configuration means (see definition 87) fluidly connected to a refrigerant-circuit configuration, whereas the latter term is used to denote a refrigerant-circuit configuration of a system of the invention fluidly connected to supplementary-configuration means.
127. The term ‘evacuated refrigerant-circuit configuration’ denotes a refrigerant-circuit configuration having a refrigerant enclosure
128. The term ‘evacuated configuration’ denotes an airtight refrigerant configuration, or an evacuated refrigerant-circuit configuration.
129. The term ‘active’, where used to indicate the state of an evacuated configuration, denotes that refrigerant is circulating at a significant rate around at least one of the evacuated configuration's refrigerant circuits.
130. The term ‘inactive’, where used to indicate the state of an evacuated configuration, denotes that refrigerant is circulating at a significant rate around none of the evacuated configuration's one or more refrigerant circuits.
A first general purpose of the invention is to devise airtight configurations (see definitions) and control techniques for endowing airtight two-phase heat-transfer systems with a property named ‘minimum-pressure maintenance’. This property ensures, broadly speaking, that the pressure inside an entire airtight configuration, or inside a part of an airtight configuration, is maintained at or above a preselected minimum pressure, higher than the refrigerant's lowest saturated-vapor pressure while the airtight configuration's principal configuration is inactive and while the airtight configuration is in thermal equilibrium with its environment. For example, in the case where an airtight configuration's lowest thermal equilibrium temperature with its environment is 0° C. while it is inactive, and where the configuration's refrigerant is water, the refrigerant's lowest saturated-vapor pressure is 0.61 kPa, and the preselected minimum pressure would be higher than 0.61 kPa. (0.61 kPa is the saturated-vapor pressure of water corresponding to 0° C.) I distinguish, as explained in section III,D, between ‘complete minimum-pressure maintenance’ and ‘partial minimum-pressure maintenance’.
A second general purpose of the invention is to devise airtight configurations and control techniques for endowing airtight configurations with one or more of the properties named ‘freeze protection’, ‘self regulation’, ‘refrigerant-controlled heat release’, ‘gas-controlled heat release’, ‘refrigerant-controlled heat absorption’, and ‘evaporator liquid-refrigerant injection’; and to devise evacuated configurations and control techniques for endowing evacuated configurations with the property named ‘evaporator liquid-refrigerant injection’.
Other important purposes of the invention will be disclosed later in this DESCRIPTION.
The eight properties cited in this section III,B are disclosed and discussed in sections III,D to III,H. I note that the three properties named ‘complete minimum-pressure maintenance’, ‘partial minimum-pressure maintenance’, and ‘freeze protection’, pertain to airtight configurations while their principal configuration is inactive. The other five of the eight properties cited in this section pertain to airtight configurations while their principal configuration is active.
The invention disclosed in this DESCRIPTION covers two-phase heat-transfer systems that include an airtight configuration, or an evacuated configuration, and associated control system for transferring heat from one or more heat sources to one or more heat sinks and for achieving at least one of the eight properties cited in section III,B. The term ‘two-phase heat-transfer systems includes ‘two-phase heat-transfer heating systems’ and ‘two-phase heat-transfer cooling systems’ where the qualifiers ‘heating’ and ‘cooling’ indicate the primary purpose of a two-phase heat-transfer system. I shall, in this DESCRIPTION, use the terms ‘two-phase heating systems’ and ‘two-phase cooling systems’ as abbreviations for respectively the terms ‘two-phase heat-transfer heating systems’ and ‘two-phase heat-transfer cooling systems’. It follows that the two last-cited abbreviations do not include heat pumps and refrigerators.
The airtight configurations used in systems of the invention are combinations of
I shall refer to the combination specified under (a) (in the immediately-preceding minor paragraph) as a ‘type A combination’; to the combination specified under (b) as a ‘type B combination’; and to the combination specified under (c) as a ‘type C combination’.
All airtight configurations of the invention have, by definition, only a single principal configuration. However type A combinations may have one or more ancillary configurations; type B combinations may have one or more ancillary configurations and one or more IG or IGP configurations; and type C combinations may have one or more IG or IGP configurations.
Many systems of the invention, in addition to including one or more airtight configurations, also include the parts of other material structures cooperating with the airtight configurations to achieve at least one or more of the eight properties recited in section III,B. Those parts include control units and components (including their associated supporting structures) cooperating with the one or more airtight configurations. Examples of such cooperative components include equipment generating certain heat sources, such as the burners of boilers, hot-fluid pumps such as the burners' blowers, and cold-fluid pumps such as the fans of fan-coil units and the radiators of internal-combustion-engines.
An airtight configuration of the invention, or an evacuated configuration of the invention, has one or more hot heat exchangers and one or more cold heat exchangers. I shall refer to the heat source from which the refrigerant in (the one or more refrigerant passages of) a hot heat exchanger absorbs heat as the ‘hot heat exchanger's heat source’; and, where the heat exchanger is an evaporator, a preheater, or a superheater, I shall refer to the heat source as the ‘evaporator's heat source’, as the ‘preheater's heat source’, or as the ‘superheater's heat source’, respectively. And I shall refer to the heat sink to which the refrigerant in (the one or more refrigerant passages of) a cold heat exchanger releases heat as the ‘cold heat exchanger's heat sink’; and where the heat exchanger is a condenser, a subcooler, or a desuperheater, I shall refer to the heat sink as the ‘condenser's heat sink’, as the ‘subcooler's heat sink’, or as the ‘desuperheater's heat sink’, respectively. The hot heat exchangers of an airtight configuration of the invention may have the same heat source or different heat sources; and similarly the cold heat exchangers of an airtight configuration of the invention may have the same heat sink or different heat sinks.
All hot heat exchangers of an airtight configuration of the invention have, by definition, one or more refrigerant passages wherein the refrigerant absorbs heat, released by the hot heat exchanger's heat source, while the airtight configuration to which the hot heat exchanger belongs has an active principal configuration. And all cold heat exchangers of an airtight configuration have one or more refrigerant passages wherein the refrigerant releases heat, absorbed by the cold heat exchanger's heat sink, while the airtight configuration to which the cold heat exchanger belongs has an active principal configuration.
In applications where the heat source of a hot heat exchanger is a hot fluid which is at least in part in direct contact with the walls of the hot heat exchanger's (one or more) refrigerant passages, the hot heat exchanger usually has one or more surfaces which bound one or more enclosed spaces or one or more open spaces, named ‘fluid ways’, to which the hot fluid—while the airtight configuration to which the hot heat exchanger belongs is active—releases heat absorbed by refrigerant in the hot heat exchanger's refrigerant passages. Similarly, in applications where the heat sink of a cold heat exchanger is a cold fluid which is at least in part in direct contact with the walls of the cold heat exchanger's (one or more) refrigerant passages, the cold heat exchanger usually has one or more surfaces which bound one or more enclosed spaces or one or more open spaces, named ‘fluid ways’, from which the cold fluid—while the airtight configuration to which the cold heat exchanger belongs is active—absorbs heat released by refrigerant in the cold heat exchanger's refrigerant passages. Examples of enclosed spaces, in the sense intended by me, are the space inside a tube or inside a rectangular duct; the space inside an annulus formed by concentric tubes; the space between the internal surface(s) of an open or a closed cylinder and the external surfaces of several interconnected tubes inside the cylinder; and the space between the internal surface(s) of an open or a closed rectangular duct and the external surfaces of several rectangular ducts inside the rectangular duct. And examples of open spaces, in the sense intended by me, are the space inside a building or the space inside a room of a building, the space outside a building, the space inside a water reservoir, and the space occupied by a lake.
A heat source of a hot heat exchanger of an airtight configuration of the invention is always also a heat source of the airtight configuration, or more specifically of the airtight configuration's principal configuration; and a heat sink of a cold heat exchanger of the airtight configuration is always also a heat sink of the airtight configuration, or more specifically of the airtight configuration's principal configuration. Thus the set of one or more heat sources of an airtight configuration of the invention, or equivalently of the airtight configuration's principal configuration, is the set of the one or more heat sources of the airtight configuration's one or more hot heat exchangers; and the set of one or more heat sinks of the airtight configuration, or equivalently of the airtight configuration's principal configuration, is the set of the one or more heat sinks of the airtight configuration's one or more cold heat exchangers.
The heat source of a hot heat exchanger may be a material substance remote from the hot heat exchanger. Examples of remote heat sources are the sun, flames, and high-temperature metal slabs and rods not in contact with the refrigerant passages of the hot heat exchanger. The heat source may also be a material substance at least in part contiguous to, or in the fluid ways of, a hot heat exchanger. Examples of the latter heat source include
The heat sink of a cold heat exchanger may be, for example, a material substance, such as an extra-terrestrial body or a terrestrial body (such as the wall of a room) remote from the system: or it may be a material substance, at least in part, contiguous to or in the fluid ways of the cold heat exchanger. Examples of the latter heat sink include
Heat may be transmitted from a hot heat exchanger's heat source to refrigerant in the hot heat exchanger, and from refrigerant in a cold heat exchanger to the cold heat exchanger's heat sink, by radiation, convection, or conduction, or by a combination of any two, or of all three, of the foregoing heat-transmittal mechanisms. For example, in the case where the heat source is the sun and the one or more refrigerant passages of a hot heat exchanger are made of glass transparent to thermal radiation, heat is transmitted from the heat source to the refrigerant in the hot heat exchanger essentially only by radiation; and, in the case where the heat source is the flame and combustion gas in a fired steam boiler (having refrigerant passages exposed to radiation from the flame), heat is transmitted from the heat source to the refrigerant in the boiler by radiation, convection, and conduction.
Airtight configurations, or evacuated configurations, of the invention not only include configurations employing a refrigerant whose refrigerant pressure is below ambient atmospheric pressure while they are inactive, but also configurations employing a refrigerant whose pressure stays below ambient atmospheric pressure while they are active. In particular, airtight refrigerant configurations of the invention include airtight refrigerant configurations, employing H2O as their refrigerant, that operate exclusively at subatmospheric pressures. Such configurations, in contrast to non-airtight refrigerant configurations employing H2O as their refrigerant, need no vacuum pump to operate at sub-atmospheric pressures.
The refrigerant used in an airtight configuration, or an evacuated configuration, of the invention may be, in principle, any fluid whose liquid and vapor phases can coexist over the entire range of operating refrigerant evaporation temperatures of interest in the particular application considered. The phrase ‘any fluid’ is intended to include not only single-component fluids, and (multi-component) azeotropic fluids, which evaporate at a single (sensible) temperature at a given pressure, but also (multi-component) non-azeotropic fluids which evaporate over a range of temperatures at a given pressure.
Examples of single-component or azeotropic refrigerants which are in principle suitable for the systems of the present invention include refrigerants suitable for heat pipes, tube thermo-siphons, loop thermosiphons, and heat pumps.
A partial list of single-component and azeotropic refrigerants which have been considered for, or used in, heat pipes and heat pumps is given respectively in P. D. Dunn and D. A. Reay, ‘Heat Pipes’, 2nd Edition, published 1969 by Pergamon Press (London), see page 293; and ‘Thermodynamic Properties of Refrigerants’, published 1969 by ASHRAE (New York), see Table of Contents. And a partial list of non-azeotropic, non-aqueous refrigerants which have been considered for heat pumps is given in a paper by Prof. Thore Bentsson and Dr. Hans Schnitzer, ‘Some Technical Aspects on Nonazeotropic Mixtures as Working Fluids’, presented in September 1984 at the International Symposium on ‘The Large Scale Applications of Heat Pumps’ organized and sponsored by BHRA, The Fluid Engineering Centre, Cranfield, Bedford, England. In addition to the fluids listed in the papers cited in this minor paragraph, a number of non-azeotropic aqueous refrigerants are in principle suitable for the systems of the present invention. These include aqueous solutions of glycol, ethanol, methanol, or acetone. Some of the foregoing azeotropic-like refrigerants—such as chlorofluorocarbons—are no longer acceptable, but I envisage the evacuated configurations of the invention employing acceptable substitutes such as Isceon 69S.
In practice, the usefulness of a refrigerant for a given application is limited by a number A of constraints. For example, the refrigerant evaporation pressures, and the refrigerant saturated-vapor specific volumes, corresponding to the refrigerant evaporation and condensation temperatures of interest must not be unacceptably high; the refrigerant must not decompose chemically at the highest temperatures which may occur while the system, in which the refrigerant is employed, is active or is inactive; and the cost of the system's refrigerant must not be unacceptably high.
The materials from which the inside surfaces of the walls of the refrigerant passages of an airtight configuration, or an evacuated configuration, of the invention are made must be compatible with their refrigerant. And, where heat-exchanger refrigerant passages of the configuration come into direct contact with a heat source or a heat sink, the materials from which the outside surfaces of the walls of these refrigerant passages are made must also be compatible with the heat source or the heat sink. The term ‘compatible’ is used herein to indicate that the materials from which refrigerant passages are made have no unacceptable adverse effect on the refrigerant, the heat source, or the heat sink; and also, conversely, to indicate that the refrigerant, the heat source, or the heat sink, have no unacceptable adverse effect on the materials from which the walls of refrigerant passages are made.
A system of the invention having several airtight configurations may use
The systems of the invention may be used in a land vehicle, a surface vehicle, a submerged vehicle, or an airborne vehicle—as well as in a fixed ground installation—provided these systems are not required to operate efficiently whilst the vehicle in which they are installed is undergoing a steady-state acceleration having a substantial component normal to the local gravitational field or a substantial component parallel and opposite to this field. What constitutes a ‘substantial’ component depends on the particular system considered, but a component, to be substantial, might often have to be as large as 0.5 g, 0.75 g, or even larger.
Systems of the invention comprise systems having a heat source controlled in part or entirely by them as well as a heat source not controlled by them. The equipment associated with the former heat source is usually a part of a system of the invention; whereas the equipment associated with the latter heat source is usually not a part of a system of the invention. Examples of heat sources which are controlled by, and which—together with their associated equipment—are entirely a part of, a system of the invention comprise finite thermal-capacity heat sources such as the combustion gases of a steam boiler of the invention used to heat buildings or to supply heat to industrial processes. And examples of heat sources which are not controlled by a system of the invention include
Systems of the invention also comprise systems having a heat sink controlled by them as well as heat sinks not controlled by them. The former heat sink—and its associated equipment—is usually a part of a system of the invention; whereas the latter heat sink—and most or all of its associated equipment—is not a part of a system of the invention.
1. General Remarks
Minimum-pressure maintenance may, as mentioned in section III,B, be complete or partial. The qualifier ‘complete’ denotes that the internal pressure inside an entire airtight configuration always stays at or above a preselected minimum pressure; and the qualifier ‘partial’ denotes that the internal pressure inside only a part of an airtight configuration always stays at or above a preselected minimum pressure. The latter property is useful where only a part of an airtight configuration would be subjected to an unacceptably high net external pressure, or would ingest air, if its internal pressure fell substantially below a preselected minimum pressure. Examples of such a part are an air-cooled condenser which would be subjected to unacceptably high crushing pressures, or a refrigerant pump with mechanical seals through which air would be ingested, if the internal pressure of those parts fell substantially below a preselected minimum pressure above the lowest refrigerant saturated-vapor pressure inside an airtight configuration while the configuration is inactive.
2. Type A Combinations
Complete minimum-pressure maintenance is achieved with type A combinations by
Partial minimum-pressure maintenance is achieved with type A combinations by
Complete minimum-pressure maintenance is achieved with type B and C combinations by
Partial minimum-pressure maintenance is achieved with type C combinations by
Inert-gas transfer between a principal and an IG configuration is controlled primarily by one or more controllable elements of the IG configuration; whereas inert-gas transfer between a principal and an IGP configuration is controlled primarily by the total pressure in the one or more refrigerant circuits of the principal configuration.
Partial minimum-pressure maintenance is achieved in type B combinations either in the way it is achieved in type A combinations or in the way it is achieved in type C combinations.
The purpose of freeze protection is to prevent liquid refrigerant freezing in the principal configuration of an airtight configuration while the entire principal configuration, or while one or more parts of the principal configuration, are exposed to refrigerant subfreezing temperatures.
Freeze protection with a type A or with a type B combination is achieved in essence by
I note that the kind of freeze-protection method just outlined differs considerably from the freeze-protection method recited in section III,F of my co-pending U.S. patent application Ser. No. 400,738, filed 30 Aug. 1989: the former method stores liquid refrigerant that could be exposed to subfreezing temperatures outside the principal configuration whereas the latter method stores liquid refrigerant that could thus be exposed inside the principal configuration.
1. General Remarks
Techniques, named ‘self-regulation techniques’ have been devised by me to ensure, broadly speaking, that a principal configuration transfers heat—under pre-prescribed operating conditions—efficiently over the entire range of those operating conditions. I have named the property achieved by using self-regulation techniques ‘self regulation’.
Self regulation of a principal configuration is achieved by
Self regulation of a principal configuration is defined precisely in terms of a preselected set of ‘specific self-regulation conditions’ formulated for a particular heat-transfer application. However, these specific conditions always satisfy collectively, in the case of a principal configuration with an FRC principal circuit, four conditions, named ‘universal self-regulation conditions’, which do not depend on the particular application considered. Only the first three of the four universal self-regulation conditions apply to a principal configuration with an NRC principal circuit. The four universal self-regulation conditions are discussed next.
2. Universal Self-regulation Conditions
The four universal self-regulation conditions require—for a pre-prescribed set of operating conditions—the refrigerant flow, in a principal configuration with a principal refrigerant pump, to be controlled so that, with the principal configuration charged with an appropriate amount of refrigerant mass,
I shall refer individually to the four universal self-regulation conditions just recited as ‘self-regulation conditions (A), (B), (C), and (D)’, respectively. And I shall say that a principal configuration with an FRC principal circuit ‘achieves self regulation’, or alternatively ‘is in its self-regulation mode’, when the four self-regulation conditions are satisfied irrespectively of whether all preselected specific self-regulation conditions for that configuration are satisfied. And I shall further say that an airtight configuration ‘achieves self regulation’ or alternatively ‘is in its self-regulation mode’ if the airtight configuration's principal configuration achieves self regulation or is in its self-regulation mode; and that an airtight configuration satisfies a particular self-regulation condition when the airtight configuration's principal configuration satisfies that particular condition.
The foregoing four conditions, irrespectively of the specific self-regulation conditions selected for a particular heat-transfer application, can be achieved without using a refrigerant-vapor throttling valve; thereby allowing—for the entire pre-prescribed operating conditions—the absolute value of the difference between
I note that self-regulation conditions (A) to (C) can be achieved by principal configurations, having an FRC principal circuit, with far fewer spatial constraints than by principal configurations having an NRC principal circuit. In particular, the former configurations can satisfy self-regulation conditions (A) to (C) with their condenser below as well as above, or at the same height as, their evaporator; whereas the latter configurations cannot satisfy self-regulation conditions (A) to (C) with their condenser below their evaporator, and this makes the latter systems unsuitable for many important applications.
I also note that a principal configuration having an FRC principal circuit, may often be preferable to a principal configuration having an NRC principal circuit even in applications where the configuration's condenser may be, or is required to be, placed above the configuration's evaporator. Examples of such applications include applications where the condenser of a principal configuration with an NRC principal circuit would have to be placed at an unacceptably-great height—say at a height of over ten meters—above the evaporator of the principal configuration to allow the net refrigerant static head in the NRC principal circuit to overcome the total friction-induced pressure drop around this circuit. (The total friction-induced pressure drop around an NRC principal circuit may be high because the refrigerant mass-flow rate per unit refrigerant passageway cross-sectional area is high in the evaporator refrigerant passages, or in the condenser refrigerant passages, or in both, because of system requirements.)
3. Specific Self-regulation Conditions
Each specific self-regulation condition is expressed in terms of a preselected quantity, named a ‘self-regulation quantity’, and a preselected constraint on the current value of that quantity. This constraint may be expressed in any one of the following four ways:
The self-regulation quantities chosen for a set of specific self-regulation conditions may, even in the absence of a refrigerant auxiliary circuit, include
The foregoing four specific self-regulation quantities are intended to be only illustrative examples of self-regulation quantities and not to constitute an exhaustive list of these quantities.
The pre-selected specific self-regulation quantity may be
Internal characterizing parameters are those characterizing the state of a thing which is a part of an airtight configuration. This thing is usually the airtight configuration's enclosure or the airtight configuration's refrigerant. Examples of parameters characterizing an airtight configuration's enclosure are its temperature at a location of the enclosure. And examples of parameters characterizing the state of the refrigerant are
External characterizing parameters are those characterizing the state of a thing which is not a part of an airtight configuration. Examples of things which are not a part of an airtight configuration are a heat source, a heat sink, and ambient air, of the configuration. In applications where a heat source is a fluid, referred to henceforth as a ‘hot fluid’, and a heat sink is also a fluid, referred to henceforth as a ‘cold fluid’, examples of parameters characterizing the hot fluid and the cold fluid are:
The measures of internal or of external characterizing parameters recited in the immediately-preceding two minor paragraphs may be direct measures or indirect measures.
Examples of indirect measures are:
Most techniques used for satisfying a set of specific self-regulation conditions consist in essence in
The choice of a set of specific self-regulation conditions for a particular heat-transfer application depends greatly, but not solely,
Examples of spatially substantially-uniform temperature heat sources are a fluid which releases heat while undergoing a change in phase with no significant pressure drop, and a metal slab being cooled. Examples of a spatially substantially-uniform heat sink are a fluid which absorbs heat while undergoing a change in phase with no significant pressure drop, and a water reservoir, with no significant temperature gradient, within which a cold heat exchanger is immersed. Examples of heat sources which release heat while undergoing a substantial drop in temperature, and of heat sinks which absorb heat while undergoing a substantial rise in temperature, are fluids which respectively release and absorb heat without changing phase at low mass-flow rates.
1. Preliminary Remarks
The rate at which radiant energy is transmitted from a high-temperature refrigerant in cold heat exchanger to remote substances, such as the walls or floors of a building or extraterrestrial bodies, can be changed by a shutter opaque to thermal radiation. This shutter is used to intercept partly, or even totally, thermal radiant energy emitted by the refrigerant itself or by the cold heat exchanger's heat-transfer surfaces. In the former case, the cold heat-exchanger heat-transfer surfaces are transparent to thermal radiant energy and, in the latter case, those heat-transfer surfaces are made of heat-conducting material.
The rate at which heat is transmitted from a refrigerant in a cold heat exchanger to a contiguous cold fluid can be changed by cold-fluid valves (including dampers or shutters), and/or by cold-fluid pumps. Where the cold fluid absorbs heat without changing phase, the two last-cited devices are used to change the cold fluid's mass-flow rate. And, where the cold fluid absorbs heat by changing from a liquid to a vapor, those two devices are used to change the amount of liquid cold fluid in direct contact with the cold heat exchanger's external heat-transfer surfaces.
I shall hereinafter use the term ‘externally-controlled heat release’, or more briefly the term ‘EC heat release’, to denote the methods of heat-release control outlined in the immediately-preceding two minor paragraphs. (The qualifier ‘externally-controlled’ refers to the fact that the means used to achieve heat-release control are not a part of an airtight configuration.) The techniques for controlling shutters opaque to thermal radiation, and cold-fluid valves (including dampers or shutters) and pumps, are well known. They shall therefore not be discussed in this DESCRIPTION.
The rate at which a refrigerant in a cold heat exchanger releases heat to remote substances or to a contiguous cold fluid can—where a cold heat exchanger is a condenser—also be changed by controlling the amount of liquid refrigerant in the condenser's refrigerant passages. I shall hereinafter use the term ‘refrigerant-controlled heat release’, or more briefly ‘RC heat release’, to denote heat-release control achieved by changing the amount of liquid refrigerant in the condenser refrigerant passages of a principal configuration.
I note that RC heat release is an operating mode of an airtight configuration, and is achieved by controlling the refrigerant of an airtight configuration in a way which differs from the way it would be controlled to achieve self regulation. By contrast, EC heat release is not an operating mode of an airtight configuration and is not achieved by controlling the refrigerant of an airtight configuration. Consequently, self regulation and RC heat release are two mutually exclusive operating modes of a two-phase heat-transfer system; whereas self regulation and EC heat release are not mutually exclusive operating modes of a two-phase heat-transfer system, and can therefore coexist. Furthermore, RC heat release and EC heat release are also not mutually exclusive operating modes of a two-phase heat-transfer system, and can therefore also coexist.
The rate at which a refrigerant in a cold heat exchanger releases heat to remote substances, or to a contiguous fluid, can—where the cold heat exchanger is a condenser—alternatively be changed by controlling the amount of inert-gas mass in the condenser's refrigerant passages. I shall hereinafter use the term ‘gas-controlled heat release’, or more briefly ‘GC heat release’ to denote heat-release control achieved by changing the amount of inert-gas mass in the condenser refrigerant passages of a principal configuration.
I note that GC heat-release, in contrast to RC heat release, can coexist with self regulation; and that GC heat release, like RC heat release, can coexist with EC heat release.
2. Refrigerant-controlled Heat Release
The purpose of RC heat release is usually to control the rate at which refrigerant releases heat in the condenser refrigerant passages of a principal configuration at a preselected refrigerant pressure or equivalently, in the case of an azeotropic-like refrigerant, at a preselected refrigerant saturated-vapor temperature. The preselected refrigerant pressure may be fixed or may change in a pre-prescribed way.
RC heat release is achieved with type A, or with type B, combinations by controlling the amount of liquid refrigerant in the one or more condenser refrigerant passages of their principal configuration in pre-prescribed ways, which fall into three general categories.
The first general category of RC heat-release techniques achieve heat-release control by satisfying self-regulation condition (B) and violating self-regulation condition (C); namely by supplying a condenser's refrigerant passages with essentially dry refrigerant, and by increasing the amount of liquid refrigerant, backing-up into those passages, above that allowed by self-regulation condition (C).
The second general category of RC heat-release techniques achieve heat-release control by violating self-regulation condition (B) and satisfying self-regulation condition (C); namely by supplying wet refrigerant vapor to a condenser's refrigerant passages whilst not allowing liquid refrigerant to back-up into those passages by an amount exceeding that allowed by self-regulation condition (C).
The third general category of RC heat-release techniques achieve heat-release control by violating self-regulation conditions (B) and (C).
In the particular case where a condenser is a split condenser including several component condensers (see section V,B,12), liquid refrigerant can be inserted into, and extracted from, component condensers independently by using several ancillary configurations or even by using a single ancillary configuration.
3. Gas-controlled Heat Release
Broadly speaking, the purpose of GC heat release is the same as that as RC heat release. However, where GC heat release is used, the preselected pressure at which the rate of heat release is controlled is usually the total pressure of the refrigerant and inert gas. (This total pressure is of course essentially equal to the refrigerant pressure at a point, inside an airtight configuration, where the partial pressure of the inert gas is negligible.)
GC heat-release is achieved with type B, or with type C, combinations having an IG configuration by transferring inert gas from their IG reservoir to their condenser's refrigerant passages, and from their condenser's refrigerant passages to their IG reservoir, in a pre-prescribed way. Inert gas can be inserted into, or extracted from, those passages through the condenser's refrigerant inlet, through the condenser's refrigerant outlet, or through one or more ports along the condenser's refrigerant passages.
In the particular case where a condenser is a split condenser including several component condensers, inert gas can be inserted into, or extracted from, component condensers independently by using several IG configurations, or even by using only a single IG configuration.
1. Preliminary Remarks
The rate at which radiant thermal energy is transmitted from a remote high-temperature material substance, such as a flame or the sun, to a refrigerant in a hot heat exchanger can be changed by a shutter opaque to thermal radiation. This shutter is used to intercept partly, or even totally, thermal radiant-energy absorbed by the refrigerant itself or by the hot heat exchanger's heat-transfer surfaces. In the former case, the hot heat exchanger heat-transfer surfaces are transparent to thermal radiation and, in the latter case, those heat-transfer surfaces are made of heat-conducting material.
The rate at which heat is transmitted from a hot fluid to a contiguous refrigerant in a hot heat exchanger can be changed by hot-fluid valves (including dampers or shutters), and/or by hot-fluid pumps. Where the hot fluid releases heat without changing phase, the two last-cited devices are used to change the hot fluid's mass-flow rate. And, where the hot fluid releases heat by changing from a vapor to a liquid, those two devices are used to change the amount of liquid hot fluid in direct contact with the hot heat exchanger's external heat-transfer surfaces.
I shall hereinafter use the term ‘externally-controlled heat absorption’, or more briefly the term ‘EC heat absorption’, to denote methods of heat absorption control outlined in the immediately-preceding two minor paragraphs.
The rate at which a refrigerant in a hot heat exchanger absorbs heat from a remote substance, or from a contiguous hot fluid, can—where the hot heat exchanger is an evaporator—also be changed by controlling the amount of liquid refrigerant in, and/or the refrigerant mass-flow rate through, the evaporator's refrigerant passages. I shall hereinafter use the term ‘refrigerant-controlled heat absorption’, or more briefly the term ‘RC heat absorption’, to denote heat-absorption control recited in the immediately-preceding sentence.
I note that RC heat absorption—like RC heat release—is an operating mode of an airtight configuration, and is achieved by controlling the refrigerant of an airtight configuration in a way which differs from the way it would be controlled to achieve self regulation. By contrast, EC heat absorption—also like EC heat release—is not achieved by controlling the refrigerant of an airtight configuration. Consequently, self regulation and RC heat absorption—like self regulation and RC heat release—are two mutually-exclusive operating modes of a two-phase heat-transfer system; whereas self regulation and EC heat absorption—also like self regulation and EC heat release—are not mutually-exclusive operating modes of a two-phase heat-transfer system, and can therefore coexist. Furthermore, RC heat absorption and EC heat absorption are not mutually-exclusive operating modes of a two-phase heat-transfer system, and can therefore also coexist.
2. Refrigerant-controlled Heat Absorption
The purpose of RC heat absorption is usually to control the rate at which refrigerant absorbs heat in all, or in a part of, the evaporator refrigerant passages of a principal configuration at a preselected refrigerant saturated-vapor pressure or equivalently, in the case of an azeotropic-like refrigerant, at a preselected refrigerant saturated-vapor temperature. The preselected refrigerant pressure may be fixed or may change in a pre-prescribed way.
RC heat absorption is achieved with type A, or with type B, configurations by controlling the amount of liquid refrigerant in the one or more evaporator refrigerant passages of their principal configuration in pre-prescribed ways. Pre-prescribed ways for achieving heat absorption often violate self-regulation condition (A); namely they decrease the amount of liquid refrigerant in the evaporator refrigerant passages below that allowed by self-regulation condition (A).
In the particular case where an evaporator is a split evaporator including several component evaporators (see section V,B,12), liquid refrigerant can be inserted into, and extracted from, component evaporators independently by using several ancillary configurations, or even by using only a single ancillary configuration.
The purpose of liquid-refrigerant injection is to achieve at least one of several objectives. These objectives include (1) preventing refrigerant vapor being trapped in one or more parts of the evaporator refrigerant passages of a system of the invention, thereby eliminating potential hot spots; (2) increasing the refrigerant's heat-transfer coefficients in the system's evaporator refrigerant passages; and (3) increasing the refrigerant's critical flux in those passages.
To achieve one or more of the foregoing several objectives, the systems of the invention use liquid-refrigerant injectors, or more briefly LR injectors, to inject liquid refrigerant into the system's evaporator refrigerant passages. LR injectors are usually passive devices having one or more orifices whose total cross-sectional area is smaller than the cross-sectional area of the inlet through which liquid refrigerant is supplied to them. LR injectors achieve their objectives by one or more of several techniques. These techniques include (1) promoting turbulence in evaporator refrigerant passages; (2) distributing liquid refrigerant in preselected vapor spaces inside evaporator refrigerant passages; and (3) distributing liquid refrigerant over preselected internal surfaces of those passages.
FIG. 24 and
FIG. 74 and
FIG. 83 and
The symbol ‘⊙’ used in certain FIGURES denotes that the signal represented by a letter with one or more superscripts which include a ‘dash’, and with one or more subscripts, is transmitted (1) from a transducer to a control unit, where the arrow associated with the signal points toward the signal, and (2) from a control unit to a control table element or means—such as a pump or a valve—where the arrow associated with the signal points away from the symbol. And a first of the two symbols
inside the block representing a heat exchanger, represents the one or more refrigerant passages of the heat exchanger; and a second of the two symbols
inside the block representing a heat exchanger, represents the one or more fluid ways of the heat exchanger.
Several numerals occur often in the FIGURES. Elements designated by certain of those numerals are listed for convenience below.
The optimal number and kind of airtight configurations used in a system of the invention, the desired properties of those configurations, and the particular refrigerant—and where applicable inert-gas—control techniques employed to achieve those properties, depend on the particular heat-transfer application considered. It follows that the best mode for carrying out the invention, namely the preferred embodiment of a system of the invention, depends on the particular heat-transfer application considered.
In this part (part V) of this DESCRIPTION I first describe principal, ancillary, and IG, configurations suitable for various preferred embodiments of the invention, and then give examples of those embodiments in the context, for specificity, of a particular category of applications. Each of these embodiments is expected to be a preferred embodiment for some specific useful application. The statements made about the principal configurations of airtight configurations apply to the refrigerant-circuit configurations of evacuated configurations.
All principal configurations include only one refrigerant principal circuit. A refrigerant principal circuit includes, by definition, the one or more refrigerant passages of an evaporator, the one or more refrigerant passages of a condenser, means for transferring refrigerant vapor exiting the one or more refrigerant passages of an evaporator to the one or more refrigerant passages of a condenser, and means for transferring liquid refrigerant exiting the one or more refrigerant passages of a condenser to the one or more refrigerant passages of an evaporator. The refrigerant-vapor transfer means may transfer in part, or even over its entire length, only liquid refrigerant under certain special operating conditions. And the principal configuration may also include refrigerant auxiliary circuits around which only liquid refrigerant circulates.
Almost all principal configurations of preferred embodiments of the invention can be divided into twelve groups designated by roman numerals I to XII. In grouping principal configurations of the invention, I distinguish between
I shall hereinafter refer to the former kind of evaporators as ‘pool evaporators’, or more briefly as ‘P evaporators’; and to the latter kind of evaporators as ‘non-pool evaporators’, or more briefly as ‘NP evaporators’.
Most P evaporators have a single-level liquid-vapor interface surface while they are active as well as while they are inactive. However, P evaporators also include evaporators which have a multi-level liquid-vapor interface surface while they are active. Electrode-type electric steam boilers are examples of P evaporators having a two-level liquid-vapor interface surface while they are active.
I note that, by definition, group I to VI configurations have NP evaporators and group VII to XII configurations have P evaporators. I also note that NP and P evaporators may have a single, or may have several, bottom, top, or multi-level refrigerant inlet ports, and/or several bottom, top, or multi-level refrigerant outlet ports.
I further note that, in classifying principal configurations belonging to a given group, I distinguish between configurations having a preheater and those having no preheater, and between configurations having a superheater and those having no superheater, only if the principal configurations have an evaporator refrigerant auxiliary circuit. I therefore, for simplicity, show no preheater and no superheater in FIGURES used in classifying principal configurations and having no evaporator refrigerant auxiliary circuit.
Examples of known P evaporators are, in the steam-generating industry, fire-tube steam boilers, cast-iron steam boilers, resistance-type electric steam boilers, and electrode-type electric steam generators; and, in the refrigeration industry, flooded shell-and-tube coolers and flooded evaporators. And examples of known NP evaporators are, in the steam-generating industry, water-tube steam boilers and coil-type steam boilers; and in the refrigeration industry, direct-expansion air-cooled evaporators, direct-expansion shell-and-tube coolers, direct-expansion shell-and-coil coolers, tube-in-tube coolers, plate coolers, and Baudelot coolers.
By contrast with evaporators, I shall not distinguish, in grouping principal configurations, between condensers in which
I note that certain condensers, such as shell-and-tube condensers, in which refrigerant flows through the space between the shell and the tubes, can be used for storing liquid refrigerant without flooding or submerging even part of the condensers’ heat transfer surfaces, and can therefore also perform the function of a 2-port or feed-through receiver, one of the ports of the 2-port receiver being the condensers' horizontal cross-section just below their lowest heat-transfer surface. Thus the receiver (of a principal configuration) may be an integral part of a condenser. I also note that, in classifying principal configurations belonging to the same group, I do not distinguish between principal configurations having a desuperheater and those having no desuperheater. I therefore, for simplicity, show no desuperheater in the FIGURES used in grouping principal configurations.
1. Group I Configurations
The key distinctive characteristic of group I configurations, compared to other groups of configurations with an NP evaporator, is that they have no auxiliary circuit.
I distinguish between group I configurations having a refrigerant pump and those that have no refrigerant pump; and designate the former subgroup of configurations by the symbol IF and the latter subgroup of configurations by the symbol IN, where the subscripts ‘F’ and ‘N’ stand, respectively, for forced refrigerant circulation and natural refrigerant circulation. I also distinguish between group I configurations having a subcooler and group I configurations having no subcooler. However, I do not distinguish between group I configurations having a preheater and those having no preheater (or between group I configurations having a superheater and those having no superheater).
I use a superscript to indicate the absence or the presence of a subcooler. Thus the symbols IFo and INo designate classes of group I configurations with no subcooler and the symbols IFs and INs designate classes of group I configurations with a subcooler.
A class IFo configuration, with a 2-port or feed-through receiver, is shown in FIG. 1. NP evaporator 1, hereinafter referred to as evaporator 1, has a refrigerant inlet 2 and a refrigerant outlet 3; condenser 4 has a refrigerant inlet 5 and a refrigerant outlet 6; 2-port condensate receiver 7 has an inlet 8 and an outlet 9; refrigerant pump 10 has an inlet 11 and an outlet 12; and refrigerant circulates around refrigerant principal circuit 2-3-5-6-8-9-11-12-2 primarily under the a 10. A class IFo configuration may have a 1-port receiver instead of a 2-port receiver as shown in
A class IFs configuration with a 2-port receiver is shown in
2. Group II Configurations
The key distinctive characteristic of group II configurations, compared to other groups of configurations with an NP evaporator, is that they have a separator and a single refrigerant auxiliary circuit of the kind named a type 1 evaporator refrigerant auxiliary circuit, (and therefore, in particular, have no subcooler refrigerant auxiliary circuit). Group II configurations may have no refrigerant pump, a CR pump, an EO pump, or both a CR pump and an EO pump.
I distinguish between group II configurations having a refrigerant pump, and those that have no refrigerant pump and are designated by the symbol IINN. (In the symbol IINN, the first subscript indicates natural refrigerant circulation around the refrigerant principal circuit, and the second subscript indicates natural refrigerant circulation around their evaporator refrigerant auxiliary circuit.)
I use the symbol IIFN to designate the subgroup of group II configurations in which the refrigerant circulates around their refrigerant principal circuit primarily under the forced action of a CR pump, and around their evaporator refrigerant auxiliary circuit solely under the combined natural action of gravity and heat absorbed from the evaporator's heat source. I also use the symbol IIFF to designate the subgroup of group II configurations in which their refrigerant circulates around the refrigerant principal circuit primarily under the forced action of a CR pump, and around their evaporator refrigerant auxiliary circuit primarily under the forced action of an EO pump. I further use the symbol IINF to designate the subgroup of group II configurations in which the refrigerant circulates around the refrigerant principal circuit solely under the combined natural action of gravity and heat absorbed from a heat source, and around the refrigerant auxiliary circuit primarily under the forced action of an EO pump.
I use a first superscript to indicate the absence or the presence of a subcooler; a second superscript to indicate the presence or absence of a superheater; and a third superscript to indicate the absence or presence of a preheater. In the case of the first superscript, a ‘o’ (zero), an ‘s’, an ‘s′’, and an ‘s″’, indicate that group II configurations, designated by the symbols with these superscripts, have respectively
A class IIFNooo configuration with a 3-port (type 1) separator and a 2-port receiver is shown in FIG. 3. Type 1 separator 21 has a vapor inlet 22 connected to evaporator refrigerant outlet 3, vapor outlet 23 connected to condenser refrigerant inlet 5, and liquid port 24 connected to node or mergence point 25 at some point along refrigerant line 12-2. Refrigerant circulates around the refrigerant principal circuit 2-3-22-23-5-6-8-9-11-12-25-2 primarily around the evaporator refrigerant auxiliary circuit 2-3-22-24-25-2 solely under the combined action of gravity and heat absorbed from a heat source (not shown). A class IIFNooo configuration with a 4-port (type 1) separator is shown in FIG. 4. In this case, separator 21 has a liquid inlet 26—in addition to vapor inlet 22, vapor outlet 23, and liquid port 24—and refrigerant-pump outlet 12 is connected to liquid inlet 26 instead of to a point along refrigerant line 12-2 as shown in FIG. 3. Whereas the evaporator refrigerant auxiliary circuit in the case of a 4-port separator is—except for the absence of node 25—the same as that for a 3-port separator, the refrigerant principal circuit in the case of a 4-port separator also includes liquid inlet 26 and liquid port 24 so that refrigerant flows (under steady state conditions) primarily under the action of pump 10 around refrigerant principal circuit 2-3-22-23-5-6-8-9-11-12-26-24-2. A class IIFNooo configuration with 3-pot separator and a 1-port receiver is shown in
A class IINFooo configuration with a 3-port separator and a 2-port receiver is shown in FIG. 6. This configuration differs from that shown in
A class IIFFs″sp″ configuration with a 2-port receiver is shown in FIG. 7. Class IIFFs″sp″ configurations differ from class IIFNooo configurations with a 3-port separator by the addition, in the manner shown in
Class IINNs″sp″ configurations read on
I note that, in the refrigerant-circuit configuration shown in
3. Group III Configurations
The key distinctive characteristic of group III configurations, compared to other groups of configuration with an NP evaporator, is that they have a separator and a single refrigerant auxiliary circuit of the kind named a type 2 evaporator refrigerant auxiliary circuit, (and therefore, in particular, have no subcooler refrigerant auxiliary circuit). Group III configurations have a DR pump only, or a DR pump and a CR pump.
I use the symbol IIIFN to designate group III configurations having no CR pump, and the symbol IIIFF to designate group III configurations having a CR pump.
I distinguish between four classes of subgroup IIIFN configurations, and use the symbols IIIFNoo, IIIFNso, IIIFNos, and IIIFNss, to designate these four classes. In the last four symbols, the subscript F is used to indicate that refrigerant circulates around both the refrigerant principal circuit, and around the evaporator refrigerant auxiliary circuit, under the forced action of a DR pump; and the first and second superscripts are used to indicate the absence or presence of a subcooler and a superheater, respectively. I also distinguish between type 2 separators used in group III configurations and type 1 separators used in group II configurations because the former separators perform a significantly different function from the latter; and can, in particular, also perform the function of a receiver. However, I do not distinguish between subgroup IIIFN configurations having a preheater and those having no preheater.
A class IIIFNooo configuration with a 3-port (type 2) separator and no separate receiver is shown in FIG. 8. Type 2 separator 42 has a vapor inlet 43 connected to evaporator refrigerant outlet 3, a vapor outlet 44 connected to condenser refrigerant inlet 5, and a liquid outlet 45. DR pump 46 has an inlet 47 connected to condenser refrigerant outlet 6 and an outlet 48 connected to evaporator refrigerant inlet 2; and separator liquid outlet 45 is connected to refrigerant line 6-47 at point 49. Refrigerant circulates around refrigerant principal circuit 2-3-43-44-5-6-49-47 around evaporator refrigerant auxiliary circuit 2-3-43-45-49-47-48-2, primarily under the force action of DR pump 46. A class IIIFNooo configuration with a 4-port separator having a liquid inlet 50 is shown in FIG. 8A.
A class IIIFNss configuration with a 3-port separator and no separate receiver is shown in
I note that, in the class IIIFNss configuration shown in
A class IIIFFss configuration with a 3-port separator and a receiver is shown in FIG. 9A. Class IIIFFso configurations are obtained from
4. Group IV Configurations
The key distinctive characteristic of group IV configurations, compared to other groups of configurations with an NP evaporator, is that they have a single refrigerant auxiliary circuit of the kind named a subcooler refrigerant auxiliary circuit which, by definition, always includes the one or more refrigerant passages of a subcooler, and the one or more refrigerant passages of a pump, and which may also include the one or more refrigerant passages of a preheater; but which always excludes the one or more refrigerant passages of an evaporator, and the one or more refrigerant passages of a condenser. Broadly speaking, group IV configurations are combinations of a group I configuration with a subcooler refrigerant auxiliary circuit.
I use the symbols IVFF, IVFF*, IVF*F, and IVNF, where the subscript ‘F*’ denotes the presence of an HF pump, to designate subgroups of group IV configurations with respectively
A class IVFFs configuration with a 2-port receiver is shown in FIG. 10. Subcooler 57 has a refrigerant inlet 58 and a refrigerant outlet 59; preheater 60 has a refrigerant inlet 61 and a refrigerant outlet 62; and refrigerant circulates, under the forced action of SC pump 63, around subcooler refrigerant auxiliary circuit 66-58-59-61-62-67-64-65-66, where 64 and 65 are the inlet and outlet, respectively, of SC pump 63, and where node 66 is located along the refrigerant line connecting CR pump outlet 12 to subcooler refrigerant inlet 58, and where node 67 is located along the refrigerant line connecting preheater refrigerant outlet 62 to evaporator refrigerant inlet 2. A class IVFFs configuration can be looked at as a class IFs configuration to which has been added a subcooler refrigerant auxiliary circuit whose subcooler and preheater refrigerant passages are a part of the configuration's refrigerant principal circuit.
A class IVFFo configuration is obtained by deleting subcooler 18 from a class IVFFs configuration; and class IVNFs and IVNFo configurations are obtained by deleting SC pump 63 from respectively class IVFFs and IVFFo configurations.
A subgroup IVFF* configuration differs from a subgroup IVFF configuration in that SC pump 63 is replaced, in the manner shown in
Subgroup IVNF configurations are obtained by deleting CR pump 10 from sub-group IVFF configurations. Refrigerant outlet 6 of condenser 4 must be no lower than refrigerant inlet 2 of evaporator 1 in all group IV configurations having no CR pump. This is a necessary and not a sufficient requirement for operability. (In fact, the requirements for operability on the height of outlet 6 are more complex in group IV configurations with no CR pump than in group II configurations with no CR pump and no EO pump.)
Examples of group IV configurations having a subcooler refrigerant auxiliary circuit with no preheater refrigerant passages are obtained by deleting preheater 60 in
5. Group V Configurations
The key distinctive characteristic of group V configurations, compared to other groups of configurations with an NP evaporator, is that they have, in addition to a subcooler refrigerant auxiliary circuit, a type 1 evaporator refrigerant auxiliary circuit. Broadly speaking, group V configurations are combinations of group II configurations with a subcooler refrigerant auxiliary circuit which may include the one or more refrigerant passages of a preheater.
I distinguish between group V configurations with a subcooler refrigerant auxiliary circuit having a subcooler whose refrigerant passages are a part of the configurations' refrigerant principal circuit (as well as of the subcooler refrigerant auxiliary circuit) and group V configurations with a subcooler refrigerant auxiliary circuit having a subcooler whose refrigerant passages are not a part of the configurations' refrigerant principal circuit. I shall refer to the former subcooler refrigerant auxiliary circuit as an ‘interactive-type subcooler refrigerant auxiliary circuit’, or more briefly as an ‘I-type subcooler refrigerant auxiliary circuit’; and to the latter subcooler refrigerant auxiliary circuit as a ‘non-interactive-type subcooler refrigerant auxiliary circuit’, or more briefly as a ‘NI-type subcooler refrigerant auxiliary circuit’. Group V configurations with an I-type subcooler refrigerant auxiliary circuit have a 3-port (type 1) separator and group V configurations with an NI-type subcooler refrigerant auxiliary circuit have either a 5-port (type 1) or a 6-port (type 1) separator.
I use a first superscript to indicate the absence or presence of a subcooler, other than a subcooler having one or more refrigerant passages that are a part of the subcooler refrigerant auxiliary circuit; a second superscript to indicate the absence or presence of a superheater; and a third superscript to indicate the presence or absence of a preheater other than a preheater having one or more refrigerant passages that are a part of the subcooler refrigerant auxiliary circuit. In the case of the first superscript, a ‘o’ (zero), an ‘s’, an ‘s′’, and an ‘s″’, indicate that group V configurations, designated by the symbols with these superscripts, have respectively
In the case of group V configurations with an I-type subcooler refrigerant auxiliary circuit, I use the symbols VFF, VFF*, VF*F, and VNF, to designate subgroups of group V configurations with respectively
A class VFFFs″sp′ configuration with an I-type subcooler refrigerant auxiliary circuit and with a 2-port receiver is shown in
A class VFF*Fs″sp′ configuration differs from a class VFFFs″sp′ configuration in that SC pump 63 is replaced, in the manner shown in
Sub-subgroup VFFN, VFF*N, and VF*FN, configurations are obtained by deleting EO pump 27 from subgroup VFFF, VFF*F, and VF*FF, configurations, respectively; subgroup VNFF and VNF*F configurations are obtained by deleting CR pump 10 from subgroup VFFF and VFF*F configurations, respectively; and subgroup VNFN configurations are obtained by deleting EO pump 27 from subgroup VNFF configurations. (Refrigerant outlet 6 of condenser 4 must be no lower than refrigerant inlet 2 of evaporator 1 in all group V configurations that do not have a CR pump or an HF pump.)
Examples of group V configurations having an I-type subcooler refrigerant auxiliary circuit with no preheater refrigerant passages are obtained by deleting preheater 60 in
In the case of group V configurations with an NI-type subcooler refrigerant auxiliary circuit, I use the symbols VFF and VNF to designate subgroups of group V configurations with a CR pump, and no CR pump, respectively; the symbols VFFF and VFFN to designate subgroups of subgroup VFF configurations with an EO pump, and no EO pump, respectively; and the symbols VNFF and VNFN to designate subgroups of subgroup VNF configurations with an EO pump, and no EO pump, respectively.
A class VFFNssp configuration with an NI-type subcooler refrigerant auxiliary circuit, a 6-port (type 1) separator, and a 2-port receiver, is shown in
Examples of group V configurations having an NI-type subcooler refrigerant auxiliary circuit with no preheater are obtained by deleting preheater 74 in
6. Group VI Configurations
The key distinctive characteristic of group VI configurations, compared to other groups of configurations with an NP evaporator, is that they have, in addition to a subcooler refrigerant auxiliary circuit, a type 2 evaporator refrigerant auxiliary circuit. Broadly speaking, group VI configurations are combinations of group III configurations with a subcooler refrigerant auxiliary circuit.
I distinguish—as in the case of group V configurations—between group VI configurations with an I-type subcooler refrigerant auxiliary circuit and group VI configurations with an NI-type subcooler refrigerant auxiliary circuit. Group VI configurations with an I-type subcooler refrigerant auxiliary circuit may—unlike group V configurations with an l-type subcooler refrigerant auxiliary circuit—have a 4-port (type 2) separator as well as a 3-port (type 2) separator; and group VI configurations with an NI-type subcooler refrigerant auxiliary circuit may—like group V configurations with an NI-type subcooler refrigerant auxiliary circuit—have either a 5-port (type 2) separator or a 6-port (type 2) separator. However, the differences between group VI configurations with 3-port and 4-port separators, and between group VI configurations with 5-port and 6-port separators, are only minor; and therefore only 3-port separator and 5-port separator group VI configurations are shown. (4-port separator group VI configurations and 6-port separator group VI configurations can be deduced easily respectively from the three 3-port separator group VI configurations shown (see
In the case of group VI configurations with an I-type subcooler refrigerant auxiliary circuit, I use the symbol VIFF to designate group VI configurations with a DR pump and an SC pump; the symbol VIFF* to designate the subgroup of group VI configurations with a DR pump and an HF pump, and the symbol VIF*F to designate the subgroup of group VI configurations with an HF pump and an SC pump. I use a first superscript to indicate the absence or presence of a subcooler, other than a subcooler having one or more refrigerant passages that are a part of the subcooler refrigerant auxiliary circuit; and a second superscript to indicate the absence or presence of a superheater. In the case of the first superscript, a ‘o’ (zero), and an ‘s’ indicate that group VI configurations, designated by the symbols with these superscripts, have respectively
A class VFFss configuration with a 3-port separator and a 2-port receiver is shown in
A class VIFF*ss configuration differs from a class VIFFss configuration in that SC pump 63 is replaced, in the manner shown in
Subgroup VINF configurations are obtained by deleting DR pump 46 from subgroup VIFF configurations.
Examples of group VI configurations having an I-type subcooling refrigerant auxiliary circuit with no preheater refrigerant passages are obtained by deleting preheater 60 from
In the case of group VI configurations with an NI-type subcooler refrigerant auxiliary circuit, there exist only subgroup VIFF configurations.
A class VIFFss configuration with an NI-type subcooler refrigerant auxiliary circuit and a 5-port separator is shown in FIG. 9B. This configuration can be looked at as a class IIIFFss configuration in which the 3-port separator has been replaced by a 5-port separator and to which NI-type subcooler refrigerant auxiliary circuit 79-72-73-64-65-75-76-80 has been added, where numerals 79 and 80 designate respectively a liquid-refrigerant outlet and a liquid-refrigerant inlet of separator 42.
Examples of group VI configurations having an NI-type subcooler refrigerant auxiliary circuit with no preheater refrigerant passages are obtained by deleting preheater 74 in
7. Group VII and VIII Configurations
Group VII and VIII configurations are derived from respectively group I to VI configurations by replacing the NP evaporator in the latter configurations by a P evaporator. Thus, for example, a class VIIFs configuration is a class IFs configuration in which NP evaporator 1 has been replaced by P evaporator 81 (see FIGS. 16 and 16A), and a class XFFs configuration is a class IVFFs configuration in which NP evaporator 1 has been replaced by P evaporator 81 (see FIG. 16B). Numeral 123 designates the refrigerant liquid-vapor interface inside a P evaporator.
8. Group II*, III*, V*, VI*, VIII*, IX*, XI*, and XII*, Configurations
Group II*, V*, VIII*, and XI*, configurations are, by definition, principal configurations derived from respectively group II, V, VIII, and XI, configurations by replacing type 1 separator 21 by type 1 separating assembly 21*; and group III*, VI*, IX*, and XII*, configurations are, by definition, principal configurations derived from respectively group III, VI, IX, and XII, configurations by replacing type 2 separator 42 by type 2 separating assembly 42*, and by adding a receiver whenever the four last-cited groups have no receiver and a receiver is required. (A receiver is usually required unless condenser 4 can also perform the function of a receiver. An example of such a condenser is a shell-and-tube condenser with refrigerant outside its tubes.) Thus, for example, a class VIII*FNooo configuration is a class VIIIFNooo configuration in which separator 21 has been replaced by separating assembly 21* (see FIG. 17); and a class IX*FNoo configuration is a class IXFNoo configuration in which separator 42 has been replaced by separating assembly 42* (see FIG. 18), and to which—where the class IXFNoo configuration has no receiver—a receiver has been added. (The receiver may be a 1-port or a 2-port receiver.)
However, whereas in symbols designating classes belonging to group III, VI, IX, and XII, configurations, the first superscript is either a ‘o’ (zero) or an ‘s’; in symbols designating classes belonging to group III*, VI*, IX*, and XII*, configurations the first superscript can, in addition to a ‘o’ or an ‘s’, also be an ‘s′’, an ‘s″’, or an ‘s′″’. A ‘o’ indicates classes belonging to group III* and IX* configurations having no subcooler; and classes belonging to group VI* and XII* configurations having no subcooler other than a subcooler whose one or more refrigerant passages are a part of the configurations' subcooler refrigerant auxiliary circuit. An ‘s’ indicates classes belonging to group III* and IX* configurations having a subcooler whose one or more refrigerant passages are a part of the configurations' principal refrigerant circuit and evaporator refrigerant auxiliary circuit, and of no other refrigerant circuit; and classes belonging to group VI* and XII* configurations having a subcooler whose one or more refrigerant passages are a part of the configurations' principal refrigerant circuit and evaporator refrigerant auxiliary circuit, and of no other refrigerant circuit other than a subcooler whose one or more refrigerant passages are a part of the configurations' subcooler refrigerant auxiliary circuit. An ‘s′’ indicates classes belonging to group III* and IX* configurations having a subcooler whose one or more refrigerant passages are part of the configurations' one or more evaporator refrigerant auxiliary circuits and of no other refrigerant circuit; and classes belonging to group VI* and XII* configurations having a subcooler whose one or more refrigerant passages are part of the configurations' one or more evaporator refrigerant auxiliary circuits, and of no other refrigerant circuit other than a subcooler whose one or more refrigerant passages are a part of the configurations' subcooler refrigerant auxiliary circuit. An ‘s″’ indicates classes belonging to group III* and IX* configurations having a first subcooler whose one or more refrigerant passages are part of the configurations' refrigerant principal circuit and of the configurations' evaporator refrigerant auxiliary circuit; and a second subcooler whose one or more refrigerant passages are part of the evaporator refrigerant auxiliary circuit, and of no other refrigerant circuit; and classes belonging to group VI* and XII configurations having a first subcooler whose one or more refrigerant passages are part of the configurations' refrigerant principal circuit and of the configurations' evaporator refrigerant auxiliary circuit; and a second subcooler whose one or more refrigerant passages are part of the evaporator refrigerant auxiliary circuit, and of no other refrigerant circuit other than a subcooler whose one or more refrigerant passages are a part of the configurations' subcooler refrigerant auxiliary circuit. And an ‘s′″’ indicates classes belonging to group III* and IX* configurations having a subcooler whose one or more refrigerant passages are part of the refrigerant principal circuit, and of no other refrigerant circuit; and classes belonging to group VI* and XII* configurations having a subcooler whose one or more refrigerant passages are part of the refrigerant principal circuit, and of no other refrigerant circuit other than a subcooler whose one or more refrigerant passages are a part of the configurations' subcooler refrigerant auxiliary circuit.
9. General Remarks on Principal Configurations
In
10. Specialized Configurations
The present invention includes, in addition to the groups of principal configurations discussed in section V, several specialized groups of principal configurations which may be preferred for certain special applications.
A first specialized group of principal configurations consists of principal configurations having a type 1′ separator. The principal configuration shown in
A second group of specialized principal configurations consists of principal configurations having an upper special header through which liquid refrigerant is distributed to the one or more refrigerant passages of their evaporator, the special header being, under most operating conditions, filled only partially with liquid refrigerant. The principal configuration shown in
A third specialized group of principal configurations consists of principal configurations having a liquid-refrigerant auxiliary transfer means for transferring by gravity liquid refrigerant in a pool evaporator to the liquid-refrigerant principal transfer-means segment upstream from the principal configuration's refrigerant principal pump. Three examples of such specialized configurations are shown in
11. Integral Evaporator-separator Combinations
The principal configurations of the invention include configurations in which a type 1 or a type 1′ separator is physically an integral part of an NP evaporator. Any integral evaporator-separator combination, employed in conventional (namely in airtight) steam generators and in refrigeration equipment, can also be employed in the airtight configurations of the invention—provided the evaporator-separator combination used is constructed with materials and joining techniques compatible with the refrigerant employed and suitable for airtight configurations. Examples of evaporator-separator combinations range from an evaporator-separator combination having a single evaporator refrigerant passage, and a separator whose separator vessel is a small sphere, to an evaporator-separator combination having, like the so-called four-drum Stirling-type boilers, hundreds of refrigerant passages. I give here just enough examples of evaporator-separator combinations to show how they fit into the principal configurations. I use in these examples a class IIFNooo configuration with a 2-port receiver and certain principal configurations with a type 1′ separator; but other—although not all—principal configurations with a type 1, or a type 1′, separator could also have been used.
The integral evaporator-separator combination shown in
In
12. Component Heat Exchangers, Component Heat Sources, and Component Heat Sinks
Each of the heat exchangers represented by a rectangle in the FIGURES may be a ‘unitary heat exchanger’ consisting, by definition, of a single unit; or may be a ‘split heat exchanger’ that includes, by definition, several separate and physically-distinct heat-exchanger units I shall hereinafter refer to as ‘component heat exchangers’. Component heat exchangers of the same split heat exchanger may have their refrigerant passages connected in series, in parallel, or both in series and in parallel; the refrigerant passages of all component heat exchangers of a given split heat exchanger constituting the split heat exchanger's refrigerant passages. In the particular case where a heat exchanger is a hot heat exchanger, a cold heat exchanger, an evaporator, a preheater, a superheater, a condenser, a subcooler, and a desuperheater, I shall refer to the heat exchanger's component heat exchangers respectively as ‘component hot heat exchangers’, ‘component cold heat exchangers’, ‘component evaporators’, ‘component preheaters’, ‘component superheaters’, ‘component condensers’, ‘component subcoolers’, and ‘component desuperheaters’.
A heat source of a given split hot heat exchanger may be either a ‘unitary heat source’, consisting, by definition, of a single not readily-divisible heat source; or may be a split heat source consisting of several readily-distinguishable component heat sources. A unitary hot heat exchanger has almost always a unitary heat source, but a split hot heat exchanger may have either a unitary heat source or a split heat source. In the former case all the component heat exchangers of the split hot heat exchanger have the selfsame heat source, whereas in the latter case at least two of the component heat exchangers of a split hot heat exchanger have readily-distinguishable component heat sources of the split heat source. Similarly, a heat sink of a given split cold heat exchanger may be either a ‘unitary heat sink’, consisting, by definition, of a single not readily-divisible heat sink; or may be a split heat sink consisting of several readily-distinguishable component heat sinks. A unitary cold heat exchanger has almost always a unitary heat sink, but a split cold heat exchanger may have either a unitary heat sink or a split heat sink. In the former case all the component heat exchangers of the split cold heat exchanger have the selfsame heat sink, whereas in the latter case at least two of the component heat exchangers of a split cold heat exchanger have readily-distinguishable component heat sources of the split heat sink.
I note that in certain embodiments of the invention the selfsame heat exchanger is under certain operating conditions a hot heat exchanger, and is under certain other operating conditions a cold heat exchanger.
13. Compoment Seperating Devices, Component Receivers, Component Refrigerant Pumps, and Component Refrigerant Valves
Briefly—as in the case of heat exchangers—separating devices, receivers, refrigerant pumps, refrigerant valves, and other elements of airtight configurations, may be a ‘unitary element’ consisting, by definition, of a single unit; or may be a ‘split element’ that includes, by definition, several separate and physically-distinct units I shall hereinafter refer to, in general, as ‘component elements’ and for example specifically as ‘component separating devices’, ‘component receivers’, ‘component refrigerant pumps’, and ‘component refrigerant valves’. In particular, a refrigerant principal circuit, or a refrigerant auxiliary circuit, may include one or more refrigerant-circuit segments with several sets of parallel branches. Each branch of a set of parallel branches may, for example, include a component preheater, a component NP evaporator, and a component separating device; and another set of parallel branches may include a component condenser, a component receiver, and a component subcooler. Furthermore, a refrigerant principal circuit, or a refrigerant auxiliary circuit of a principal configuration may have sub-branches merging into branches which in turn merge into a single refrigerant-circuit segment. Thus, for example, a single principal configuration may—as in a system reduced to actual practice by S. Molivadas—have sixteen parallel branches, each of which contains four component NP evaporators, connected in series to four parallel branches, each of which contains a component separating device and more specifically a component separator; and four parallel branches, each of which contains a component condenser.
14. Split Refrigerant Principal Configurations
All the principal configurations discussed so far have a single evaporator and a single condenser, either of which may be a unitary or a split heat exchanger. I shall refer to the foregoing principal configurations as ‘unitary principal configurations’. In certain applications, the refrigerant principal circuit of a principal configuration may consist of several branches which have either (1) a common component evaporator and different condensers, or (2) a common component condenser and different evaporators. I shall refer to the last-cited principal configurations as ‘split principal configurations’. Examples of split principal configurations are given later in this DESCRIPTION.
The branches of a spiit principal configuration have the selfsame refrigerant and a common refrigerant principal-circuit segment. However, each of these branches can often be thought of conceptually as belonging to distinct principal configurations which can be grouped and classified in the same way as unitary principal configurations.
15. Subatmospheric Airtight Configurations
I use the term ‘subatmospheric airtight configurations’ to denote airtight configurations whose refrigerant pressure always stays—except in the case of a failure—below ambient atmospheric pressure while they are active and while they are inactive.
I note that prior-art so-called vapor, vacuum, and variable-vacuum, steam systems have non-airtight configurations. Consequently, their configurations, while inactive, ingest air until their refrigerant pressure reaches ambient atmospheric pressure, and therefore this pressure does not always stay below ambient atmospheric pressure. Furthermore, all of the foregoing three prior-art systems are operated at positive as well as negative gauge pressures; typically at positive gauge pressures up to 5 psig (0.345 bar gauge) in the case of vacuum and variable-vacuum systems. It follows that the refrigerant-circuits (water-steam circuits) of the three prior-art systems cited above must use components designed to withstand internal working pressures as well as external working pressures, and to meet the requirements of applicable pressure-vessel and pressure-piping codes.
By contrast, airtight configurations having aqueous solutions as their refrigerant, and maximum heat-sink temperatures substantially below (say at least 15° C. below) the refrigerant's saturated-vapor temperature at ambient atmospheric pressure, can be operated so that their refrigerant pressure always stays below ambient atmospheric pressure. It follows that the refrigerant circuits of such configurations can be equipped with one or more pressure-relief devices that release refrigerant (into the ambient air) when their refrigerant exceeds only slightly ambient atmospheric pressure (because, for example, of a system malfunction).
Subatmospheric airtight configurations—namely airtight configurations whose refrigerant pressure always stays below ambient atmospheric pressure—are not restricted to a particular kind of refrigerant, and may use any azeotropic-like or non-azeotropic refrigerant. Neither are subatmospheric airtight configurations restricted to transferring heat to heat sinks at temperatures below the boiling point of their refrigerant at ambient atmospheric pressures. For example, a subatmospheric airtight configuration whose refrigerant is lithium can transfer heat to heat sinks well above 1000° C. even at ambient atmospheric pressures at sea level.
The refrigerant passages of subatmospheric airtight configurations need not be capable of withstanding net internal pressures. This allows heat exchangers to be made with techniques which greatly reduce their cost, and which make them immune to damage by frozen refrigerants that, like H2O, expand when they change from their liquid to their solid phase. For example, a heat exchanger can be made of two flat or tubular sheets of material—such as copper, copper-plated steel, or aluminum—joined together only around their perimeter by, for instance, brazing or welding. One or both sheets have corrugations, waffle-like patterns, or hybrid patterns, that form, when the two sheets are held against each other, a panel having refrigerant passages connected to a refrigerant inlet, and to a refrigerant outlet, located at opposite ends of the two sheets' common perimeter.
16. Principal-configuration Controllable Elements
‘Principal-configuration controllable elements’, referred to in the CLAIMS as ‘principal-configuration controllable means’, are, by definition, elements of a principal configuration controlled by the two-phase heat-transfer system to which the principal configuration belongs. Principal-configuration controllable elements include refrigerant pumps and refrigerant valves.
The control of controllable elements of an airtight configuration, and in particular of a principal configuration, of the invention may be two-step, multi-step, or proportional: usually two-step or proportional in the case of unidirectional controllable pumps, and usually three-step or proportional in the case of bidirectional controllable pumps such as certain refrigerant pumps, certain hot-fluid pumps and cold-fluid pumps, and certain other fluid pumps. I shall say that a set of one or more system-controllable elements is controlled so that the characterizing parameter controlled by the set stays close to, or tends to, a preselected value where the preselected value is a single value. Proportional control can be achieved by using a modulated analog signal or by using a modulated pulsed signal.
1. Liquid-refrigerant Reservoirs
The liquid-refrigerant (LR) reservoirs used in type A or in type B combinations can be any known kind of suitable fixed-volume reservoir, or any known kind of suitable variable-volume reservoir having an internal volume which can be changed by the two-phase heat-transfer system to which the variable-volume reservoir belongs. The word ‘suitable’ in the immediately preceding sentence denotes properties such as compatibility with the liquid refrigerant stored in an LR reservoir, and the ability to withstand the range of refrigerant pressures and temperatures over which the LR reservoir is to be used.
Examples of variable-volume reservoirs are (1) structures, such as bellows-type and bladder-type devices, and (2) combinations of a deformable and a rigid structure, such as a rigid cylinder with for instance a cylindrical or a spheroidal shape, which together form a space within which liquid refrigerant is stored. I shall refer to the former reservoirs as ‘type 1 variable-volume reservoirs’ and to the latter reservoirs as ‘type 2 variable-volume reservoirs’, but I shall designate all variable-volume LR reservoirs by the same numeral.
In the case cited under (1) in the immediately-preceding minor paragraph, an external rigid structure may have to be used to constrain lateral and/or longitudinal motions of the variable-volume reservoir. Whether or not such an external frame is necessary depends on (1) the material or materials from which the reservoir is made, and (2) the tilts and accelerations to which it will be subjected. In type IIR ancillary configurations (see section V,C,2,b,ii), the rigid structure mentioned in the immediately-preceding sentence may be provided by the mechanism used to provide the external mechanical force; and in type IIIR ancillary configurations (see section V,C,2,b,iii), the rigid structure may be provided by a rigid container in which the deformable structure is partly or entirely enclosed.
2. Types of Ancillary Configurations
a. Definitions of Type IR to VIR Configurations
Most of the (refrigerant) ancillary configurations which can be used in type A and type B combinations (of the invention) can be grouped into six general types:
b. Typical Type IR to VIR Configurations
i. Type IR Configurations
The LR reservoir shown is a type 1 bellows-type reservoir with a flexible corrugated cylindrical wall 403, but the LR reservoir could, as mentioned earlier, be any kind of variable-volume reservoir.
Numeral 404 (in
In the particular case where
ii. Type IIR Configurations
The foregoing rudimentary vise-type mechanism may be optimal in the case where the maximum absolute value of the difference (pR−pA) is a fraction of a bar and the diameter of the reservoir 401 is only a couple of centimeters. However, in cases where the maximum value of the last-cited pressure difference, or the last-cited diameter, is substantially larger, a pair of screws on opposite sides of the bellows, in a plane containing the bellows' center line, would usually be preferred. These two screws would be driven through gears, by a single motor.
In the case of a type 2 variable-volume reservoir, the position of the deformable device could be controlled by a motor driving a single screw as shown in FIG. 29. In
Screw 412 may be turned manually, instead of by an electric motor or other non-manually controlled device.
iii. Type IIIR Configurations
iv. Type IVR Configurations
v. Type VR Configurations
vi. Type VIR Configurations
c. Alternative Type IR To VIR Ancillary Configurations
One of several alternative forms of each of the type IR to VIR ancillary configurations shown in
In certain applications it may be desirable for port 407—where liquid refrigerant in the ancillary configuration merges with liquid refrigerant in the principal configuration—to be replaced (see, for example,
An example of applications where two-port ancillary configurations may be desirable are those where the preferred refrigerant is a non-azeotropic fluid such as an aqueous glycol solution. The reasons for which two-port ancillary configurations may be desirable, where the last-cited solutions are employed as a refrigerant, are given in section V,F,2.
Bidirectional LT pumps, air-transfer pumps, and hydraulic pumps, may be unavailable, or may be too costly, for the particular requirements of certain applications. Where this is true, two unidirectional LT pumps, air-transfer pumps, or hydraulic pumps, as applicable, can obviously be employed instead of a single bidirectional LT pump, air-transfer pump, or hydraulic pump, respectively.
A first alternative to employing two unidirectional LT pumps, air-transfer pumps, or hydraulic pumps (where a bidirectional LT pump, air-transfer pump, or hydraulic pump, is not available or is too costly) is to employ a single unidirectional LT pump, air-transfer pump, or hydraulic pump, in parallel with a bidirectional (two-way) valve. This alternative is shown, for the particular case of a refrigerant pump and a variable-volume LR reservoir, in
A second alternative to employing two unidirectional LT pumps, air-transfer pumps, or hydraulic pumps, is to use known means for reversing the direction of flow induced by a unidirectional LT pump, between two points. Examples of such means are described in section V,N of my co-pending U.S. patent application Ser. No. 400,738, filed 30 Aug. 1989.
In cases where liquid refrigerant leaks through bidirectional LT pump 404, unidirectional LT pump 404A, or unidirectional LT pump 404B, while it is not running, a bidirectional LT valve can be used in series with any one of the three last-cited pumps to eliminate, or to help reduce, the rate at which refrigerant leaks through each of those pumps while it is not running. Bidirectional valves can be used for a similar purpose, in series with an air-transfer pump or a hydraulic pump. The particular case where a bidirectional LT valve is used in series with a bidirectional LT pump is shown in
1. Inert-gas Reservoirs
The inert-gas (IG) reservoirs used in type B and in type C combinations (of the invention) can be any kind of suitable fixed-volume reservoir, or any kind of suitable variable-volume reservoir having an internal volume which can be changed by the two-phase heat-transfer system to which the variable-volume reservoir belongs. The word ‘suitable’, in the immediately-preceding sentence, denotes properties such as compatibility with the inert gas and with the refrigerant (which may be contained in the inert gas), and the ability to withstand the range of inert-gas pressures and temperatures over which an IG reservoir is to be used.
Variable-volume IG reservoirs may, like variable-volume LR reservoirs, be divided into type 1 variable-volume reservoirs and into type 2 variable-volume reservoirs.
2. Types of Inert-gas Configurations
a. Definitions of Type IG to VG Configurations
Most of the inert-gas (IG) configurations which can be used in type B and type C combinations can be grouped into five general types:
The five types of IG configurations listed under (a) to (e) in this section V,D,2 are usually employed to insert inert gas in, and to extract inert gas from, a principal configuration, but can also be used to insert inert gas in, and to extract inert gas from, an LR reservoir. The foregoing five types of IG configurations are described in section V,D,2,b.
b. Typical Type IG to VG Configurations
i. Type IG Configurations
GT pump 443 is a bidirectional GT pump having ports 444 and 445 through which it can induce inert-gas flow either from port 444 to port 445 or from port 445 to port 444. Alternatively, a unidirectional GT pump can be used together with means for reversing the direction of inert-gas flow between ports 444 and 445.
When GT pump 443 induces inert gas to flow from port 440 toward port 442, it will at times be mixed with a small amount of refrigerant vapor. Condensate-type refrigerant-vapor trap 446, having inlet-outlet gas port 447, and inlet-outlet gas port 448, is used to help ensure no significant amount of refrigerant vapor enters GT pump 443 and reservoir 441. To this end, trap 446 includes means for cooling, and thereby condensing, refrigerant vapor contained in inert gas. Liquid refrigerant, generated by condensation in trap 446, is returned by gravity from liquid outlet 449 of trap 446 to principal-configuration inlet 450.
ii. Type IIG Configurations
iii. Type IIIG Configurations
iv. Type IVG Configurations
v. Type VG Configurations
c. Alternative Type IG to VG Inert-gas Configurations
One of several alternative forms of each of the type IG to VG configurations shown in
In certain applications it may be desirable for port 440 to be replaced by inlet 470 (see, for example,
In certain applications bidirectional GT pumps, air-transfer pumps, and hydraulic pumps, may be unavailable, or may be too costly, for the particular requirements of those applications. Where this is true, two unidirectional GT pumps, air-transfer pumps, or hydraulic pumps, as applicable, can obviously be employed instead of a single bidirectional GT pump, air-transfer pump, or hydraulic pump, respectively.
A first alternative to employing two unidirectional GT pumps, air-transfer pumps, or hydraulic pumps (where a bidirectional GT pump, air-transfer pump, or hydraulic pump, is not available or is too costly) is to employ a single unidirectional GT pump, air-transfer pump, or hydraulic pump, in parallel with a bidirectional (two-way) gas-transfer valve, or more briefly a bidirectional GT valve. This alternative is shown, for the particular case of a GT pump and a fixed-volume IG reservoir, in
A second alternative to employing two unidirectional GT pumps, air-transfer pumps, or hydraulic pumps, is to use known means for reversing the direction of flow, induced by a unidirectional GT pump, between two points.
In cases where inert gas leaks through bidirectional GT pump 443, unidirectional GT pump 404A, or unidirectional GT pump 404B, while it is not running, a bidirectional GT valve can be used in series with any one of the three last-cited pumps to eliminate, or to help reduce, the rate at which inert gas leaks through each of those pumps while it is not running. The particular case where a bidirectional GT valve is used in series with a bidirectional GT pump is shown in
In cases where inert gas entering an IG reservoir contains some refrigerant vapor, condensed refrigerant vapor, accumulating in the IG reservoir, can be removed by providing a bidirectional drain valve, in parallel with a GT pump.
3. Condensate-type Refrigerant-vapor Traps
The complexity of the condensate-type refrigerant-vapor traps employed in inert-gas configurations depends on the particular application in which they are being used.
The condensate-type refrigerant-vapor trap shown in
Most of the refrigerant vapor—where present—in inert gas entering condenser 456 is condensed in condenser 456. The resulting liquid refrigerant is entrained by the inert gas in which the liquid refrigerant is contained toward LR diverter 462, where the entrained liquid refrigerant is diverted to liquid outlet 465. Inert gas, entering LR diverter 462 at 463, exits at 464. Residual refrigerant vapor, in inert gas exiting at 464, is condensed in condenser 459 and the resulting liquid refrigerant is returned by gravity in gas line 460-464 which has a cross-sectional area large enough for liquid refrigerant and gas to flow in opposite directions.
Condensers 456 and 459 may, for example, be air-cooled condensers, water-cooled condensers, or (liquid) refrigerant-cooled condensers. In the first case, condensers 456 and 459 may merely be a finned tube; and, in the second and third cases, condensers 456 and 459 may merely be a tube with a coil, wrapped around the tube, carrying a cold fluid, and LR diverter 462 may, for example, be a small vessel or a tee, whose ports are inlet-outlet gas ports 463 and 464, and liquid outlet 465.
Condensers 456 and 459, and LR diverter 462, may be combined into a single unit. A first example of a single-unit condensate-type refrigerant-vapor trap is shown in
I note that condenser 456 is often not necessary, and that, in this case, a principal configuration's receiver may replace LR diverter 462. Where a receiver is used also as a diverter, condenser 456, inert-gas lines 440-447-457 and 458-463, and liquid-refrigerant line 449-450, are eliminated.
4. Inert-gas Special Configurations and Inert-gas Passive Configurations
Inert-gas (IG) special configurations differ from IG configurations essentially only in that they transfer inert gas between the LR reservoir of a type VIR ancillary configuration and an IG reservoir instead of between a principal configuration and an IG reservoir. Liquid refrigerant exiting trap 446 at outlet 449 is returned by gravity either
Inert-gas passive (IGP) configurations can, like IG configurations, be one-port configurations or two-port configurations. An IGP configuration can, also like an IG configuration, have a condensate-type refrigerant-vapor trap.
I mentioned in section V,P of my U.S. patent application Ser. No. 400,738, filed 30 Aug. 1989, the need to remove a non-condensable gas, and in particular hydrogen, which may be generated inside the refrigerant passages of an airtight refrigerant configuration or of an evacuated refrigerant-circuit configuration. And I mentioned the use of membranes permeable to a non-condensable gas, but not permeable to the airtight configuration's refrigerant, as a means for getting rid of a non-condensable gas. Another means for getting rid of a particular non-condensable gas is to use, at one or more locations inside an airtight refrigerant configuration, a solid or a liquid, not miscible with the refrigerant, which will absorb that non-condensable gas; for example, to use hydrazine to absorb hydrogen. Still another means for getting rid of a non-condensable gas is to use a non-condensable-gas trap similar to that described on page 48 of NASA Technical Briefs, December 1990.
In
The immediately-following text in this minor paragraph is an excerpt from page 48 cited in the first minor paragraph of this section V,E. “The trap . . . includes a tube of stainless steel or other poorly thermally conductive material . . . attached to a tap on top of the main vapor line where the vapor flows toward the condenser. Subcooled liquid from the outlet of the condenser cools the upper end of the tube below the vapor temperature. A small fraction of the flow in the main vapor line enters the trap and travels to the upper end. There, the vapor condenses, and the liquid is returned to the main line by gravity. (In the absence of gravity, it could be returned by the capillary action of a wick.) Noncondensable gas . . . entrained in the upward flow of vapor accumulates gradually, thereby increasing the effective thermal conductance of the upper end of the trap and decreasing the temperature T1 measured by a thermocouple near the upper end. When T1 decreases to a preset differential above T2, the temperature of the incoming coolant, a solenoid valve at the upper end opens momentarily to vent the noncondensable gas.”
In
The trap shown on page 48 of the cited NASA document uses, as mentioned in the above quotation, subcooled refrigerant to condense refrigerant vapor in the trap. However, cold water can be used, instead of subcooled refrigerant. (The qualifier ‘cold’, in the immediately-preceding sentence, indicates that water, flowing in the last-cited coil, is substantially colder than refrigerant entering the trap.) Alternatively, in certain applications, where air surrounding the trap is cold enough, coil 492 in
I note that the only significant difference between non-condensable gas traps and the condensate-type refrigerant-vapor traps discussed in this DESCRIPTION is that the former traps include a unidirectional device for causing and controlling the discharge of gas into an airtight configuration's, or into an evacuated configuration's, surroundings; whereas the latter traps include no such device.
1. Preliminary Remarks
I discuss in this section V,F applications where the properties complete minimum-pressure maintenance and self regulation are required, and where refrigerant-controlled heat release, or more briefly RC heat-release, is usually also required.
Piston-engine cooling applications provide good examples of applications where a heat source (1) requires the evaporator refrigerant passages of a principal configuration to have sharp bends and non-uniform cross-sections; (2) subjects those passages to spatially highly non-uniform heat fluxes; and (3) has temperatures far above the maximum-permissible temperatures for those passages and the refrigerant in them.
By contrast, piston-engine intercooling systems provide good examples of applications where a heat source (1) does not require the evaporator refrigerant passages of a principal configuration to have sharp bends and non-uniform cross-sections; (2) does not subject those passages to spatially highly non-uniform heat fluxes; and (3) has no temperatures above the maximum-permissible temperatures for those passages and the refrigerant in them.
In sections V,F,2 and V,F,3 I describe type A combinations, and their associated control techniques, for the case where the combinations' condenser is an air-cooled condenser. The most prominent examples of piston-engine cooling and intercooling systems with air-cooled condensers are probably those installed in automobiles and trucks. However, piston-engine cooling and intercooling systems with air-cooled condensers are also suitable for other automotive vehicles such as locomotives, for certain industrial fixed installations, and for certain passenger and cargo planes.
In section V,F,4 I describe type A combinations, and their associated control techniques, for the case where the combinations' condenser is a water-cooled condenser. The most prominent examples of piston-engine cooling systems with water-cooled condensers are probably those installed in ships and motor boats, and those installed in industrial fixed installations adjacent to a large body of water such as the sea.
Because all the type A combinations discussed in this section V,F have no partial minimum-pressure maintenance, I shall for brevity refer in this section V,F to complete minimum-pressure maintenance simply as ‘minimum-pressure maintenance’. This property, as mentioned in section III,D, is achieved in type A combinations by filling completely their principal configuration with liquid refrigerant.
2. Cooling Systems with an Air-cooled Condenser
a. Cooling Systems with a Pool Evaporator
i. Refrigerant Configuration and Control System
Refrigerant vapor, generated in the evaporator and exiting at 83″, is transferred from the evaporator to type 1 separator 21 by vapor manifold 506, having four refrigerant vapor lines (see FIG. 43A). Separator 21 has a vapor inlet 22, which has four ports. Alternatively, for example, inlet 22 may have a single port. In this second case, the four vapor-lines of manifold 506 merge into a single vapor line connected to that single port. Under most operating conditions, essentially dry refrigerant vapor exits separator 21 at 23 and enters upper header 507 of air-cooled condenser 508 at 5. (I use the numeral 5 to designate the refrigerant inlet of any condenser.) Refrigerant vapor entering header 507 flows through several condenser refrigerant passages 399 and condensed refrigerant vapor, generated in passages 399, exits lower header 509 of condenser 508 at 6 and enters 2-port condensate receiver 7 at inlet 8. (I use the numeral 6 to designate the refrigerant outlet of any condenser.) Liquid refrigerant, accumulating in receiver 7, exits at outlet 9, enters inlet 11 of CR pump 10, exits outlet 12 of CR pump 10, and enters at 82′ the evaporator formed by the coolant passages of engine 500. Liquid refrigerant, separated from refrigerant vapor in separator 21, exits at liquid outlet 24 and, after by-passing refrigerant passages 399 of condenser 508, merges at 25 with liquid refrigerant exiting pump 10 at 12. Under most operating conditions, liquid refrigerant in those coolant passages forms liquid-vapor interface surface 123. Interface surface 123 may consist of several separate and distinct segments.
The class VIIIFNooo principal configuration described in the immediately preceding two minor paragraphs has a refrigerant principal circuit 82′-83″-22-23-5-6-8-9-407-11-12-25-82′ and a type 1 evaporator refrigerant auxiliary circuit 82′-83″-22-24-25-82′. The refrigerant configuration shown in
The cooling system shown in
The control system includes central control unit 513 (see FIG. 44), or more briefly CCU 513, which, on the basis of signals received from several transducers and preselected instructions stored in CCU 513, controls pump 10, pump 404, and fan 510. The particular transducers used by the system in
The signals generated by the six last-listed transducers are supplied to CCU 513 which computes, on the basis of preselected instructions stored in CCU 513, see
The control system also includes Minimum-Pressure Maintenance Control Unit 518, or more briefly MPMCU 518, see FIG. 45. This unit operates only while engine 500 in
ii. Unsafe and Safe States
I shall say that a piston-engine cooling system is in an ‘unsafe state’, when running the engine being cooled by the system is unsafe in the sense that the engine could be damaged, by inadequate cooling, if it started running, or if it continued running. And I shall say that a piston-engine cooling system is in a ‘safe state’ when running the engine being cooled by the system is safe in the sense that the engine would not be damaged by inadequate cooling if it started running, or if it continued running. More precisely, I shall say that the system is in an unsafe state when any one of the following four relations is true:
LP<LP,SAFE; LR<LR,SAFE; pR>pR,SAFE; and TR>TR,SAFE; (1), (2), (3), (4)
and that the system is in a safe state when all of the following four relations are true:
LP≧LP,SAFE; LR≧LR,SAFE; pR≦pR,SAFE; and TR≦TR,SAFE. (5), (6), (7), (8)
Symbols LC, LR, LP, pR, and TR, were defined earlier in this section V,F,2,a. The remaining symbols in the last eight relations are defined next: the symbol LP,SAFE denotes the minimum value of LP at which the engine should be allowed to run; the symbol LR,SAFE denotes the minimum value of LR for which the cooling system's refrigerant pump does not cavitate significantly; and symbols pR,SAFE and TR,SAFE denote the maximum values of pR and TR, respectively, at which the engine should be allowed to run. (Although condition (6) would not damage the engine directly, it would usually do so indirectly in the sense that it would soon cause the value of LP to fall below LP,SAFE.)
iii. Typical Operating Method
I now outline a typical method of operating the system shown in
I start at an instant in time when the engine being cooled by the system is not running and is started, say, by an operator manually. When the engine is started, CCU 513 and all its associated transducers and controllable elements are energized, if they are not already energized.
CCU 513, as soon as it is energized, and subsequently at frequent preselected periodic time intervals while it remains energized, performs a system safety check to determine whether the system is in a safe state. If it is not, an audible and/or visual warning signal is generated to indicate that the system is in an unsafe state, and the engine, after being stopped by the operator, is inhibited from being started. If the unsafe state has occurred because pR or TR, or both, have exceeded their safe values, CCU 513 runs fan 510 at its maximum capacity until their safe values are no longer exceeded, and then de-energizes itself automatically. Thereafter MPMCU 518, which is always energized while the system is in a safe state, remains energized and controls LT pump 404 in the same way as in control mode 0. (See next major paragraph.) If the system has become unsafe because of an insufficient refrigerant charge, MPMCU 518 will de-energize itself automatically. (The refrigerant charge is insufficient when relation (1) or (2) is satisfied.)
I shall describe the operation of systems of the Invention, while they are in their safe state, in terms of ‘control modes’ and ‘transition rules’ between control modes (see definitions 115 and 116 in section III,A). In
A first mode, mode 0, of the four different control modes, is used to achieve minimum-pressure maintenance.
A second control mode, mode 1, is used, in the case of a non-azeotropic refrigerant, to achieve quasi-uniform refrigerant-component concentrations after the refrigerant temperature TR falls below a preselected temperature TR,MIN, which is (1) lower than the refrigerant's lowest saturated-vapor temperature, while the system's principal configuration is active, and which is (2) higher than the freezing temperature of the refrigerant-component with the highest freezing temperature. The elapsed time Δt, from the instant at which TR falls below TR,MIN, is determined by a clock, usually a software clock incorporated in CCU 513. This clock is stopped and reset after a preselected time interval unless the engine is running or starts running. If the engine was stopped and starts running before Δt is equal to the preselected time interval, the clock is stopped and reset at the instant the engine starts running.
A third control mode, mode 2, is used to achieve refrigerant-controlled heat release, or more briefly RC heat release, which is the particular form of Internally-controlled heat release, or more briefly IC heat release, used in type A combinations.
A fourth control mode, mode 3, is used to achieve self regulation and, whenever required, also to achieve simultaneously EC heat release. The particular EC heat-release technique used by the system employs a fan (fan 510).
In mode 0, pump 10 and fan 510 do not run; and MPMCU 518 ensures pump 404 is controlled so that pR tends to pRDo, where pRDo is a preselected desired current value for pR while the system is in mode 0.
In mode 1 (used only where the refrigerant is a non-azeotropic refrigerant), CCU 513 ensures: (1) pump 10 runs at a preselected effective capacity, usually near or equal to the pump's full effective capacity; (2) pump 404 is controlled so that pR tends to pRD, where pRD is a preselected desired current value for pR while the system is in modes 1 to 3; and (3) fan 510 does not run.
In mode 2, CCU 513 ensures: (1) pump 10 is controlled so that LP tends to LPD, where LPD is a preselected desired current value for LP high enough for all high heat-flux zones of the cylinder-head coolant passages to be covered with liquid refrigerant when the value of LP is close to LPD, and low enough for refrigerant vapor exiting separator 21 at 23 to be essentially dry; (2) pump 404 is controlled so that pR tends to pRD; and (3) fan 510 does not run.
In mode 3, CCU 513 ensures: (1) pump 10 is controlled so that pR tends to pRD; (2) pump 404 is controlled so that LR tends to LRD; and (3) fan 510 is controlled so that pR tends to pRD.
The preselected desired current value pRDo, pRD, or LPD, (of respectively pR, pR, or LP) may be a constant, or may be a value which changes in a pre-prescribed way as a function of one or more preselected characterizing parameters.
In the case of pRDo, a typical preselected characterizing parameter is the ambient atmospheric pressure pA, and a typical pre-prescribed way is the relation
pRDo=pA+Δop, (9)
where Δop is usually, but not necessarily, a fixed quantity. In the case of pR, typical preselected characterizing parameters and typical pre-prescribed ways are discussed in section V,H. And, in the case of LP, the desired current value LPD is usually a constant unless the condenser overfeed techniques described in section V,F,2,d are used, or unless the vehicle-tilt compensating techniques described in section V,F,2,f are used.
In the case of a non-azeotropic refrigerant, the transition rules between modes 0, 1, 2, and 3 are (where ‘eng.’ is an abbreviation for ‘engine’):
In rule (I), the value of ΔpR must be chosen large enough for the value of (pRD−ΔpR) to be smaller than the value of pR at which CCU 513 stops fan 510 running while the system is in mode 3.
In the case of an azeotropic-like refrigerant, mode 1 is eliminated and therefore transitions 0 to 1, 1 to 2, 2 to 1, and 1 to 0, are eliminated and the transition rule under (h) is changed to:
(h′) modes 2 to 0: eng. not running and TR<TR,MIN.
I note that, when the engine is started, the system may be in control mode 1, 2, or 3; but not in control mode 0 since, with the postulated transition rules, the system cannot be in control I mode 0 while the engine is running.
iv. Comments on Refrigerant Configuration and Control System
In this section V,F,2,a,iv I make miscellaneous comments on the refrigerant configuration and control system described in section V,F,2,a,i.
Where CR pump 10 is a high-slip positive displacement pump or a centrifugal pump, it is usually highly desirable, particularly in the case of two-step (on-off) control, to use unidirectional (one-way) valve 220, as shown in
Liquid refrigerant, exiting separator 21 at 24, can be returned to one or more points of refrigerant passages 504 or to one or more points of refrigerant passages 505, instead of to point outside the engine's refrigerant passages 504 and 505.
Proportional liquid-level transducer 113 can be used for three-step control, namely for controlling pump 404 so that it induces an essentially constant positive flow rate, an essentially constant negative flow rate, or no flow rate. If only three-step control of pump 404 is acceptable, a possibly less expensive three-step liquid-level transducer could be used provided the dead zones between steps are large enough to prevent unacceptably-fast cycling of pump 404. Similarly, a two-level (on-off) liquid-level transducer could be used to control two-step (on-off) operation of pump 10. (Three-step and two-step control of respectively pumps 404 and 10 has—among other disadvantages—the disadvantage of making it impracticable to control LR in mode 3, and LP in modes 2 and 3, as accurately as with proportional control.)
Although not essential, the control system may also include two-step liquid-level transducer 517 (see
Also, although not essential, the control system may further include two-step liquid-level transducer 519 (see
Finally, in several applications, MPMCU 518 is not required. In this case, mode 0 denotes that the system is in a safe state and that the system's CCU is de-energized. The value of pR while CCU 513 is de-energized may, for example, be chosen equal to the value of pRD at the instant TR falls below TR,MIN.
v. Other Refrigerant Configurations and Control Systems.
It should be clear, from the teachings so far in this DESCRIPTION, that the class VIIIFNooo principal configuration shown in
I would explain that principal configurations with a subcooler are, in some installations, desirable, or even necessary, to increase the amount of subcool of liquid refrigerant exiting, as applicable, receiver 7, and/or separator 42*, while the system is in control mode 3—to increase, for example, the net positive suction head available, as applicable, to pump 10 or to pump 46. The subcooler used may merely be a quasi-horizontal section of a refrigerant line which is located roughly in the same plane as refrigerant passages 399, and which is exposed to ram air and/or to the airflow induced by fan 510. An example of such a refrigerant line, in the case of a class VIIIFNsoo configuration, is finned refrigerant-line segment 9-522 shown in FIG. 43E.
I would also explain that in some installations having a principal configuration with a type 1 separator, a refrigerant pump may be desirable, or may be necessary, to return liquid refrigerant from the separator to the configuration's pool evaporator. Examples of installations where this is necessary are those where the desired location of separator 21 results in the level of the refrigerant liquid-vapor interface surface in it being below the level of the refrigerant liquid-vapor interface surface in refrigerant passages 505.
I would further explain that the control-mode rules of CR pump 10 and LT pump 404 can be reversed in control modes 2 and 3 if node 407, where the principal and the ancillary configuration join, were for example located (see
It should also be clear from the teachings so far in this DESCRIPTION that a type IIR, type IIIR, type IVR, type VR, or type VIR, ancillary configuration could have been used instead of the type IR ancillary configuration shown in FIG. 43. With types IIR to VIR configurations, the same control modes and transition rules as those described in section V,F,2,a,iii would apply, except that the controllable element (pump 404) of a type IR configuration would be replaced by the controllable element of one of the other five types of ancillary configurations; namely, for example, by motor 413 in the case of a type IIR configuration and, as applicable, by handwheel 479, by air-transfer pump 420, or by hydraulic pump 422, in the case of a type IIIR configuration.
Type IR to VIR two-port ancillary configurations are often desirable where the refrigerant employed is a two-component non-azeotropic fluid. A typical example of the locations of inlet 431 and outlet 432 are shown in
A damper or shutter with a controllable aperture upstream from an air-cooled condenser (with respect to the direction of airflow through the condenser) can be used to regulate the volumetric airflow of air through the condenser, and thereby control the rate at which the condenser releases heat to the air surrounding the condenser. I shall refer in this DESCRIPTION to this last-cited kind of heat-release control as ‘shutter-controlled heat-release’, or more briefly ‘SC heat-release’. SC heat release can be used with a system of the invention having an air-cooled condenser instead of, or in addition to, RC heat release. SC heat release is a particular form of externally-controlled passive heat release, or more briefly EC passive heat release.
I choose the refrigerant configuration shown in
Where SC heat release is used, in addition to RC heat release, mode 2 is replaced by modes 2A(s) and 4B(s). In mode 2A(s), the system's CCU (not shown) ensures (1) pump 10 is controlled so that LP tends to LPD; (2) pump 404 is controlled so that pR tends to pRD; (3) fan 510 does not run; and (4) motor 581 is controlled so that TR tends to a preselected value TRD higher than TR,MIN. And, in mode 2B(s), the system's CCU ensures (1) pump 10 is controlled so that LP tends to LPD; (2) pump 404 is controlled so that pR tends to pRD; (3) fan 510 does not run; and (4) shutter 580 is (completely) open.
The transition rules between modes 2A(s) and 2B(s) are
b. Cooling Systems with a Non-pool Evaporator
i. Preliminary Remarks
The evaporator in
I have therefore devised two-phase engine-cooling systems with no liquid-vapor interface surface in refrigerant passages 505; namely I have devised engine-cooling systems having a non-pool evaporator, or more briefly an NP evaporator. I next give examples of such cooling systems for the case of a V-engine, but similar systems can also be used with an in-line engine, an engine with opposed cylinders, or a radial engine.
ii. First Refrigerant Configuration, Control System, and Operating Method
The cooling system shown in
Liquid refrigerant, after flowing through node 522, enters at 530′a the NP component evaporator formed by the coolant passages of cylinder bank 500a, and very low-quality refrigerant vapor exits at 3″a; and liquid refrigerant enters at 530′b the NP component evaporator, formed by the coolant passages of cylinder bank 500b, and very low-quality refrigerant vapor exits at 3″b. Substantially dry refrigerant vapor exits separator 21 at 23 and liquid refrigerant in separator 21 exits at 24 and is returned to refrigerant passages 505a and 505b at points 523″a and 523″b, respectively, after flowing through node 524. Each of the alphanumeric symbols 530′a, 530′b, 3″a, 3″b, 523″a, and 523″b, designates a set of ports. The number of ports in each set need not be the same and can range from one to several ports. In the latter case, the number of ports in each set would typically be equal to the number of cylinders in a bank of cylinders, or to a multiple or submultiple of the number of cylinders in a bank of cylinders.
The location of vapor inlets 22a and 22b of separator 21 below liquid-vapor interface surface 521 helps ensure the refrigerant vapor quality is always low enough to assure potential hot spots in cylinder-head refrigerant passages 505a and 505b are always essentially wetted everywhere with liquid refrigerant without locating separator 21 at heights unacceptable—even in a fixed ground installation—to get the required evaporator overfeed. (See section V,F,2,b,iii.) The cross-sectional area of interface surface 521 is large enough to ensure the velocity of refrigerant vapor passing through that interface surface is small enough for refrigerant vapor exiting separator vapor outlet 23 to be substantially dry without using, in separator 21, separating surfaces that would cause an unacceptably high pressure drop, for example a pressure drop in excess of say 0.01 bar in the case of an aqueous glycol solution at a pressure of one bar.
Relations (1) to (8) in section V,F,2,a,ii can also be used to determine whether the cooling system shown in
The refrigerant configuration shown in FIG. 46—although preferred for certain installations—has, for many installations, at least two handicaps compared to alternative refrigerant configurations having a forced-circulation evaporator refrigerant auxiliary circuit. Firstly, refrigerant lines 3″a-22a and 3″b-22b must have a large-enough cross-sectional area to allow ‘sewer flow’, namely to allow liquid refrigerant and refrigerant vapor to flow in opposite directions; and secondly, separator 21 must be located above refrigerant outlets 3″a and 3″b.
iii. Second Refrigerant Configuration, Control System, and Operating Method
The engine-cooling system shown in
The system shown in
Pump 27 can be controlled by any technique which, explicitly or implicitly, maintains the value of qEV at or below a preselected value qEV,MAX low enough to prevent burn-out. This can, for example, be accomplished by controlling pump 27 so that the value of qEV tends toward a desired preselected value qEV,D which may be fixed, or which may change in a pre-prescribed way as a function of one or more preselected characterizing parameters.
Because, under steady-state conditions
where {dot over (m)}C is the refrigerant mass-flow rate induced by pump 10, where {dot over (m)}EO is the refrigerant mass-flow rate induced by pump 27, and where {dot over (m)}E is the refrigerant mass-flow rate exiting at 3″a and 3″b, it follows that the quality qEV of refrigerant vapor exiting the component evaporators, formed by the coolant passages of cylinder banks 500a and 500b, is—under steady-state conditions—a single-valued function of the evaporator-overfeed ratio
Consequently, the desired preselected value qEV,D can be obtained by controlling rEO or, almost equivalently, by controlling the ratio of the volumetric-flow rates FCR and FEO induced respectively by pumps 10 and 27. Techniques for controlling the ratio of FCR and FEO are disclosed in section V,B,3,e of my co-pending U.S. patent application Ser. No.400,738, filed 30 Aug. 1989. (Where pumps 10 and 27 are low-slip positive displacement pumps driven by stepping motors, or by pulse-width controlled motors, CCU 526 can use the signals generated by it, to control those motors, as a measure of the volumetric flow rates FCR and FEO induced respectively by pumps 10 and 27. Consequently no flow-rate transducers are necessary to obtain a measure of FCR and a measure of FEO.) The foregoing techniques for controlling the ratio FEO over FCR, and thus almost equivalently the value of rEO, are used whenever pump 10 is running. However, pump 10 may not always run while the engine shown in
Whenever the engine-cooling system shown in
iv. Other Refrigerant Configurations and Control Systems
It should be clear, from the teachings so far in this DESCRIPTION, that the class IIFNooo principal configuration shown in
All suitable principal configurations for piston-engine cooling systems with an NP evaporator must have sewer flow, or a substantial evaporator-overfeed ratio, or both. This is achieved in the case of subgroup IIFN and IIFF configurations in the way described in respectively sections V,F,2,b,ii and V,F,2,b,iii. I note that an alternative version of the class IIFFooo principal configuration shown in
A substantial evaporator-overfeed ratio can also be obtained by operating the DR pump of subgroup IIIFF and III*FF configurations like the EO pump of a subgroup IIFF configuration; namely by operating DR pump 46 so that the volumetric-flow rate FDR induced by it varies in a pre-prescribed way as a function of the volumetric-flow rate FCR induced by CR pump 10.
The EO and DR pump control techniques described so far in this section V,F,2,b may often be unsatisfactory because of unacceptably large differences between the current value of {dot over (m)}C and the current value of {dot over (m)}V during transients, where {dot over (m)}V is the mass-flow rate of essentially-dry refrigerant vapor in the principal configuration's refrigerant-vapor transfer means. In cases where such unacceptably large differences would occur, the EO and DR pump control techniques described so far can
The last-cited control technique—which can, with obvious changes, be used with either an EO or a DR pump—is described in this major paragraph using as an example a system, hereinafter referred to in this major paragraph as ‘the system’, consisting of the class III*FNoo principal configuration, and the type IVR ancillary configuration, shown in
The particular dual flow-rate control technique employed by the refrigerant configuration shown in
In
The system has, like all systems of the invention discussed so far, four control modes (in the case of a non-azeotropic refrigerant) which I refer to, in general, as modes 0, 1, 2, and 3. (I use dashes, as in section V,F,2,b,iii, only where I need to distinguish between different versions of those control modes.) Briefly, to recapitulate, modes 0, 1, 2, and 3, designate modes I shall refer to respectively as a minimum-pressure-maintenance mode; a mixing mode; an RC heat-release mode; and a combined self-regulation and EC heat-release mode. (The term ‘mixing mode’ refers to the action of mixing the components of a non-azeotropic refrigerant to achieve a more spatially-uniform concentration of its components.) The system has four controllable elements: DR pump 46, LT pump 404, condenser fan 510, and refrigerant bidirectional valve 528.
In mode 0, pump 46 and fan 510 do not run; valve 528 is open; and MPMCU 518 ensures pump 404 is controlled so that pR tends to pRDo.
In mode 1, CCU 527 ensures (1) pump 46 runs at a preselected capacity, usually near or equal to the pump's full capacity; (2) pump 404 is controlled so that pR tends to pRD; (3) fan 510 does not run; and (4) valve 528 is closed.
In mode 2, CCU 527 ensures (1) pump 46 is controlled so that qEV does not exceed qEV,MAX; (2) pump 404 is controlled so that pR tends to pRD; (3) fan 510 does not run; and (4) valve 528 is open.
In mode 3, CCU 527 ensures (1) pump 46 is controlled so that qEV does not exceed qEV,MAX; (2) pump 404 is controlled so that LR tends to LRD; (3) fan 510 is controlled so that pR tends to pRD; and (4) valve 528 is open.
The transition rules between the four modes recited in this major paragraph can be identical to those given in section V,F,2,a,iii.
I note that there is no identifiable liquid level in separating assembly 42*. Therefore, CCU 527 determines whether the refrigerant-circuit configuration shown in
I also note that the location of the inlet and outlet of the two-port ancillary configuration shown in
I further note that, where the signal C′DR used to control DR pump 46 provides a sufficiently accurate measure of FDR, transducer 142 can be eliminated.
It should be clear from the teachings so far in this DESCRIPTION that a type IIR, type IIIR, type IVR, or type VIR, ancillary configuration can be used instead of the type IR ancillary configuration shown in
Shutter-controlled heat release can be used with a cooling system of the invention having an NP evaporator in the same way as with a cooling system of the invention having a P evaporator.
c. Location of Evaporator Refrigerant Inlets and Outlets
Everywhere in this DESCRIPTION I distinguish between NP-evaporator liquid-refrigerant inlets and P-evaporator liquid-refrigerant inlets, and between NP-evaporator refrigerant-vapor outlets and P-evaporator refrigerant-vapor outlets. And I also everywhere in this DESCRIPTION distinguish, where applicable, between cylinder-block evaporator (liquid-refrigerant) inlets and (refrigerant-) vapor outlets on the one hand, and cylinder-head evaporator (liquid-refrigerant) inlets and (refrigerant-) vapor outlets on the other hand, by adding to numerals designating cylinder-block evaporator inlets and vapor outlets a single-dash superscript, and by adding to numerals designating cylinder-head evaporator inlets and vapor outlets a double-dash superscript. I further distinguish in this section V,F (and I have already done this in
An evaporator liquid-refrigerant inlet, or an evaporator refrigerant-vapor outlet, may consist of one or more ports. In the case where that inlet, or that outlet, consists of several ports, the several ports may be located at the same level or at different levels.
I stated in section V,F,2,b,iii that liquid refrigerant may have to be supplied to several locations in the coolant passages of a bank of cylinders. This is true not only with the class IIFFooo principal configuration discussed in the last-cited section, but also with any principal configuration. Preferred locations depend not only on the orientation of a piston engine's bank of cylinders but also on design details such as the precise configuration of cylinder-block and cylinder-head coolant passages. Liquid refrigerant can be delivered to these passages by nozzles to increase the velocity with which liquid refrigerant is injected into them, thereby generating turbulence and eliminating hot spots. I shall refer to the last-cited nozzles as ‘liquid-refrigerant injection nozzles’ or more briefly as ‘LR injection nozzles’. I use numeral 531 to designate a set of one or more LR injection nozzles.
A typical example of LR injection-nozzle locations is given in
In the typical example shown in
Turbulence promoters in the form of fins inside an engine's coolant passages, and/or in the form of grooves in the internal surfaces of those passages, are used by the invention, where desirable, to promote or to enhance turbulent refrigerant flow inside the engine's coolant passages.
The typical example shown in
d. Supplementary Control Techniques for Non-azeotropic Refrigerants
i. General Remarks
The refrigerants envisaged by me for piston-engine cooling and intercooling systems exposed to subfreezing water temperatures include azeotropic-like and non-azeotropic refrigerants. The former refrigerants include ethanol, methanol, acetone, HCFCs, and HFCs; and the latter include aqueous glycol, ethanol, methanol, and acetone, solutions.
Most of the non-azeotropic refrigerants I have in mind are—like the four last-cited solutions—two-component non-azeotropic refrigerants. I shall therefore, in this section V,F,2,d, consider only two-component non-azeotropic refrigerants. However, the techniques described in this same section also apply to non-azeotropic refrigerants with more than two components.
In the particular case of a two-component non-azeotropic refrigerant, the spatial distribution of the concentration of one of its components at a given point automatically determines the spatial distribution of the concentration of its other component at that point. I therefore need to consider the spatial distribution of the concentration of only one component.
Let c(x,y,z) be the concentration, at a point (x,y,z) of the liquid phase of the refrigerant's component with the higher evaporation (boiling) temperature (at a given pressure); let c be the concentration of the liquid phase of that component when its concentration is spatially uniform throughout a refrigerant-circuit configuration; and let {overscore (c)}E(x,y,z), or more briefly {overscore (c)}E, be the mean value of the concentration of the liquid phase of that component in the configurations evaporator. Then, while a principal configuration is active, the value of {overscore (c)}E will in general exceed the value of c, and consequently the mean value {overscore (T)}RS,E (of the refrigerant's saturated-vapor temperature TRS in a configuration's evaporator) will exceed the value of the refrigerant's saturated-vapor temperature TRS,O corresponding to the value of c. The difference ({overscore (T)}RS,E−TRS,O), if substantial, is undesirable, and I have therefore devised supplementary control techniques for reducing it. I distinguish between two-component non-azeotropic refrigerants, which I shall refer to as ‘group H refrigerants’, whose component with the lower freezing temperature has—as in aqueous glycol solutions—the higher evaporation temperature; and other two-component non-azeotropic refrigerants, which I shall refer to as ‘group L refrigerants’, whose component with the lower freezing temperature has—as in ethanol, methanol, and acetone, solutions—the lower evaporation temperature. I also note that the foregoing supplementary control techniques are essentially, but not necessarily exactly, the same for both group H and group L refrigerants.
ii. Cooling Systems with no Evaporator Refrigerant Auxiliary Circuit
In the just-cited case, the value of {overscore (c)}E−c) depends, for a given refrigerant and a given evaporator-overfeed ratio rEO, on the value of the ratio
and decreases as rM increases. In relation (13), ME is the mass of liquid refrigerant in the evaporator and ML is the mass of liquid refrigerant in the principal configuration outside the evaporator.
The value of ({overscore (c)}E−c), and the corresponding value of {overscore (T)}RS,E (at a given refrigerant pressure), may be acceptable, for certain two-component non-azeotropic refrigerants, for values of rM as low as unity—even where the evaporator-overfeed ratio is high. Examples of such two-component refrigerants are those which—like aqueous ethanol solutions—have component evaporation temperatures which do not differ greatly. (The boiling temperature at standard pressure of water and ethanol are respectively 100° C. and 77.7° C., and therefore differ by only 22.3° C.) By contrast, the value of {overscore (c)}E, and the corresponding value of {overscore (T)}RS,E, may not be acceptable for certain other two-component non-azeotropic refrigerants, even for values of rM as high as 3 or even higher—even where the evaporator-overfeed ratio is high. Examples of such two-component non-azeotropic fluids are ethylene glycol solutions and propylene glycol solutions. (The evaporation temperature, at standard pressure, of the former solution is 198° C. and of the latter solution is 187° C., and therefore these two temperatures differ from the boiling temperature of water by 98° C. and 87° C., respectively.) I consider as an example, in greater detail, a spatially uniform concentration of ethylene glycol equal to 0.5. Then, when rM is equal to unity, the value of ({overscore (c)}c−c) is, with a high value of rEO (say over 10), about 0.34, which at one atmosphere corresponds to a value of {overscore (c)}E of about 0.84 and to a value of {overscore (T)}RS,E of about 127° C. This temperature corresponds to an often undesirably-high rise in temperature above the boiling temperature of water at standard atmospheric pressure. With a design I have in mind, I expect the value of rM to be as high as 7 while some piston-engine cooling systems of the invention are in mode 3. This value corresponds, for c equal to 0.5, to a value of {overscore (c)}E equal to about 0.57, and to values of {overscore (T)}RS,E of about 109° C. and 105° C. at respectively one atmosphere and 0.8 atmosphere. This is usually acceptable. By contrast, when the system is in mode 2 and the system's condenser is almost completely filled with liquid refrigerant, the value of rM may approach unity and {overscore (T)}RS,E may approach 127° C. at one atmosphere, which is usually undesirable. I have therefore devised the techniques disclosed next to reduce, where necessary, the value of {overscore (c)}E and {overscore (T)}RS,E while the system is in mode 2. (These techniques can also be used for the same purpose in mode 3 at the expense of a larger condenser.)
All the techniques devised by me for reducing the concentration {overscore (c)}E and the temperature {overscore (T)}RS,E are based on the fact that, for a given value of rM, the value of {overscore (c)}E decreases as the value of the ratio qCV decreases, where
where qCV, {dot over (m)}V, and {dot over (m)}L, are respectively the quality of refrigerant vapor, the mass-flow rate of dry refrigerant vapor, and the mass-flow rate of liquid refrigerant, entering condenser 508 at refrigerant inlet 5; and where
is a ratio I shall refer to as the ‘condenser-overfeed ratio’.
The purpose of separator 21 is to ensure the value of {dot over (m)}L is essentially zero in mode 3. However, the purpose of operating the engine-cooling system in mode 2 is to decrease condenser effectiveness. This was achieved with the techniques described in sections V,F,2,b, and V,F,2,c, by backing-up liquid refrigerant in condenser refrigerant passages 399. Because condenser effectiveness decreases as rCO increases, the same result can be achieved by causing liquid refrigerant to enter passages 399 through condenser refrigerant inlet 5 instead of through condenser refrigerant outlet 6. This second way of decreasing condenser effectiveness decreases the value of {overscore (c)}E for a given value of rM, thereby also decreasing the value of {overscore (T)}E for a given value of pR. The value of {dot over (m)}L can be made to have a substantial value with several techniques.
The first set of techniques for achieving a required value of rM includes using a liquid-level independent-control technique to raise the level LP of interface surface 123 sufficiently for, as applicable, separator 21, separating assembly 21*, or separating assembly 42*, to become ineffective and cause wet refrigerant, instead of essentially dry refrigerant, to be supplied to air-cooled condenser 508. To this end, the value LPD3 of LPD in mode 3 would still be chosen low enough for separator 21, separating assembly 21*, or separating assembly 42*, to supply essentially dry refrigerant to condenser refrigerant passages 399, but the value LPD2 of LPD in mode 2 would be chosen high enough to cause separator 21, separating assembly 21*, or separating assembly 42*, to become sufficiently ineffective for the ratio rCO to tend toward a value high enough to prevent the value of {overscore (c)}E, or of {overscore (T)}RS,E, exceeding a preselected maximum value. A measure of {overscore (c)}E can be obtained by measuring the value of cE inside refrigerant passages 505 at a point below interface surface 123, and a measure of {overscore (T)}RS,E can be obtained by measuring the refrigerant temperature TR also at a point in refrigerant passages 505 below that interface surface. Then, for example, in the case of the subgroup VIIIFN, VIIIFF, IIFN and IIFF, configurations shown in respectively
TR−TRS,O≦εRS, (16)
where the value of TRS,O, as a function of the values of TR and c, can be computed for a given refrigerant and stored in a system's CCU; where εRS is a preselected positive quantity equal to a few degrees Celsius; and where LT pump 404 would usually be controlled so that pR tends to pRD.
The second set of techniques for achieving a required value of {dot over (m)}L includes by-passing, as applicable, separator 21, separating assembly 21*, or separating assembly 42*, with a liquid-refrigerant line connecting directly liquid refrigerant in refrigerant passages 504, or refrigerant passages 505, at a point below interface surface 123, to a point of refrigerant-vapor line 23-5, or to a point of condenser header 507; and to cause liquid refrigerant to flow in that liquid-refrigerant line when the engine-cooling system is in mode 2. This can be done in several ways. One of these ways is shown in
rCO≧rCO,MIN (17)
where rCO,MIN is a precomputed quantity, not necessarily fixed, stored in the cooling system's CCU (not shown). (For example, rCO,MIN may be a function of pR.) To this end, the cooling system's CCU determines the current value of {dot over (m)}V from a signal F′V generated by refrigerant vapor-flow transducer 136, and the cooling system's CCU generates a signal C′CO which controls pump 539 so that
{dot over (m)}L≧{dot over (m)}V·rCO,MIN (18)
iii. Cooling Systems with an Evaporator Refrigerant Auxiliary Circuit
In the just-cited case, the values of ({overscore (c)}EA−c) and ({overscore (T)}RS,EA−TRS,O) depend, for a given refrigerant and a given evaporator-overfeed ratio rE,O, on the value of the ratio
and decrease as rMA increases. In the expression ({overscore (c)}EA−c), the quantity {overscore (c)}EA is the mean value of the concentration cEA, in a principal configuration's evaporator refrigerant auxiliary circuit, of the liquid phase of the refrigerant's component with the higher evaporation temperature in the expression ({overscore (T)}RS,EA−TRS,O); the quantity {overscore (T)}RS,EA is the mean value of the refrigerant saturated-vapor temperature TRS in the principal configuration's evaporator refrigerant auxiliary circuit; and in relation (19), MEA is the mass of liquid refrigerant in the evaporator refrigerant auxiliary circuit, and MLA is the mass of liquid refrigerant in the principal configuration outside that auxiliary circuit.
The ratio rMA is—like the ratio rM—expected usually to be sufficiently high while the engine-cooling system is in mode 3, but not high enough while the system is in mode 2, and I have therefore devised several sets of techniques, similar to those devised for the case of cooling systems with P evaporators, to reduce, where necessary, the values of {overscore (c)}EA and {overscore (T)}RS,EA while engine-cooling systems with an NP evaporator are in mode 2. I next describe only essential differences between the two sets of techniques.
The essential difference between the first set of supplementary control techniques devised for engine-cooling systems with a P evaporator and the first set of supplementary control techniques devised for engine-cooling systems with an NP evaporator, is that in the former systems the effectiveness of, as applicable, separator 21, separating assembly 21*, or separating assembly 42*, is reduced indirectly by raising the level of liquid refrigerant in their P evaporator; whereas in the latter systems the effectiveness of separator 21 is reduced directly by raising the level of liquid refrigerant in their separator.
The essential difference, between the second set of supplementary control techniques devised for engine-cooling systems with a P evaporator and the second set of supplementary control techniques devised for engine-cooling systems with an NP evaporator, is that in the former systems liquid refrigerant is transferred to a point of refrigerant (vapor) line 23-5, or of condenser header 507, from the evaporator; whereas in the latter systems liquid refrigerant is transferred to that line, or to that header, from—as applicable—separator 21, liquid-refrigerant line 24-25, refrigerant line 21*-25, or refrigerant line 45*-49.
e. Location of Inlet and Outlet Ports of Two-port Ancillary Configurations
The supplementary control techniques disclosed in section V,F,2,d are, as mentioned in that section, essentially the same for group H and group L refrigerants. However, the control techniques, for helping ensure the concentration of the components of a two-component non-azeotropic refrigerant are spatially quasi-uniform throughout a cooling system's configuration before it cools down, depend in part on whether the refrigerant is a group H or a group L refrigerant. The reason for this is that
whereas that concentration will be low in the evaporator and high outside the evaporator where a group L refrigerant is employed; and similarly
It follows that a two-port ancillary configuration should preferably usually be connected to the principal configuration associated with it, so that
It also follows that a one-port ancillary configuration should preferably usually be connected, to the principal configuration associated with it, so that
f. Vehicle-tilt Compensating Techniques
i. Preliminary Remarks
Piston-engine cooling systems having an NP evaporator can be designed so that their performance is not affected adversely during large tilts, with respect to a local horizontal plane, of the vehicle on which they are installed. For example, such cooling systems can be made immune to tilts of up to at least 30 degrees, in any direction, where, as applicable, their separator has, or their receiver is, in the absence of tilt, a vertical cylindrical vessel with a length-to-diameter ratio of, say, no less than 2. By contrast, the performance of cooling systems having a P evaporator, a shallow separator, or a shallow receiver, may be affected adversely by tilts of 15 degrees or less. I use the term ‘shallow’ to denote, in the case of a vertical cylindrical vessel, a length to diameter ratio of less than one.
In the case of automobiles designed for road-only service, and in the case of ships, tilts exceeding, say, 15 degrees are, while the engine is running, unusual for long time intervals, but may occur for short time intervals. For such short time intervals (say less than one minute), I have devised the vehicle-tilt compensating techniques, described in the next two subsections of this section V,F,2,f, for two-phase engine-cooling systems with a P evaporator, an NP evaporator and a shallow separator, or with a shallow receiver.
ii. Cooling Systems with a Pool Evaporator
The vehicle-tilt compensating techniques devised for cooling systems with a P evaporator are based on the premise that whereas potential hot spots of the walls of cylinder-head passages 505 must remain immersed continuously in liquid refrigerant, a temporary degradation in cooling-system performance is acceptable if it causes the temperature of liquid refrigerant in passages 505 to rise temporarily by only a few degrees Celsius. The last-cited techniques are disclosed using the refrigerant configuration shown in FIG. 43.
I assume, for specificity only, that the one or more cylinder axes of the engine being cooled are vertical when the vehicle on which the engine is mounted is placed on a horizontal surface, and that therefore the angle θ of the cylinder axes, with respect to the normal to interface surface 123 (see FIG. 43), is equal to the vehicle tilt angle with respect to a local horizontal plane. And I use the letter φ to designate the azimuth angle of the vertical plane, containing the angle θ, with respect to a vertical plane fixed to the engine.
The value LP,MIN of LP at which potential hot spots of the walls of refrigerant passages 505 remain just immersed in liquid refrigerant is a function of 0 which, in general, is in turn a function of φ or, in symbols
LP, MIN=LP,MIN {θ(φ)} (20)
Relation (20) is stored in the engine-cooling system's CCU. This CCU uses relation (20) to compute a current value LPD high enough for LP to stay above LP,MIN, and then generates a signal L′P which controls CR pump 10 so that LP tends to LPD.
This action will ensure the potential hot spots cited earlier remain immersed in liquid refrigerant at the expense of a degradation in cooling-system performance whenever the level of interface surface 123 rises sufficiently for separator 21 to be unable to deliver essentially dry refrigerant vapor to condenser 508.
Suitable tilt transducers include two inclinometers at right angles to each other in a plane, fixed to the engine, which is horizontal when the engine's cylinder axes are vertical. Typical examples of inclinometers are LVDT-type transducers. Inclinometers 548 and 549 (see
In cases where tilt in only one vertical plane is of interest only one inclinometer is used. The signal generated by it could, in some applications, only be a two-step, or at most a three-step signal.
iii. Cooling Systems with a Non-pool Evaporator and Shallow (Type 1) Separator Having Vapor Inlets Below Liquid Level
The vehicle-tilt compensating techniques devised for cooling systems with an NP evaporator and with a shallow separator, and in particular with a shallow type 1 separator, having a set of one or more vapor inlets below interface surface 521, are similar to those devised for cooling systems having a P evaporator. Namely, the techniques devised to ensure the potential hot spots of the walls of a P evaporator remain immersed in liquid refrigerant during vehicle tilts are used to ensure the last-cited separator's set of vapor ports remains covered by liquid refrigerant during those tilts. The only essential difference is that signals θ′1 and θ′2 provide measures of the inclination, with respect to a local horizontal plane, of separator 21, and not of the inclination of a bank of cylinders which may, as in the case of a V engine, have a different inclination from another bank of cylinders of the same engine.
g. Cabin-heating
Cabin heating, when desired, can be performed by using one or more refrigerant circuits which are an integral part of the refrigerant configuration used to cool an internal-combustion piston engine. This can be done in several ways which can be divided into two sets: ways which use single-phase heat-transfer and ways which use two-phase heat-transfer.
In the former case, the class VIIIFNooo configuration shown in
In the cabin-heating refrigerant circuits described next, I shall use alphanumeric symbols to denote components and points. The numeral in these symbols, where already used in this DESCRIPTION, designates the same kind of component as, or the corresponding point to, respectively the component, or the point, already designated by that numeral in this DESCRIPTION; and the letter ‘h’ in those alphanumeric symbols signifies that those symbols designate a component or a point belonging either exclusively or primarily to a cabin-heating circuit.
An example of a cabin-heating circuit employing an NI subcooler refrigerant auxiliary circuit is shown in
An example of a cabin-heating circuit employing an NI subcooler refrigerant auxiliary circuit is shown in
I note that the subcooler refrigerant auxiliary circuit shown in
I also note that SC pump 63h can also be used to perform the function of a CO pump. To this end, outlet 73h of subcooler 551h is connected, whenever required, to a point of the refrigerant principal circuit between, as applicable, separator vapor outlet 23, separator vapor outlet 44, separating-assembly vapor outlet 23*, or separating-assembly vapor outlet 44*, on the one hand; and condenser refrigerant passages 399 on the other hand.
An example of one or more cabin-heating circuits which are a branch of a split principal configuration, with two parallel branches sharing the selfsame P evaporator, is shown in FIG. 43M. The cabin-heating branch of the split principal configuration shown in
An example of one or more cabin-heating circuits, which are a branch of a split principal configuration with two parallel branches sharing the selfsame NP evaporator and the selfsame separator, is shown in FIG. 46G. The cabin-heating branch of the split principal configuration shown in
I note that a cabin-heating branch, using two-phase heat transfer, could use other refrigerant circuits; and, in particular, (1) refrigerant circuits with a constant-capacity DR pump, or (2) a natural refrigerant-circulation circuit, with a refrigerant valve, where interface surface 116h is above interface surface 521.
3. Intercooling Systems with an Air-cooled Condenser
a. General Remarks
Certain internal-combustion piston engines use a supercharger, which may be either a mechanically-driven supercharger or a turbocharger. The efficiency and shaft power of such engines can be, and has been, increased by intercooling; namely by cooling the compressed air discharged by a supercharger before it is supplied to the engine's one or more combustion chambers.
Intercoolers of the present invention may, like prior-art intercoolers, be independent of (namely separate and distinct from) a piston-engine's cooling system, or be an integral part of a piston-engine's cooling system. Independent intercoolers are generally preferred because they can be used to lower the air delivered, by a piston-engine's supercharger to the engine's cylinders, below the temperature of the engine's coolant. However, intercoolers which are an integral part of a piston-engine's cooling system, in the sense that they share the system's condenser, are within the scope of the invention disclosed in this DESCRIPTION. Such intercoolers would be a branch of a split principal configuration with two branches sharing the radiator (condenser) of the engine being intercooled.
At this time (1991), an aqueous ethylene glycol solution is the generally preferred refrigerant for single-phase piston-engine cooling systems exposed to temperatures below zero degrees Celsius; and is one of the preferred refrigerants for two-phase piston-engine cooling systems exposed to such temperatures. By contrast, the generally preferred refrigerant for cooling compressed air discharged by the supercharger of a piston engine is often not an aqueous ethylene (or an aqueous propylene) glycol solution. The reason for this is that it is often desirable to cool the air discharged by such a supercharger down to at least 60° C. with a refrigerant that boils, at acceptable absolute pressures, down to at least 55° C. Minimum acceptable refrigerant absolute pressures for intercooling are considerably lower than those for piston-engine cooling primarily because of the absence of cylinder-head gaskets. Nevertheless, I expect the cost of an intercooler to start rising rapidly as the minimum pressure, to which the system's principal configuration is subjected, falls below about 0.5 bar. The temperature at which an aqueous 50% ethylene or propylene glycol solution starts to boil exceeds 70° C. at even 0.3 bar. Consequently, refrigerants with lower boiling points than those of aqueous ethylene and propylene glycol solutions may be preferable for independent intercoolers.
Suitable refrigerants in freezing climates for independent intercoolers with a minimum-pressure-maintenance capability include ethanol, methanol, acetone, and their aqueous solutions.
The purpose of an intercooler is to maintain the temperature TIi of air exiting the intercooler and supplied to the engine's cylinders, at a preselected desired temperature TIDi; where TIDi may have a fixed value or may have a value which varies in a pre-prescribed way as a function of one or more preselected parameters which include ambient air temperature, ambient air pressure, supercharger-output air temperature, supercharger-output air pressure, and parameters characterizing the state of the engine.
Where no minimum-pressure-maintenance and no refrigerant-controlled heat-release capabilities are required, several of the refrigerant-circuit configurations and control techniques disclosed in my co-pending U.S. patent application Ser. No. 400,738, filed 30 Aug. 1989, can be used for piston-engine intercoolers.
I would mention that, particularly where a non-azeotropic refrigerant is used, it is sometimes desirable to confirm that the intercooler's principal configuration is filled completely with liquid refrigerant. Several methods can be used to do this. A first of those several methods is to use a two-step liquid-level transducer at the highest point of the principal configuration, and to determine whether liquid refrigerant has reached that transducer, but this first method is impracticable for intercoolers subjected to substantial tilts. A second of those several methods, which is also applicable to intercoolers subjected to substantial tilts, is to use an absolute-pressure transducer to obtain a measure of refrigerant pressure, and a refrigerant-temperature transducer in the same neighborhood to obtain a measure of refrigerant (sensible) temperature; to compute the refrigerant saturated-vapor temperature corresponding to the measured refrigerant pressure; to compare the measured refrigerant sensible temperature and the computed refrigerant saturated-vapor temperature; and to use the fact that the latter temperature exceeds the former temperature by a preselected amount as confirmation that the principal configuration is completely filled with liquid refrigerant.
Preferred intercoolers of the invention usually have NP evaporators. I note that a substantial evaporator-overfeed ratio is not needed to prevent hot spots in the evaporator refrigerant passages of an intercooler; and is usually also not needed to prevent, in the case of a non-azeotropic refrigerant, a refrigerant saturated-vapor temperature rise in those passages. The reason for this is that such a temperature rise usually has no adverse effect comparable to that which would be caused by it if it occurred in a piston-engine's coolant passages. Consequently, preferred principal configurations for piston-engine intercoolers of the invention need not include means for overfeeding their evaporator. Therefore, in principle, any group I to III, VII to IX, II*, III*, VIII*, or IX*, configuration, usually with a refrigerant principal pump, and no preheater, superheater, or desuperheater, might be a preferred principal configuration.
b. A Fast-response Intercooler
I shall describe typical ways of operating an independent fast-response intercooler using
The numeral in the alphanumeric symbols used in
I note that the rise in the intake-air temperature entering an intercooler's evaporator can be very rapid just after the supercharger starts operating, and therefore that the intercooler—if it were completely filled with liquid refrigerant while the engine is running and the supercharger is not running—would often not be able to reach mode 3 fast enough to maintain TIi at its preselected desired value TIDi unless, as applicable, pump 404, motor 413, air-transfer pump 420, or hydraulic pump 422, is unacceptably large. Consequently, the invention includes, where desirable, means for preventing the refrigerant pressure of an intercooler using a type A (or incidentally also a type B) combination from falling below a preselected minimum value—while the engine is running and the supercharger is not running—without requiring the combination's principal configuration to be filled completely with liquid refrigerant. To this end, I use heat available in the engine's exhaust. (I could alternatively use an electric heating element. This, however, would consume a substantial amount of utilizable power whereas using exhaust-gas heat does not.)
Many piston engines have means for heating their intake air with their exhaust gases during cold weather. Instead of using the exhaust gases of a piston engine to heat its intake air directly, I use those gases to heat its intake air indirectly through the engine's intercooler by heating a refrigerant-circuit segment of its principal configuration. I can thus achieve minimum-pressure maintenance with a principal configuration only filled partially with liquid refrigerant while, at the same time, transferring heat to the engine's intake air through the intercooler's principal configuration.
In
In
I now describe a first typical control technique for reducing the response time of the heat-transfer rate of an intercooler of the invention, immediately after the supercharger (with which the intercooler is associated) starts running, by utilizing heat from the engine's exhaust gases. To this end, while the engine is running and the supercharger is not running, I use heat from the engine's exhaust gases to allow minimum-pressure maintenance to be achieved with no liquid-refrigerant in refrigerant-vapor transfer-means segment 23i-5i, and with the value of LCi equal to LCDi. This ensures the intercooler (1) can start releasing heat, without significant delay, when the supercharger starts running (provided the engine has been running for a few seconds, or at most for a few tens of seconds, before the supercharger starts running); and (2) can change to mode 3 much faster than it could if the refrigerant circuits of the intercooler's principal configuration were filled completely with liquid refrigerant.
I shall, in this section V,F,3,b, refer (1) to the intercooler shown in
The intercooler has five control modes: control modes 0E, 0S, 1, 2, and 3. Control modes 1, 2, and 3, designate—as in the case of engine-cooling systems—respectively a mixing mode (used only with non-azeotropic refrigerants); an RC heat-release mode; and a combined self-regulation and EC heat-release mode, where the EC heat-release mode is a fan-controlled heat-release mode. Control mode 0E designates a minimum-pressure-maintenance mode during which the engine (with the intercooler) is not running, and corresponds to control mode 0 in the case of an engine-cooling system; and control mode 0S designates a combined minimum-pressure-maintenance mode during which the engine is running and the engine's supercharger is not running, and is a combined minimum-pressure-maintenance and fast-response-separation mode. CCU 563i (shown in
The first typical control technique does not use intercooler shutter 580i; and thus has only four system-controllable elements: CR pump 10i, LT pump 404i, fan 510i, and damper 567i.
In mode 0E, pump 10i and fan 510i do not run; damper 567i is in a preselected position (say closed, open, or half open); and MPMCU 564i controls pump 404i so that pRi tends to pRDoi where pRDoi is a preselected value of pRi. (CCU 563i places damper 567i in that preselected position at the instant in time when it is de-energized.)
In mode 0S, CCU 563i ensures (1) pump 10i is controlled so that the level LXi tends to LXDi; (2) pump 404i is controlled so that a refrigerant liquid-vapor interface surface forms in header 507i and thereafter has a level which tends to LCDi (while the intercooler is in mode 0S); (3) fan 510i does not run; and (4) damper 567i is controlled so that TIi tends to TIDi.
In mode 1 (used only with a non-azeotropic refrigerant), CCU 563i ensures (1) pump runs at a preselected capacity, usually near or equal to the pump's full capacity; (2) pump 404i is controlled so that pRi tends to pRDoi, (3) fan 510i does not run; and (4) damper 567i is closed.
In mode 2, CCU 563i ensures (1) pump 10i is controlled so that LSi tends to LSDi; (2) pump 404i is controlled so that TIi tends to TIDi; (3) fan 510i does not run; and (4) damper 567i is closed.
In mode 3, CCU 563i ensures (1) pump 10i is controlled so that LSi tends to LSDi; (2) pump 404i is controlled so that LRi tends to LRDi; (3) fan 510i is controlled so that pRi tends to pRDoi; and (4) damper 567i is closed.
The transition rules between the last-cited five modes are:
In transition rules (a), (c), and (f), small delays between the event specified and the corresponding transition may be desirable and can be preselected. For example, a small delay may be desirable in transition rule (a) between the time the engine starts and the cited transition occurs to allow the exhaust gases, after a cold start, to be hot enough to ensure the refrigerant pressure does not momentarily fall below its minimum-permissible value.
I note that while the intercooler is in mode 2, the refrigerant pressure might, in very cold climates and under certain operating conditions, fall below its minimum-permissible value. If such an occurrence is possible, an additional control mode can be added during which damper 567i is partially opened to allow exhaust gases to supplement heat supplied by the intake air entering evaporator 561i, and thereby ensure the refrigerant's pressure does not fall below its minimum-permissible value.
I now describe a second typical control technique for reducing the response time of the intercooler. The second typical control technique allows the intercooler to achieve, after a given small time interval (say a few seconds) after the supercharger starts running, a much larger heat-transfer rate than that achievable with the first typical control technique after the same time interval. To this end, I use the engine's exhaust gases to allow the refrigerant liquid-vapor interface surface, upstream from CR pump 10i, to be located in receiver 7, instead of in condenser header 507i, while the intercooler is in mode 05. Whenever required, or whenever desirable, the invention (see for example
To implement the intercooler second typical control technique, I need only
The CCU for implementing the second typical control technique differs in essence from CCU 563 only in that it also generates a signal C′CSi for controlling shutter motor 581i; and the MPMCU for implementing that technique is the same as MPMCU 518.
4. Cooling Systems with a Water-cooled Condenser
a. General Remarks
A first principal difference, in piston-engine cooling applications, between type A combinations having a water-cooled condenser and type A combinations having an air-cooled condenser is that
A second principal difference, in piston-engine cooling applications, between type A combinations with a water-cooled condenser and type A combinations with an air-cooled condenser is that the former combinations are often installed in a building or on a ship and that consequently their refrigerant is usually never exposed to water-freezing temperatures, whereas the refrigerant of the latter combinations is in most applications exposed, at some time, to such temperatures. It follows that, for piston-engine cooling installations, the preferred refrigerant for type A combinations with a water-cooled condenser is often water (with, where required, passivation and anti-corrosion additives). Exceptions include installations in motor boats with no permanent heated engine room.
Because of the facts mentioned in the two immediately-preceding minor paragraphs. I shall limit my discussion of type A combinations with a water-cooled condenser to combinations having no freeze-protection capability (even where their refrigerant is water) and no RC heat-release capability.
b. Refrigerant Configuration and Control System
Liquid refrigerant entering at 2″ is supplied to refrigerant passages 505 inside one or more spaces bounded by one or more weirs 599. In the particular case where inlet 2″, consists of a number of ports equal to the number of cylinders of engine 500 in
I said in the first minor paragraph of this section V,F,4,b that the refrigerant configuration shown in
The refrigerant configuration shown in
c. Unsafe and Safe States
I shall say that the system to which the refrigerant configuration shown in
d. Typical Operating Method
I now outline a typical method of operating a system having the refrigerant configuration shown it)
The system-controllable elements of the system are DR pump 46, LT pump 404, and ‘cold-water pump’ 598. (Pump 598 is a particular kind of cold-fluid pump.) The system has three control modes: modes 0, 20, and 3, where—as in sections V,F,2 and V,F,3—mode 0 is a minimum-pressure-maintenance mode while the engine is not running; where mode 20 is in essence a minimum-pressure-maintenance mode while the engine is running; and where mode 3 is a combined self-regulation and EC heat-release mode. The particular EC heat-release technique employed uses CW pump 598. The system includes CCU 590 shown in FIG. 55 and MPMCU 518 shown in FIG. 45.
In mode 0, pump 46 and pump 598 do not run and MPMCU 518 ensures pump 404 is control led so that pR tends to pRDo.
In mode 20, CCU 590 ensures (1) pump 46 is controlled in a pre-prescribed way as a function of the engine's fuel mass-flow rate {dot over (m)}F, or almost equivalently as a function of the engine's fuel volumetric-flow rate FF; (2) pump 404 is controlled so that pR tends to pRDo; and (3) pump 598 does not run. (The sensor providing a measure of the fuel-flow rate is not shown.)
In mode 3, CCU 590 ensures (1) pump 46 is controlled in a pre-prescribed way as a function of the current value of the engine fuel-flow rate FF; (2) pump 404 is controlled so that the level LD of liquid-vapor interface surface 139, as indicated by signal L′D generated by liquid-level transducer 145, tends to a preselected, usually fixed, value LDD; and (3) pump 598 is controlled so that pR tends to pRD.
The transition rules between modes 0, 20, and 3, are:
In the foregoing transition ΔpR1 and ΔpR2 are small positive quantities, and ΔpR1 is larger than ΔpR2.
Refrigerant-pump control as a function of fuel-flow rate is discussed in section V,H.
e. Other Refrigerant Configurations and Control Systems
Any of the other refrigerant configurations and control systems described or mentioned in section V,F,2 can also be used with piston engines cooled by a system of the invention using a type A combination and a water-cooled condenser. The preferred refrigerant configuration and control system depends on the details of the particular application of interest.
5. Elimination of Minimum-pressure-maintenance Control Unit
a. Preliminary Remarks
In discussing minimum-pressure-maintenance with type A combinations having no MPMCU, I distinguish between combinations having (1) a type IR, or a type IIIR, configuration; and (2) a type IIR, a type IVR, or a type VR, configuration. Combinations having no MPMCU and a type IR, or a type IIIR configuration often can, while their principal configuration is inactive, maintain the current value (pR−pA) at a preselected value accurately over a wide range of environmental temperatures. By contrast, type A combinations having no MPMCU and a type IIR, a type IVH, or a type VR, configuration usually cannot do so.
In a system of the invention with a type A combination and no MPMCU, control mode 0 is eliminated and is replaced by a control mode 00 in which by definition none of the controllable elements of the type A combination, and in particular of its ancillary configuration, are controlled by the system.
I next give two examples of operating methods where an MPMCU is not employed, and where control mode 0 is replaced by control mode 00. The first example is a type A combination having a type IR configuration with a spring. The second example is a type A combination having a type IIIR configuration which can maintain the value of (pR−pA) between a preselected upper limit and a preselected lower limit while the combination is inactive. (If the AT pump used allows air to flow through it at a sufficient rate while it is not running, a type IIIR configuration could be used to make the value of (pR−pA) equal to zero while the combination is inactive.)
b. Example with a Type IR Ancillary Configuration
The principal configuration employed in the first example, see
I assume, for specificity only, that, in mode 00, the minimum-permissible refrigerant pressure is the current ambient atmospheric pressure. I choose a spring (spring 478) which exerts a contracting force large enough to offset the expanding force exerted by corrugated cylindrical wall 403, and thus ensure the refrigerant pressure does not fall below ambient atmospheric pressure while the system's principal configuration is inactive. Clearly spring 478 can alternatively be chosen to exert a force which results in a preselected non-zero (positive or negative) current value of (pR−pA) while the principal configuration is inactive.
The system having the refrigerant configuration shown in
Mode 00A is a minimum-pressure-maintenance mode while engine 500 is not running, and corresponds to mode 00. And mode 00B is a minimum-pressure-maintenance mode while engine 500 is running but cold, and the effective capacity of pump 46 is zero although the engine is running. The purpose of mode 00B is to accelerate engine warm-up while TR is lower than TR,MIN.
Mode 1A is used to achieve the same purpose as mode 1, namely to mix the components of a non-azeotropic refrigerant so that the concentrations of their liquid phases are approximately spatially uniform. And mode 1B, which I name ‘the dry-up-prevention mode’, is used to continue cooling the engine, after it stops running; while TR is at or above TR,MIN.
Modes 2 and 3 have the same purposes as those recited in section V,G,2,a,iii.
Three-step liquid-level transducer 592 generates a signal L′R indicating whether LR has risen above an upper limit LR,MAX or fallen below a lower limit LR,MIN. (A proportional liquid-level transducer, or two two-level liquid-level transducers can be used instead of transducer 592.) Refrigerant-selector valve 585h and cabin-heating subcooler fan 552h are controlled manually by an operator or automatically by the cabin climate-control system. Cabin-heating SC pump 63h is controlled by the system's CCU only during modes 1A and 1B; and then only in the sense that the system's CCU causes pump 63h to run while the system is in any one of those two modes it if is not running (because the cabin-heating system has been turned off). Refrigerant-selector valve 586 has an inlet 587, an inlet 588, and an outlet 589; and is in position 1 in modes 1A and 1B, and in position 2 in all other modes, where position 1 causes liquid refrigerant to enter valve 586 through inlet 587 and where position 2 causes liquid refrigerant to enter valve 586 through inlet 588. Refrigerant-blocking valve 528 is closed only in mode 1A and bidirectional two-step (on-off) recirculation-control valve 591 is open only in mode 00B. (I note that valve 528, instead of being controlled by the system's CCU, could be a thermostatically-control led valve which closes when TR<TR,MIN, and which opens when TR>(TR,MIN+ΔTR), where ΔTR is a small positive quantity.) The system's CCU controls pump 63h, valve 586, valve 528, and valve 591, with signals C′SCH, C′RSV1, C′RBV, and C′RCV, respectively. The remaining system-controllable elements of the refrigerant configuration shown in
In mode 00A, no system-controllable elements are controlled. In mode 00B, (1) pump 404B and valve 435 are controlled only in certain applications where this is desirable, so that PR tends to pRDo, and (2) fan 510 does not run. In mode 1A, (1) pump 404B and valve 435 are controlled so that pR tends to pRDo, and (2) fan 510 does not run. In mode 1B, (1) pump 404B and valve 435 are controlled so that pR tends to pRDo, and (2) fan 510 runs at a preselected effective capacity, namely usually at a preselected speed. In mode 2, (1) pump 404B and valve 435 are controlled so that pR tends to pRD, and (2) fan 510 does not run. And in mode 3, pump 404B and valve 435 are controlled so that LR stays close to LRD, and (2) fan 510 is controlled so that pR tends to pRD.
Where an engine is a multicylinder engine installed on a platform which subjects it to substantial tilts in its longitudinal direction, the engine should usually have separate and distinct cylinder-head coolant passages for each cylinder. For example, an in-line engine with 4 cylinders should usually have four sets of separate and distinct cylinder-head coolant passages, four liquid-refrigerant inlet ports, four liquid-refrigerant outlet (overflow) ports, and four refrigerant-vapor outlet ports.
c. Example with a Type IIIR Ancillary Configuration
I use as an example the refrigerant configuration shown in
The type IIIR configuration used has a high-slip unidirectional air-transfer pump 420A and leakproof two-step bidirectional air valve 483 in series with it. Pump 420A, while not running, allows air to leak through it in the reverse direction at a high-enough rate for it (1) not to have to be bidirectional or (2) not to need a bidirectional valve in parallel with it to allow air to exit space 421 at a fast-enough rate to control pR in mode 2 and to control LR in mode 3. Valve 483 is leakproof in the sense that it does not allow air from space 421 to leak through it, while it is closed and pump 420A is not running, for pressures across it up to, say, one bar. The CCU of the refrigerant configuration shown in
An alternative refrigerant configuration to that shown in
d. Other Ancillary Configurations
The inert gas in the LR reservoir of a type A combination having a type IVR, or a type VR, ancillary configuration has essentially a constant volume in mode 00. Consequently the pressure of that inert gas will, in that mode, change its value as a function of ambient temperature TA; and therefore so will the current value of pR. Also, the value of the pressure of the inert gas in the LR reservoir is essentially unaffected by changes in ambient atmospheric pressure pA. It follows that, in applications where substantial changes in TA and/or in pA occur, the resulting changes in the current value of (pR−pA) may be unacceptable. In such applications an MPMCU would have to be used with type A combinations having a type IVR, or a type VR, ancillary configuration.
I note that the invention includes modified type IR, IIR, and IIIR, ancillary configurations which—although they have a variable-volume reservoir—contain an inert gas (like ancillary configurations with a fixed volume).
1. Preliminary Remarks
I discuss in this section V,G applications where the properties complete minimum-pressure maintenance and self regulation are required, and where gas-controlled heat release, or more briefly GC heat release, is usually also required.
In sections V,G,2 and V,G,3 I describe type C combinations, and their associated control techniques, for the case where the combinations' condenser is an air-cooled condenser. And, in section V,G,4 I describe type C combinations, and their associated control techniques, for the case where the combinations' condenser is a water-cooled condenser.
Because all the type A combinations discussed in this section V,G have no partial minimum-pressure maintenance, I shall refer for brevity, in this section V,G, to complete minimum-pressure maintenance simply as ‘minimum-pressure maintenance’. This property, as mentioned in section III,D, is achieved in type C combinations by inserting an inert gas in their principal configuration.
2. Cooling Systems with an Air-cooled Condenser
a. Cooling Systems with a Pool Evaporator
i. First Refrigerant & Inert-gas Configuration, Control System, and Operating Method
The R&IG configuration shown in
The control system includes CCU 600 and MPMCU 601 shown respectively in
Proportional absolute-pressure transducer 603 performs a different function from that performed by absolute-pressure transducer 514 in type A combinations; and it is for this reason that I have designated the former transducer by a different numeral from that used to designate the latter transducer. More specifically, transducer 603 generates a signal p*′R which provides, at a preselected location in the principal configuration of an R&IG configuration, (1) a measure of the total pressure p*R in the principal configuration, which is in general the current value of the sum of the partial refrigerant pressure and the partial inert-gas pressure; and which is in particular (2) a measure of the current value of the refrigerant pressure pR in the absence of inert gas or a measure of the current value of the inert-gas pressure pG in the absence of refrigerant.
Proportional absolute-pressure transducer 605 generates a signal p*′GR providing a measure of the current value of the total pressure p*GR in reservoir 453, and proportional gas-temperature transducer 606 generates a signal T′GR providing a measure of the current value of the inert-gas temperature TGR in reservoir 453.
The terms ‘unsafe state’ and ‘safe state’, in the case of engine-cooling systems using type C combinations, have the same meanings as those given in section V,F,2,ii. However, the set of conditions indicating whether an engine-cooling system using a type C combination is in an unsafe state, or in a safe state, are different, Namely, I shall say that the last-cited system is in an unsafe state, while the engine is running and hot, when one of the following four relations is true:
LP<LP,SAFE; LR<LR,SAFE; p*R>p*R,SAFE; and TR>TR,SAFE; (1), (2), (3*), (4)
and I shall say that the last-cited system is in a safe state, while the engine is running and hot, if all of the following four relations are true:
LP≧LP,SAFE; LR≧LR,SAFE; p*R≦p*R,SAFE; and TR≦TR,SAFE. (5), (6), (7*), (8)
An engine is, by definition, hot when the current value of TR exceeds TR,MIN, as defined earlier.
I now outline a typical method of operating the system shown in
I start at an instant in time when the engine being cooled by the system is not running and is started, say, by an operator manually. When the engine is started, CCU 600 and all its associated transducers and controllable elements are energized, if they are not already energized.
CCU 600, as soon as it is energized, and subsequently at frequent preselected periodic time intervals while it remains energized, performs a system safety check to determine whether the system is in a safe state. If it is not, an audible and/or visual warning signal is generated to indicate that the system is in an unsafe state, and the engine, after being stopped by the operator, is inhibited from being started. If the unsafe state has occurred because pR or TR, or both, have exceeded their safe values, CCU 600 runs fan 510 at its maximum capacity until their safe values are no longer exceeded, and then de-energizes itself automatically. MPMCU 601, which is always energized while the system is in a safe state, remains energized and controls LT pump 404 in the same way as in control mode 0*. (See next major paragraph.) If the system has become unsafe because of an insufficient refrigerant charge, MPMCU 601 will de-energize itself automatically. (The refrigerant charge is insufficient when relation (1) or (2) is satisfied.)
I shall describe the operation of systems of the invention with a type C combination, while they are in their safe state, in terms of control modes and transition rules. In
In mode 0*, pump 27 and fan 510 do not run, and diverter valve 555 is in position 1, namely by definition valve 555 is in a position which causes liquid refrigerant entering at inlet 556 to exit at outlet 557; and MPMCU 601 controls pump 10 so that LP tends to LPD, and controls pump 443 so that p*R tends to p*RDo, where p*RDo is the preselected desired current value for p*R while the system is in mode 0*.
In mode 1*, CCU 600 ensures (1) pump 10 is controlled so that LP tends to LPD; (2) pump 27 is controlled so that LS tends to LSD; (3) pump 443 is controlled so that p*R tends to p*RD, where p*RD is a preselected desired current value for p*R while the system is in modes 1* to 3*; (4) fan 510 does not run; and (5) valve 555 is in position 2, namely by definition valve 555 is in a position which causes liquid refrigerant entering at inlet 556 to exit at outlet 558; and (6) pump 63h runs at or near its maximum capacity.
In mode 2*, CCU 600 ensures (1) pump 10 is controlled so that LP tends to LPD; (2) pump 27 is controlled so that LS tends to LSD; (3) pump 443 is controlled so that p*R tends to p*RD; (4) fan 510 does not run; and (5) valve 555 is in position 1.
In mode 3*, CCU 600 ensures (1) pump 10 is controlled so that LP tends to LPD; (2) pump 27 is controlled so that LP tends to LPD; (3) pump 443 is control led so that p*GR tends to p*GR,3, where p*GR.3 is a preselected value of p*GR discussed in the immediately-following major paragraph; (4) fan 510 is controlled so that p*R tends to p*RD; and (5) valve 555 is in position 1.
The transition rules between the last four modes are (where ‘eng.’ is an abbreviation for ‘engine’):
In rule (f), Δp*R1 is a finite positive quantity; and, in rule (I), Δp*R2 is a finite positive quantity large enough for the value (p*R−Δp*R2) not to be larger than the value of p*R at which CCU 600 stops fan 510 running while the system is in mode 3*. The clock mentioned in rule (g) is used in the way described in the third minor paragraph of the second major paragraph of section V,F,2,a,iii.
In general, the preselected value p*GR,3 may be a fixed value, or a value which changes in a pre-prescribed way as a function of p*GR and TGR.
In the former case, proportional absolute-pressure transducer 605 can be replaced by a two-step pressure transducer indicating whether p*GR is close to p*GR,MAX, transducer 606 can be eliminated, and the value of p*GR,MAX is typically chosen equal to the design maximum operating value p*GR,MAX of p*GR.
In the latter case signal T′GR, generated by transducer 606, is used to compute the value p*GR,3 of p*GR at which the principal configuration contains essentially no inert gas at a preselected typical value of TR when the system is in mode 3*. Assuming reservoir 453 contains essentially only inert gas, the value of p*GR,3 can be computed, as a function of TGR, by using van der Waal's equation. Where the values of p*GR,3 are low enough, the equation of state of a perfect gas can be used instead of van der Waal's equation. In either case, the preselected function for computing p*GR,3 is stored in CCU 600.
Where condensate-type refrigerant-vapor trap 446 is not used, or allows a significant amount of refrigerant vapor to enter and condense in reservoir 453, the preselected function for computing p*GR,3 can be chosen so that it takes into account the effect of the presence of liquid refrigerant in reservoir 453. To this end, the independent variables of the last-cited function would also include p*GR and LGR, where LGR is the current level of liquid refrigerant in reservoir 453. The current value of LGR can be obtained by using a proportional liquid-level transducer (not shown). I note that the value of p*GR, in addition to the value of TGR, is needed to compute the solubility of the inert gas in the liquid refrigerant in reservoir 453 because that solubility affects the value of p*GR,3.
An alternative to using transition rule (f), in this section V,G,2,a,i, is to use the transition rule given next:
where T*RSθ is a measure of the actual current value of the refrigerant's saturated-vapor temperature at a location near outlet 471 where inert gas exits the principal configuration, and where T*RSθ is a measure of the saturated-vapor temperature the refrigerant would have H its pressure at outlet 471 were equal to the current value p*Rθ of the total pressure in the principal configuration near outlet 471. The value of TRSθ is lower than that of T*RSθ when inert gas is present at outlet 471 and becomes equal to T*RSθ when no inert gas is present at outlet 471. The current value of TRSθ can be obtained by locating proportional temperature transducer 616 at outlet 471, as shown in
ii. Comments on First Refrigerant & Inert-gas Circuit Configuration, Control System, and Operating Method
Pump 443 would not need to run in mode 3* if no inert gas leaked through pump 443 toward the principal configuration while pump 443 is not running and p*GR is equal to p*GR,MAX. The control-mode rule for pump 443 in mode 3* assumes pump 443 is not leakproof when p*GR is equal to p*GR,MAX, and assumes pump 443 will have to run occasionally, or even continuously (at a very low flow rate), to maintain p*GR close to p*GR,MAX while the system shown in
CR pump 10 is controlled in mode 0* (namely while the engine is not running and cold) so that LP stays close to LPD to ensure liquid refrigerant covers the engine's high heat-flux zones by the time they need to be cooled. Controlling pump 10 in mode 0* would be unnecessary if (1) pump 10 were a zero-slip positive displacement pump (or had in series with it a unidirectional valve (see
GT pump 443 is controlled in mode 0* so that p*R stays close to
p*RDo=pA+Δ*op, (9′)
(where Δ*op is usually a fixed quantity) for the following two reasons: firstly, to compensate for inert-gas leaking through pump 443 while it is not running, and secondly to compensate for changes in ambient-air temperature and pressure. Controlling pump 443 would be unnecessary if (1) it were a zero-slip positive displacement pump (or had in series with it a unidirectional valve (see
Compensating for changes in ambient-air temperature is unnecessary if, when the engine; stops running, the value of p*R is chosen high enough for the current value of p*R, at the design lowest ambient-air temperature, not to fall below the minimum permissible value for p*R. And compensating for changes in ambient-air pressure is unnecessary if, at the design lowest ambient-air pressure, the system does not ingest air and is not damaged by crushing pressures.
Pump 63h, except during mode 1*, is not controlled by the system, but is controlled manually, or automatically, by a thermostat (located in the passenger cabin in the case of a passenger automobile). If control of pump 63h in mode 1* by CCU 600 is not acceptable, an additional refrigerant pump, or a refrigerant valve, and an associated refrigerant-circuit segment, would have to be added where the system employs a non-azeotropic refrigerant to achieve refrigerant-component mixing.
I note that for liquid refrigerant to circulate around refrigerant auxiliary circuit 87h-556-558-559-5-6-8-9-11-12-550′ in mode 1*, point 560 must be higher than point 5. I also note that, where the R&IG configuration shown in
iii. Second Refrigerant & Inert-gas Configuration, Control System, and Operating Method
The specialized principal configurations shown in
The system employs water as its refrigerant, and drives for example an electric generator, installed in a heated building; except for condenser 508, fan 510, air-transfer pump 420, IG variable-volume reservoir 441, and rigid closed cylinder 419′ containing reservoir 441. Cylinder 419′ is located preferably in the shade and may have a finned outer surface. (Fins may often not be necessary.) Cylinder 419′ differs from cylinder 419 in
The system also includes (1) proportional absolute-pressure transducer 603; (2) two-step engine-wall temperature transducer 604 which generates a signal T′W,MAX indicating whether the current engine-wall temperature at a critical high heat-flux zone is close to its design maximum operating value; (3) IG reservoir contact switch 612 which generates a signal V′GR,MAX indicating whether the current value of internal volume VGR of reservoir 441 is at, or close to, its maximum value VGR,MAX; (4) two-step liquid-level transducer 613 which generates a signal L′RD used to indicate whether the liquid refrigerant level, while the principal configuration is inactive, is close to a preselected level L0L′0; (5) spring 614 capable of exerting, whilst fully extended, a force corresponding to a pressure at least equal to the maximum value of p*Ro; and (6) pressure-relief valve 615 set at a value high enough fully to compress spring 614. Transducer 604 may, for example, consist of one or more bimetallic temperature switches. In the case where several bimetallic switches are used, the number of bimetallic switches in multi-cylinder engines would be equal to, or a submultiple of, the number of cylinders. The purpose of receiver 7, which may not be needed, is (1) to keep the liquid refrigerant level substantially below the building's roof, while the system's principal configuration is inactive, and (2) to prevent liquid refrigerant backing-up into separating assembly 42*.
The system is charged with liquid refrigerant until transducer 613 indicates the level of liquid-refrigerant is close to L0L′0, where the level L0L′0 is chosen so that the amount of refrigerant mass MR in the R&IG configuration shown in
The system has three control modes: modes 0*0, 2*, and 3*, where mode 0*0 is, by definition, a minimum-pressure-maintenance mode in which the system controls none of its controllable elements.
In mode 2*, the system's CCU ensures (1) pump 420 is controlled so that p*R tends to p*RD; and (2) fan 510 does not run. And, in mode 3*, (1) pump 420 is controlled so that VGR stays close to VGR,MAX; and (2) fan 510 is controlled so that p*R tends to p*RD. The preselected desired value p*RD for p*R may be fixed, or may change in a pre-prescribed way as a function of one or more characterizing parameters. A typical characterizing parameter, when the engine cooled by the system drives an electric generator, is the mechanical load to which the generator subjects the engine.
The transition rules between modes 0*0, 2*, and 3*, are
where TW0 is low enough to prevent liquid refrigerant being evaporated, and where ΔTW is a small positive value.
I note that transducer 604 can be eliminated if transition rules (c), (d), and (f) are replaced respectively by transition rules
where Δp*R1, and Δp*R2 have fixed positive values, and where the value of TR is provided by a refrigerant temperature transducer which need only be a two-step transducer.
I also note that if points 45* and 8 are high enough above interface surface 123, pump 46 can be eliminated.
iv. Other Refrigerant & Inert-gas Configurations and Control Systems
It should be clear from the teachings so far in this DESCRIPTION, and from my pending U.S. patent application Ser. No. 400,738, filed 30 Aug. 1989, that the class XIFFooo principal configurations shown in
I would explain that principal configurations with a subcooler are desirable, or even necessary, in certain installations to increase—while the system is in mode 3*—the amount of subcool of liquid refrigerant exiting, as applicable, receiver 7, separator 21 or 42, or separating assembly 21* or 42*; and thus increase the net positive suction head available, as applicable, to pump 10, to pump 46, or to pump 27. The subcooler used may merely be a finned quasi-horizontal refrigerant-line segment located roughly in the same plane as refrigerant passages 399, and exposed to ram air and/or to the airflow induced by fan 510. An example of such a finned segment, in the case of a class VIIIFNooo configuration, is segment 9-522 shown in FIG. 43E.
It should also be clear from the teachings so far in this DESCRIPTION that type IIG, IVG, and VG, IG configurations can also be used with type C combinations and may be preferred IG configurations in certain installations.
It should further be clear from the foregoing teachings that a shutter can be used upstream from a condenser of a type C combination, as well as upstream from a condenser of a type A combination, to control the rate at which the condenser releases heat. The shutter, if desirable, can be made of thermally-insulating material to accelerate engine warm-up in cold climates. How shutter-control led heat release can be accomplished with type C combinations should be clear in view of the earlier discussion in this DESCRIPTION of shutter-control led heat release with type A combinations.
b. Cooling Systems with a Non-pool Evaporator
i. Preliminary Remarks
Type C combinations, in common with type A combinations, are suitable for a much wider range of piston-engine cooling applications when they have an NP evaporator instead of a P evaporator because they can be used with any cylinder orientation and impose much less stringent constraints on the tilts of the platform on which they are installed.
ii. Refrigerant & Inert-gas Configuration and Control System
The first system chosen as an example has the class IIIFNoo principal configuration and the type IVG ancillary configuration, shown in
The system has the following transducers: (1) two-step liquid-level transducer 613; (2) two-step liquid-level transducer 622; (3) proportional absolute-pressure transducer 603; (4) proportional engine-wall temperature transducer 634; and (5) two-step absolute-pressure transducer 626. The signals generated by the foregoing five transducers are supplied to the system's CCU.
Signal L′R0, generated by transducer 613, is used to indicate whether the system is charged with a correct amount of liquid refrigerant. (To this end, transducer 613 is located at level L0L′0, where L0L′0 is the correct liquid-refrigerant level while the system's principal configuration is inactive.) Signal L′RR, generated by transducer 622, is used to indicate whether liquid refrigerant—draining out of fixed-volume IG reservoir 453 through one or more ports 623 at the bottom of the cylindrical part 624 of reservoir 453—has reached in vessel 625 a preselected release level LRR determined by the location of transducer 622. Signal p*′GR,MAX, generated by transducer 626, indicates p*GR has reached its design maximum operating value p*GR,MAX. And signals p*′R and T′W, generated by transducers 603 and 634, respectively, are used by the system's CCU to generate signals C′PC, C′GT, C′CF, C′RDV1, C′RDV2, C′RR, and C′SC, used to control respectively DR-pump clutch 621, GT pump 443, condenser fan 510, liquid-refrigerant diverter valve 555, liquid-refrigerant diverter valve 630 having an inlet 631 and outlets 632 and 633, liquid-refrigerant release (drain) valve 487, and SC pump 63h.
The class XI*FNoooo principal configuration shown in
The type IV IG configuration shown in
The system's R&IG configuration is first charged with liquid refrigerant until transducer 613 generates signal L′R0 indicating the refrigerant liquid level in refrigerant-vapor line 44-5 has reached level L0L′0; and is then charged with inert gas until the R&IG configuration's internal pressure p*R, reaches a preselected value p*Ro, where the preselected value p*Ro may be different for different R&IG-configuration mean temperatures. Valve 477 is kept open by say a manual override, while the system's R&IG configuration is being charged with refrigerant and subsequently with inert gas.
iii. Unsafe and Safe States
I shall say that the system is in an unsafe state when relation (3′) or (4) is satisfied, and that the system is in a safe state when relations (3′) and (4) are satisfied.
iv. Typical Operating Method
The system, while in a safe state, has six control modes, namely modes 0*0A, 0*0B, 1*A, 1*B, 2*, and 3*.
Mode 0*0A is a minimum-pressure-maintenance mode while the engine is not running and corresponds to control mode 0*0 in section V,G,2,a,iii. And mode 0*0B is a minimum-pressure-maintenance mode while the engine is running but cold, and the effective capacity of pump 46 is zero although the engine is running. The purpose of mode 0*0B is to accelerate engine warm-up, while TW is below a preselected value TWD1 by supplying no liquid refrigerant to the engine's coolant passages while the value of TW is less than TTWD1. TWD1 is a preselected value of TW substantially lower than the maximum permissible value TW,MAX of TW, and higher than the value TRSo, of the saturated-vapor temperature TRS, corresponding to p*RDo. (TWD1 may, for example, be 120° C.).
Mode 1*A is used to achieve the same purpose as control mode 1*, namely is used to mix the components of a non-azeotropic refrigerant so that the concentrations of their liquid phases are approximately spatially uniform. However, the particular R&IG configuration shown in
Liquid-refrigerant diverter valve 655h and cabin-heating fan 552h are controlled manually or automatically by the cabin climate-control system. Valve 477 is operated in the same way in all modes where the system's CCU is energized, namely in all modes except mode 0*0A. And the clutch of pulley-and-clutch 621 is engaged in all control modes except mode 0*0B. The remaining system-controllable elements are controlled as described next.
In mode 0*0A, no system-controlleable elements are controlled.
In mode 0*0B, (1) pump 443 is controlled so that p*R tends to p*RDo; (2) fan 510 does not run; (3) valve 555 is in position 1, namely liquid refrigerant entering at 556 exits at 557; and (4) valve 630 is in position 1, namely liquid refrigerant entering at 631 exits at 633.
In mode 1*A, (1) pump 443 is controlled so that p*R tends to p*RDo; (2) fan 510 does not run; (3) valve 555 is in position 2, namely liquid refrigerant entering at 556 exits at 558; and (4) valve 630 is in position 1.
In mode 1*B (1) pump 443 is controlled so that p*R tends to p*RDo; (2) fan 510 runs at a preselected effective capacity, or at a preselected speed; (3) valve 555 is in position 1; and (4) valve 630 is in position 2, namely liquid refrigerant entering at 631 exits at 632.
In mode 2*, (1) pump 443 is control led so that TW tends to TWD; (2) fan 510 does not run; (3) valve 555 is in position 1; and (4) valve 630 is in position 2.
In mode 3*, pump 443 is controlled so that p*GR stays close to p*GR,MAX; (2) fan 510 is controlled to that TW tends to TWD; (3) valve 555 is in position 1; and (4) valve 630 is in position 2.
The transition rules between control modes are:
In transitions (e) and (r), ΔTW1 and ΔTW2, respectively, are small positive values.
I note that the value of TW1, and the current value of TW in modes 2* and 3*, must be high enough to ensure p*R does not fall below its minimum-permissible value p*R,MIN even during transients. If the last-cited constraint is not practicable, or is not desirable, the CCU, whenever p*R falls below (p*R,MIN+εP) where εP is a small positive quantity, causes the control signal C′GT to control pump 443 so as to maintain (the value of) p*R at or above p*R,MIN until p*R exceeds, say, (p*R,MIN+2 εP). The action described in the immediately-preceding sentence amounts to using two new modes 2*0 and 3*0 with the following transition rules:
Where condenser 510, receiver 7, and dual-return receiver 640, are mounted high enough above refrigerant passages 504 and 505 to ensure a,high-enough liquid-refrigerant flow-rate at 2′ and 2″ to prevent hot spots occurring without using pumps 46H and 46B, these pumps can, in principle, be eliminated. Whether or not the resulting R&IG configuration is a preferred configuration depends on the details of the application of interest. Examples of applications where it would be practicable to mount condenser 510, receiver 7, and dual-return receiver 640, above refrigerant passages 504 and 505 to achieve high-enough flow-rates at 2′ and 2″ include installations in certain trucks.
v. Other Refrigerant & Inert-gas Configurations and Control Systems
Depending on the application considered, other principal configurations which may be preferred include class IIFNooo, IIFNsoo, IIFFooo, IIFFsoo, IIFFs′oo, IIFFs″oo, II*FNooo, II*FNsoo, IIIFNso, IIIFFoo, IIIFFso, III*FNoo, and III*FNsoo, configurations, and other preferred IG configurations include type IG, IIG, and VG, configurations.
3. Intercooling Systems with an Air-cooled Condenser
a. General Remarks
The remarks made about piston-engine intercoolers in section V,F,3,a apply also to intercoolers whose airtight configurations are type C combinations with the exception of the remarks made in the third major paragraph of section V,F,3,a.
In the case where minimum-pressure maintenance, gas-controlled heat release, and a co fast-response capability are required, and where a non-azeotropic fluid is employed, the operation of an intercooler using a type C combination with an air-cooled condenser can be described in terms of control modes o*E, 0*S, 1*, 2*, and 3*, where control modes 0*E and 0*S correspond to control modes 0E and 0S, respectively, of a fast-response intercooler having a class A combination.
b. A First Fast-response Intercooler
I describe in this section V,G,3,b the operation of an intercooler having, see
I shall, in this section V,G,3,b, refer to the intercooling system comprising the R&IG configuration shown in
Four-way slide-type refrigerant-flow reversing valve 660i, having inlet-outlet ports 661i and 662i, is used to reverse the direction of the liquid-refrigerant flow rate induced, in refrigerant-circuit segment 49i-661i-662i-2i, by engine-driven DR pump 46i; and proportionally-controllable DR-pump recirculation valve 663i is used to control the effective capacity of pump 46i when liquid refrigerant flows from port 662i to port 661 i. Unidirectional GT pump 443Ai, and bidirectional (two-way) GT valve 475i, are used to control the transfer of inert gas (and associated refrigerant vapor) between the principal and the inert-gas configurations shown in
The intercooler has four control modes designated by the symbols 0*E, 0*S, 2*, and 3*. In mode 0*E the intercooler is in its minimum-pressure-maintenance mode while the engine is not running.
In mode 0*S the intercooler is in its combined minimum-pressure-maintenance and fast-response-preparation mode. In mode 0*S, heat from the engine's exhaust gases is used, while the engine's supercharger is not running, to ensure the current value of TIi stays close to a preselected desired value TIDi above the air's ambient temperature. This, among other advantages, allows minimum-pressure maintenance to be achieved with less inert-gas mass in the intercooler's principal configuration than that which would be required to achieve minimum-pressure maintenance at ambient temperature. And this, in turn, allows the intercooler to reach, if required, its design maximum heat-transfer capacity (under prevailing conditions) faster after the engine's supercharger starts running. During mode 0*S the engine's exhaust gases are circulated at a rate controlled by exhaust-gas damper 567i, around exhaust-gas circuit 566i-666i-667i-568i, where exhaust-gas inlet 566i is upstream from exhaust-gas return 568i with respect to the direction of flow of exhaust gas in pipe 565. Engine exhaust gas circulated in the last-cited circuit releases heat, while in exhaust-gas circuit segment 666i-667i, to liquid refrigerant in separator 42i. (Fins may be used in that segment to augment the heat-transfer rate to liquid refrigerant in separator 42i.) In mode 0*S separator 42i performs the function of a pool evaporator and evaporator 561i performs the function of a condenser.
In mode 2* the intercooler is in its combined gas-controlled heat-release and self-regulation mode. And, in mode 3*, the intercooler is in its combined fan-controlled heat-release and self-regulation mode.
The system-controllable elements in
In mode 0*E, (1) valve 660i is in position 2, namely valve 660i would cause refrigerant to flow from port 662i to port 661i if DR pump were running and valve 663i were not wide open; (2) valve 663i is in a preselected position; (3) fan 510i does not run; (4) damper 567i is in a preselected position; and (5) pump 443Ai and valve 475i are controlled by the system's MPMCU so that p*Ri tends to p*RDoi.
In mode 0*S, the system's CCU ensures (1) valve 660i is in position 2; (2) valve 663i is controlled so that LDi tends to its desired level LDDi; (3) pump 443Ai and valve 475i are controlled so that p*Ri tends to p*RDoi; (4) fan 510i does not run; and (5) damper 567i is controlled so that TIi tends to TIDi.
In mode 2*, the system's CCU ensures (1) valve 660i is in position 1, namely valve 660i causes refrigerant to flow from port 661i to port 662i; (2) valve 663i is closed; (3) pump 443Ai and valve 475i are controlled so that TIi tends to TIDi; (4) fan 510i does not run; and (5) damper 567i is closed.
In mode 3*, the system's CCU ensures (1) valve 660i is in position 1; (2) valve 663i is closed; (3) pump 443Ai and valve 475i are controlled so that p*GRi stays close to p*GR,MAXi; (4) fan 510i is controlled so that TIi tends to TIDi; and (5) damper 567i is closed.
The transition rules between the last-cited five control modes are:
I note that where mode 1* is required because the refrigerant employed is a group H refrigerant, merely adding means for circulating the refrigerant in a way similar to the way shown in
I would add that, except for an electrical heat source, an engine's exhaust gas is usually the heat source in an automotive vehicle whose temperature rises fastest when the engine is cold. However, where a greater delay is permissible in supplying heat to an intercooler in mode 0*S or in mode 0S, as applicable, the engine's coolant or the engine's lubricating oil can be used to supply heat to an intercooler's refrigerant during either of the two last-cited modes.
C. A Second Fast-response Intercooler
Applications where the three conditions recited in the third minor paragraph of section V,F,1 are satisfied, are examples of applications where refrigerant vapor exiting an evaporator can be allowed to be dry, and where in particular superheat-control techniques can be used to control the CR pump of type A and type C combinations having group 1, or group IV, principal configurations. The last-cited techniques are described in detail in section V,B,3,b,ii of my pending U.S. patent application Ser. No. 400,738, filed 30 Aug. 1989. In this section V,G,3,c I describe a particular way of implementing those techniques in the case of a type C combination having the R&IG configuration shown in FIG. 62B. However, it should be clear from my teachings so far in this DESCRIPTION that the superheat-control techniques described in section V,B,3,b,ii of the last-cited application, and the particular way of implementing those control techniques described in this section V,G,3,c, can also be used with type A combinations having a group I or a group IV principal configuration.
The particular way of implementing superheat-control techniques shown in
The system having the R&IG configuration shown in
In mode 0*E, pump 10i and fan 510i do not run, valve 677i is closed, and switch 684i is open; and the system's MPMCU (not shown) controls pump 443Ai and valve 475i so that p*Ri tends to p*Roi.
In mode 0*S, the system's CCU (not shown) ensures (1) pump 10i is controlled (on-off) so that the liquid-refrigerant level Lyi of liquid-vapor interface surface 687i in vessel 686i stays within an upper limit Ly,MAXi and a lower limit Ly,MINi with the help of signal L′yi generated by three-step liquid-level transducer 688i; (2) pump 443Ai and valve 475i are controlled so that p*Ri, tends to p*RDoi; (3) fan 510i does not run; (4) valve 677i is controlled so that TIi tends to TIDi; and (5) switch 684i is closed. (The heat supplied by heating element 685i to thermostatic element 679i, while switch 684i is closed, causes valve 678i to stay wide open and to allow refrigerant vapor to enter evaporator 1i where it is condensed and returned by sewer flow to vessel 686i.)
In mode 2*, the system's CCU ensures (1) pump 10i runs; (2) pump 443Ai and valve 475i are controlled so that TIi tends to TIDi; (3) fan 510i does not run; (4) valve 677i is open; and (5) switch 684i is open.
In mode 3*, the system's CCU ensures (1) pump 10i runs; (2) pump 443Ai and valve 475i are controlled so that p*GRi tends to p*GR,MAXi; (3) fan 510i is controlled so that TIi tends to TIDi; (4) valve 677i is open; and (5) switch 684i is open.
Pressure regulator 689i ensures pump 10i delivers liquid refrigerant to valve 678i at a preselected refrigerant pressure.
d. Alternative Intercoolers
In view of the extensive descriptions and discussions already given in this DESCRIPTION of the operation of piston-engine intercooling systems using type A combinations, and of the operation of piston-engine intercooling systems using type C combinations, it should be apparent, to those skilled in the art, how they could operate intercoolers using other principal configurations disclosed in this DESCRIPTION and other inert-gas configurations disclosed in this DESCRIPTION.
It should, in particular, also be apparent that an intercooler with a type C combination can, where desirable, also use shutter-controlled heat release, in addition to gas-controlled heat release, during its fast-response preparation mode to minimize the rate at which the intercooler condenser releases heat during the last-cited mode.
4. Cooling Systems with a Water-cooled Condenser
a. General Remarks
A first principal difference, in piston-engine cooling applications, between type C combinations having a water-cooled condenser and type C combinations having an air-cooled condenser, is that the former combinations can use water-controlled heat release whereas the latter combinations obviously cannot use water-controlled heat-release. Water-controlled heat release is usually adequate by itself for achieving heat-release control and therefore refrigerant-controlled heat release is usually not needed.
A second principal difference is usually the same as the second principal difference stated in the second minor paragraph of section V,F,4,a.
b. Refrigerant-circuit Configuration, Control System, and Operating Method
The R&IG configuration shown in
The DR pump of the principal configuration shown in
I would mention that the coolant flow rate entering a component evaporator need not be controlled by the injector, and may be continuous instead of being pulsed. This is particularly true in the case of cylinder-block component evaporators where pulsed injection is often unnecessary and the coolant flow rate entering a component evaporator through an LR injector—if one is used—is controlled, as in continuous fuel-injection systems, remotely through a coolant-metering device. A technique for preselecting the time-average flow rate delivered by LR injection nozzles as a function of operating conditions is discussed in section V,H,3.
The cooling system having the refrigerant-circuit configuration shown in
In mode 0*0, the system, by definition, controls none of the R&IG configuration's controllable elements.
In mode 2*0, (1) injectors 531′a and 531′b are controlled so that, in effect, the quality q′EV of refrigerant vapor exiting at 3′a and 3′b stays within a first pair of preselected limits, and injectors 531″a and 531″b are controlled so that the quality q″EV of refrigerant vapor exiting at 3″a and 3″b stays within a second pair of preselected limits; (2) pump 443 is controlled so that p*R tends to p*RD; (3) pump 598 does not run, and (4) valve 690 is in a preselected position. (Air-cooled condenser 456 is assumed to be capable of removing by itself refrigerant vapor entering at 457 while the system is in mode 2*0, and therefore pump 598 is not running.)
In mode 3*, (1) injectors 531′a and 531′b, and injectors 531″a and 531″b, are controlled in a way similar to the way they are controlled in mode 2*0, except that the preselected limits may be different; (2) pump 443 is controlled so that the value of p*GR stays close to p*GR,MAX; (3) pump 598 is controlled so that p*R tends to p*RD; and (4) valve 690 is controlled so that the value of ΔT stays above a preselected value indicating the absence of refrigerant vapor in condenser 459.
The transition rules between the foregoing three modes are:
The positive quantities of ΔTW1, ΔTW2, and ΔTW3, need not necessarily be different.
c. Other Refrigerant & Inert-gas Configurations and Control Systems
All the classes of principal configurations, and all the types of IG configurations, described or listed in section V,G,2,b can also be used with R&IG configurations having a water-cooled condenser.
5. Cabin Heating
Cabin heating with type C combinations can, where desired, be achieved either by single-phase heat transfer or by two-phase heat transfer. Techniques for cabin heating, in the case of a type C combination using single-phase heat transfer, have already been discussed in section V,G,2 where SC pump 63h, of the cabin-heating circuits shown in
Examples of the last-cited techniques were given in section V,F,2,g for the case of type combinations. The techniques used in the case of type C combinations are similar. Suitable locations in type C combinations for tapping off refrigerant vapor for cabin-heating two-phase heat-transfer circuits include a suitable point of their evaporator or a suitable point of their refrigerant-vapor transfer means including, as applicable, their separator or their separating assembly.
The cabin-heating circuit shown in
1. Preliminary Remarks
So far I have, for specificity, described control methods for embodiments of the invention in the context of either a type A or a type C combination. In this section V,H, I discuss techniques common to both the two last-cited combinations.
For brevity, where I do not wish to distinguish between pR and p*R, I shall in this section V,H refer to either pR or to p*R as PRu. Also for brevity, where I do not wish to distinguish between (control) modes 2 and 2*, between (control) modes 20 and 2*0, or between (control) modes 3 and 3*, I shall refer to either of the first two modes as mode 2u, to either of the second two modes as mode 20u, and to either of the third two modes as mode 3u.
2. Preselection of Desired Refrigerant Pressure
I stated earlier in this DESCRIPTION that the preselected desired value p*RD of the refrigerant pressure pR may be fixed, but may also change in a pre-prescribed way as a function of one or more preselected parameters. These include one or more parameters characterizing the current state of an engine and/or the current state of an engine's environment. Useful parameters characterizing the current state of the engine include (1) fuel mass-flow rate {dot over (m)}F or almost equivalently fuel volumetric-flow rate FF; (2) intake-air mass-flow rate {dot over (m)}I; (3) engine (rotational) speed ωE; (4) knocking intensity kE; (5) intake-air temperature TI; (6) intake-air pressure pI; (7) throttle position θT; and (8) the derivatives of the quantities cited under (1) to (7). And useful parameters characterizing the state of the engine's environment include (9) ambient-air pressure pA; (10) ambient-air temperature TA; (11) local solar radiation intensity TS; (12) ambient-air relative humidity HA; and (13) the derivatives of the quantities cited under (9) to (12). I note that measures of certain parameters characterizing the state of an engine can be indirect measures. For example, a suitable measure of {dot over (m)}F, in the case of an engine with pulse-width controlled fuel injection, is the pulse width of the injection-control signal; and a suitable measure of ωE, in the case of a spark-ignition engine, is the rate of the firing signal. I also note that, in the case of an unsupercharged and unthrottled engine, pA and TA may be sufficiently accurate measures of pI and TI and vice versa.
The preferred pre-prescribed way for varying PRDu as a function of one or more of the foregoing characterizing parameters depends greatly on the particular engine being cooled. A preferred pre-prescribed way, while the engine's cooling system is in mode 2u, in mode 20u, or in mode 3u, would include, in the case of an engine with a knocking-intensity sensor,
The chosen pre-prescribed way for varying pRDu as a function of preselected characterizing parameters is stored in a cooling system's CCU.
The minimum-permissible value of pRU, with most existing engines, is currently (1991) usually governed, when the current value of pRU is lower than the current value of pA, by the maximum value of |pA−pRu| for which an airtight two-phase cooling system is affordable (although it may in future be governed by other considerations). Because the value of pA decreases when altitude increases, the value of pR,MINu also decreases when altitude increases. The minimum value of pR,MINu at any altitude, can be determined by measuring pRu and pA and requiring pR,MINu to satisfy, when pRu is lower than pA, the relation
|pA−pR,MINu|≦ΔMAX,1p, (21)
where ΔMAX,1p is the maximum value of the amount by which pRu is allowed to fall below the current value of the ambient atmospheric pressure pA. The maximum-permissible value pR,MAXu of pRu, when pRu is higher than pA, is governed either by the maximum-permissible value of {overscore (T)}RS,E under specified engine and environmental conditions, where {overscore (T)}RS,E is the refrigerant's mean saturated-vapor temperature in the evaporator, or is merely governed by the maximum-permissible value of ΔMAX,2p, where ΔMAX,2p is the maximum value of the amount by which pRu is allowed to rise above the current value of pA Where a piston-engine cooling system's refrigerant and airtight-configuration design have been selected so that ΔMAX,2p is not exceeded, for the highest values of TRS at only low altitudes (say at altitudes up to 500 meters), the maximum value of pR,MAXu and the corresponding value or values of TRS must be limited at higher altitudes so that the relation
|pR,MAXu−pA|≦ΔMAX,2p (22)
is still satisfied at those higher altitudes. The maximum value of pR,MAXu, at any altitude, can be determined by measuring pRu and pA and requiring pR,MAXu to satisfy, when pRu is higher than pA, relation (22).
The invention includes providing means not only for measuring the current values of pRu and pA with two proportional absolute-pressure transducers, or of the current value of the difference (pRu−pA) with one proportional differential-pressure transducer; but also for
I note that, in general, the principal purpose for varying pRDu in a pre-prescribed way as a function of one or more parameters characterizing an engine's state is to achieve, at one or more of n preselected locations, a desired preselected (time-averaged) engine-wall temperature TWD which may be fixed, but which is usually changed in a pre-prescribed way as a function of one or more preselected characterizing parameters. However, achieving a desired value TWD of TW by controlling pRu is a very inaccurate process, particularly in the case of engines whose speed and torque vary over a wide range of values. The reason for this is that (TW−TRS), where TRS is the current value of the refrigerant saturated-vapor temperature corresponding to pRu, can be inferred, particularly during transients, only approximately from parameters characterizing the engine's state even in the case of azeotropic-like refrigerants. It follows that, where practicable, it would be desirable to measure TW at a critical point of each of the engine's one or more cylinder heads and to control pRu, while a piston-engine cooling system is in modes 2u, 20u, or 3u, so that TW, the average current value of the engine-wall temperatures at each of those critical points, tends to TWD. The invention, where practicable and affordable, comprises means for obtaining a measure of TW which includes using one or more proportional temperature transducers to generate signals T′W1 to T′Wn providing a measure of wall temperatures at the n points where they are located. Examples of suitable points in engines with two exhaust valves per cylinder are the exhaust-valve bridges. Thermistors or thermocouples, with properly protected wiring in refrigerant passages 505, could be used as the sensors of the temperature transducers used to generate signals T′W1 to T′Wn. Critical points are usually the points of an engine where the heat flux is highest. Where locating transducers at the last-cited points is impracticable or too expensive, the proportional temperature transducers cited earlier in this minor paragraph can be located at points in an engine's structure in the general neighborhood of the highest heat flux points, and the temperatures at the critical points can be estimated by the CCU of an engine-cooling system of the invention from the temperatures of the points where the proportional temperature transducers are located. Also, where it is too expensive to obtain measures of the temperatures at or near the combustion-chamber walls of each cylinder of a multicylinder engine, the number of proportional transducers used may be smaller than the number of cylinders of that engine.
The current value of TW is obtained by the CCU of a system of the invention by taking TW equal to
and by using one or more controllable elements to make TW tend to TWD in control modes 2u, 20u, or 3u. An example of such a control method was given in section V,G,2,b for the case of a type C combination with an NP evaporator.
4. Engine-driven Pumps
a. Preliminary Remarks
In the case where a system of the invention is used to cool a device generating mechanical power, namely to cool a motor, the most cost-effective means of driving a pump of the system is often to drive it by that device. This statement is true in particular where the mechanical-power generating device or motor is an internal-combustion engine or an electric motor, and applies to all the pumps of a system of the invention, including refrigerant pumps, inert-gas pumps, air-transfer pumps, hydraulic pumps, hot-fluid pumps, and cold-fluid pumps.
b. Principal-configuration Refrigerant Pumps
In general the cooling load of a variable-speed engine, or of a nominally constant-speed engine, is not only a function of the engine's speed ωE, but is also a function of one or more characterizing parameters such as the other charactizing parameters mentioned in section V,H,2. It is therefore usually—albeit not always—highly desirable that an engine-driven refrigerant pump be provided with means for changing its effective capacity, at a given engine speed. These means include a proportional bidirectional (two-way) refrigerant valve controlled by a modulated analog signal, or by a modulated pulsed signal. A pulsed signal can be modulated by varying one or more of the following three quantities: pulse width, pulse amplitude, and pulse rate (or synonymously pulse frequency). The refrigerant valve used to change the effective capacity of an engine-driven refrigerant pump may be a valve in series with the pump or a valve in parallel with the pump. In the former case, the valve is used as a throttling valve to modulate the flow rate through the pump. And, in the latter case, the valve may be used only as a recirculation valve in a circuit used exclusively as the pump's recirculation circuit; or the valve may also be used to control the flow rate of the fluid through the valve while the pump is inactive. (A pump recirculation circuit may be an integral part of the pump).
A typical method of sizing an engine-driven pump in the case of a variable-speed engine is
c. Ancillary-configuration, Inert-gas-configuration, Hot-fluid, and Cold-fluid Engine-driven Pumps
The effective capacity of the pumps cited in the immediately-preceding heading can be adjusted by using techniques similar to those used with principal-configuration refrigerant pumps. Additionally, the effective capacity of those pumps can, where they pump a gas, also be adjusted by a throttling valve upstream from the pump.
5. Evaporator Refrigerant Flow-rate Control
a. Preliminary Remarks
The proper control of the mass-flow rate {dot over (m)}E flowing through a unitary evaporator, or the mass-flow rate {dot over (m)}Ej flowing through component evaporator j of a split evaporator, is of crucial importance in most systems of the invention.
I distinguish between ‘non-overflow P evaporators’ on the one hand, and NP evaporators and ‘overflow P evaporators’ on the other hand. (For definitions of the two terms in quotation marks see the last minor paragraph of section V,B,10.) The purpose of controlling {dot over (m)}E and {dot over (m)}Ej in the case of non-overflow P evaporators is to maintain the level of interface surface 123 (see for example
b. Evaporator-overfeed Control
Evaporator-overfeed control techniques, where employed, are used in piston-engine intercooling applications, and in general in cooling and heating systems having the characteristics recited in the third minor paragraph of section V,F,1, merely to obtain, at a given instant, a mean refrigerant heat-transfer coefficient higher than that achieved with an evaporator-overfeed ratio equal to zero. By contrast, evaporator overfeed-control techniques are used in piston-engine cooling systems, and in general in cooling and heating systems having the characteristics recited in the second minor paragraph of section V,F,1, to ensure their feasibility. I next elaborate, for specificity, on the last-cited control techniques in the context of piston-engine cooling systems. But those techniques apply mutatis mutandis to all cooling and heating applications where evaporator-overfeed control is desirable.
NP evaporator (or component NP evaporator) overfeed control is used in piston-engine cooling systems, as mentioned in the first major paragraph of section V,F,2,b,ii, to ensure, with all refrigerants, that no hot spots occur; and also to ensure, with non-azeotropic refrigerants, that the concentrations of their components in an NP evaporator are sufficiently uniform spatially to prevent an unacceptably-large rise in the refrigerant's saturated-vapor temperature TRS as it flows through a unitary evaporator, or through each of the component evaporators of a split evaporator. NP evaporator overfeed can also be used, where required, to increase the mass of refrigerant in an NP evaporator, and thereby cause (see discussion in section V,F,2,d) the value of (TRS,EA−TRS,0) to be small enough for it to be acceptable.
Correct evaporator overfeed requires achieving, as applicable, one or more of the three purposes recited in the immediately-preceding minor paragraph without using undesirably-high evaporator-overfeed ratios, particularly at high engine cooling loads; where, by definition, an engine's cooling load is the rate {dot over (Q)}C at which heat generated by the engine must be removed by the engine's two-phase heat-transfer cooling system; and does not include the rate at which heat is removed from the engine by other means including (1) the rate at which heat is removed by cooler ambient air by convection, or to cooler material things surrounding the engine by radiation, and (2) the rate at which heat is removed by the engine's lubricating system where the lubricating system's oil is not cooled by the engine's two-phase heat-transfer system. Evaporator-overfeed ratios are undesirably high when they exceed the ratios required to achieve, as applicable, one or more of the foregoing three purposes and, as a result, cause (1) a larger or more expensive separator, condenser, and/or condenser fan, to be used, or (2) the condenser fan to run more often or at a higher rate.
The preselected evaporator-overfeed ratio rEO,D for achieving the applicable purposes of interest for a given engine can be obtained from tests on that engine. The value of rEO,D, under steady-state conditions, may be fixed, or may change as a function of one or more parameters characterizing the state of the engine; for example the preselected value of rEO,D may increase with {dot over (m)}F and vice versa.
The EO-pump mass-flow rate {dot over (m)}EO,D required to achieve rEO,D is given by
{dot over (m)}EO,D=rEO,D·{dot over (m)}θ (23)
where {dot over (m)}θ is the refrigerant evaporation rate in an NP evaporator; and the DR-pump mass-flow rate {dot over (m)}DR required to achieve rEO,D is given by
{dot over (m)}DR,D=(1+rEO,D)·{dot over (m)}θ. (24)
The current value of {dot over (m)}θ can be obtained, with negligible time delays, from the signal F′V generated by a refrigerant-vapor flow-rate transducer located in an airtight configuration's refrigerant-vapor transfer means as shown, for example in
The current value of {dot over (m)}θ, where the amount of refrigerant subcooling and superheating is negligible, can also be obtained quite accurately by assuming {dot over (m)}θ is equal to {dot over (Q)}C/hlg, where hlg is the latent heat of evaporation of the refrigerant. This is often the case with internal-combustion engine-cooling systems because, in those systems, the amount of refrigerant subcooling is usually negligible and the amount of refrigerant superheating is zero. Where the amount of refrigerant subcooling of the refrigerant condensate mass-flow rate {dot over (m)}Cis significant but refrigerant superheating is negligible, {dot over (m)}θ can be estimated quite accurately by using
{dot over (m)}θ({dot over (Q)}C−cplm{dot over (m)}CΔSbT)/hlg (25)
where cpl is the specific heat of liquid refrigerant and ΔSbT is the amount by which refrigerant condensate is subcooled. The value of hlg as a function of pR can, in the case of an azeotropic-like refrigerant, be determined from published tables; and, in the case of a non-azeotropic refrigerant, from published tables and from the estimated concentrations of the refrigerant's components in the NP evaporator.
The current value of {dot over (Q)}C under transient conditions as well as under steady-state conditions, could in principle be predicted by determining during tests the functional dependence of {dot over (Q)}C on a subset of applicable and non-redundant parameters selected from a set of characterizing parameters including TW and {dot over (T)}W, and the parameters listed under (1) to (13) in section V,H,2. The number of characterizing parameters employed in estimating {dot over (Q)}C depends on the desired accuracy.
In practice, determining the functional dependence of {dot over (Q)}C on parameters characterizing the state of an internal combustion engine during transients is often impracticable. Consequently, the invention envisages determining the functional dependence of {dot over (Q)}C on preselected characterizing parameters during tests conducted under steady-state conditions, and using rough empirical rules for ensuring {dot over (q)}EV does not exceed {dot over (q)}EV,MAX during transients. For example, the value of {dot over (m)}EC, or of {dot over (M)}DR, obtained by using values of {dot over (Q)}C, determined during steady-state tests, could be increased during transients by Δ{dot over (m)}EC, or by Δ{dot over (m)}DR, where either of these quantities is proportional to the absolute value of one or more of the derivatives of relevant steady-state parameters. For example, Δ{dot over (m)}EO or Δ{dot over (m)}DR may be made proportional to the absolute value |{dot over (m)}F of {dot over (m)}F, or where applicable the absolute value |{dot over (θ)}T| of θT, where the coefficient of proportionality is determined empirically. This would temporarily increase the value of {dot over (m)}EO, or of {dot over (m)}DR, above its last steady-state value when, as applicable, {dot over (m)}F or QT is increased, and would temporarily maintain the current value of the last steady-state value of {dot over (m)}EO or of {dot over (m)}DR when, as applicable, {dot over (m)}F or θT is decreased. Thus, for example, where {dot over (m)}θ is taken equal to {dot over (Q)}C,SS/hlg, and where {dot over (Q)}C,SS is the value of {dot over (Q)}C obtained from tests conducted under steady-state conditions, the expressions
{dot over (m)}EO,D=(rEO,D{dot over (Q)}C,SS)/hlgkC1|{dot over (m)}F| and {dot over (m)}DR,D={(1+rEO,D)·{dot over (Q)}C,SS/hlg}+kC2|{dot over (m)}F|, (26),(27)
where kC1 and kC2 are positive constants, can be used to offset cooling-system response lags to a sudden increase in fuel flow rate, and to offset engine thermal lags to a sudden decrease {dot over (m)}F in fuel flow rate. The same technique can be used to offset lags in the value of {dot over (m)}C with respect to the value of {dot over (m)}θ where {dot over (m)}C is used instead of {dot over (Q)}C,SS in relations (26) and (27).
The relation used to control EO pump 27, or DR pump 46, is stored in the CCU of a system of the invention; the characterizing parameters used in that relation are obtained from transducer signals and supplied to the CCU; and a signal C′EO, or signal C′DR, is generated by the CCU which controls EO pump 27 or DR pump 46, so that {dot over (m)}EO, or {dot over (m)}DR, tend respectively to {dot over (m)}EO,D or to {dot over (m)}DR,D.
Evaporator overfeed can further be used for a fourth purpose, namely to decrease the value of (TW×TRS,E) in high heat-flux zones at high cooling loads. This allows the value of TRS,E to be increased, at high cooling loads, for a given maximum value of TW and a given heat flux, thereby allowing the size of an airtight configuration's condenser to be reduced for a given cold-fluid pump power. (The cold-fluid pump is usually a fan or a water pump.) Alternatively, this allows the value of TW to be decreased, at high cooling loads for a given value of TRS,E and a given cold-fluid pump power, thereby allowing the engine's volumetric efficiency to be increased at high cooling loads and at high engine power. I shall refer to the overfeed used to achieve the foregoing fourth purpose as ‘excess overfeed’ because it exceeds the amount of overfeed required to achieve, as applicable, one or more of the three purposes cited in the first minor paragraph of the second major paragraph of this section V,H,5,b; and is undesirably high in the sense the qualifier ‘undesirably high’ is used in the second minor paragraph of the second major paragraph of this section V,H,5,b. I distinguish between ‘excess overfeed’ and ‘incorrect overfeed’. I use the latter term in the case where excess overfeed is not desired and the amount of evaporator overfeed is undesirably high.
In the case where an NP evaporator has several sets of component evaporators, and one of those sets has much higher heat-flux zones than the other one or more sets, excess overfeed is usually employed only with the set of component evaporators having the highest heat-flux zones and TW, in the expression (TW−TRS,E), is the wall temperature of the most critical of those high heat-flux zones. In the particular case of a piston engine with non-interconnected cylinder-block and cylinder-head coolant passages, an NP evaporator could, for example, have two sets of component evaporators: a set of cylinder-block component NP evaporators and a set of cylinder-head component NP evaporators. In that particular case excess overfeed would usually be employed only with the latter set of component evaporators, and TW would be the average wall temperature of a set of critical heat-flux zones of that latter set of component evaporators. In the case of an engine with a single bank of cylinders, the set of cylinder-head component evaporators may consist of only one component evaporator.
Whereas correct overfeed applies to control modes 2u and 3u, excess overfeed usually applies only to mode 3u, but need usually not be employed continuously in mode 3u. Consequently, mode 3u is in effect split into two control modes: mode 3Cu where correct overfeed is employed and mode 3Eu where excess overfeed is employed, and transition rules between those two modes must be formulated. Examples of transition rules between modes 3Cu and 3Eu are discussed next.
Assume for specificity that the system of the invention of interest has—like the system shown in FIG. 63—a set of cylinder-block component evaporators supplied collectively by liquid refrigerant at a mass-flow rate {dot over (m)}EB and a set of cylinder-head component evaporators supplied collectively by liquid refrigerant at a mass-flow rate {dot over (m)}EH. Each set of component evaporators may consist of only two component evaporators, namely one for each bank of cylinders. Alternatively, the cylinder-block refrigerant passages, and/or the cylinder-head refrigerant passages, of each bank of cylinders may be compartmentalized, and thus the refrigerant passages of each cylinder block and/or each cylinder head may form several component evaporators. In the example considered in this major paragraph, all cylinder-block component evaporators are supplied, at a given instant of time, with liquid refrigerant at essentially the same mass-flow rate, and all cylinder-head component evaporators are also supplied, at any given instant of time, with liquid refrigerant at essentially the same mass-flow rate. Also, in the example considered in this major paragraph, excess overfeed is used only for the cylinder-head component evaporators.
Suitable transition rules between modes 3Cu and 3Eu include, in the case of the specific example being considered, rules which are in essence based on the current value of {dot over (Q)}CH where {dot over (Q)}CH is the total coolant load of all the cylinder-head component evaporators; namely, for instance,
where {dot over (Q)}CH1 and {dot over (Q)}CH2 are preselected values of {dot over (Q)}CH and where {dot over (Q)}CH2<{dot over (Q)}CH1. Typical measures of {dot over (O)}CH include (1) the steady-state cylinder-head cooling load ({dot over (Q)}C,SS)H of all the cylinder-head component evaporators, which is computed by the CCU of a system of the invention in a way similar to that used in computing {dot over (Q)}C,SS (see immediately-preceding major paragraph); and (2) {dot over (m)}VH, where {dot over (m)}VH is the total refrigerant-vapor mass-flow rate exiting all cylinder-head component evaporators, where the current value of {dot over (m)}VH can be derived by the CCU of a system of the invention from one or more refrigerant-vapor flow-rate transducers. For example, in the case of the airtight configuration shown in
In the case where a measure of the current value of TW is supplied to the CCU of a system of the invention, a typical control technique in mode 3Eu is (1) to control one or more appropriate controllable elements of the system's principal configuration so that {dot over (m)}EH tends to {dot over (m)}EH,MAX, where {dot over (m)}EH,MAX is the design maximum value of the liquid-refrigerant mass-flow rate {dot over (m)}EH supplied to all the one or more cylinder-head component evaporators; and (2) to control one or more appropriate controllable elements of the system's supplementary configuration so that the current value of TW tends to TWD. In the last-cited typical control technique, the preselected value TWD of TW would be fixed where the purpose of excess overfeed is to reduce the size of the airtight configuration's condenser; and would decrease with increased cooling load where the purpose of excess overfeed is to increase volumetric efficiency at high cooling load. The current value of the cooling load {dot over (Q)}C can be estimated by the CCU from preselected characterizing parameters. Examples of techniques for obtaining an estimate of the current value of {dot over (O)}C were given earlier in this section V,H,5,b.
c. Evaporator Liquid-refrigerant Injection
i. Preliminary Remarks
I mentioned in the second major paragraph of section V,F,2,c the use of nozzles to increase the velocity with which liquid refrigerant is supplied to an NP evaporator, and I have referred to those nozzles as liquid-refrigerant injection nozzles, or more briefly as LR injection nozzles.
I shall hereinafter, in this DESCRIPTION and in the CLAIMS, use the term ‘evaporator liquid-refrigerant injector’, or more briefly in this DESCRIPTION the term ‘LR injector’, to denote a device which supplies liquid refrigerant to an NP evaporator or to a mixed evaporator (see section V,H,7) through one or more orifices whose total cross-sectional area is smaller than the cross-sectional area of the inlet through which liquid refrigerant is supplied to the LR injector. The orifices of an LR injector may be merely apertures in the injector's walls, or may be the outlets of nozzles supplied with liquid refrigerant through those apertures. LR injectors can have walls of any shape; and may, in particular, have cross-sectional areas bounded only by a single external perimeter, or may have cross-sectional areas bounded by both an external and an internal perimeter. An example of an LR injector whose cross-sectional area normal to its axis is bounded by two perimeters is an injector whose cross-sectional area is an annulus between two concentric circles. I shall hereinafter refer to LR injectors supplied with refrigerant by a liquid-refrigerant header which is in essence parallel to an engine's crankshaft (axis) as ‘transverse LR injectors’ and to LR injectors supplied with refrigerant by a liquid-refrigerant header normal to an engine's crankshaft as ‘longitudinal LR injectors’.
I distinguish between ‘liquid-refrigerant local injectors’, or more briefly ‘LR local injectors’ or just ‘local injectors’, and ‘liquid-refrigerant distribution injectors’, or more briefly ‘LR distribution injectors’ or just ‘distribution injectors’. The local injectors have one orifice, or have several orifices, close to each other, say within one or two millimeters of each other. By contrast, the distribution injectors have several orifices distributed on the injectors' one or more surfaces over an area having at least one dimension which is a significant fraction of at least one of the, dimensions of the one or more refrigerant-passage internal surfaces of the unitary evaporator, or of the split-evaporator component evaporator, in which they are located. For example, in the case of an LR distribution injector located in the cylinder-head coolant passages of a small engine (say an engine with a displacement up to 10 liters), at least one dimension of a distribution injector is typically larger than ten millimeters; and in the case of an LR distribution injector located in the cylinder-head coolant passages of a large engine (say an engine with a displacement over 100 liters), at least one dimension of the distribution injector is typically larger than 25 millimeters. I also distinguish between (1) an LR local injector I name a ‘region-injection injector’, used primarily to inject liquid refrigerant in a localized region inside the refrigerant passages of the evaporator in which the region-injection injector is located, and (2) an LR local injector I name a ‘surface-injection injector’, used primarily to inject liquid refrigerant on, and to wet, a localized area of the internal surface of the one or more refrigerant passages of the evaporator in which the surface-injection injector is located. I further distinguish between (1) an LR distribution injector I name a ‘region-distribution injector’, used primarily to distribute liquid refrigerant over one or more regions inside the refrigerant passages of the evaporator in which the region-distribution injector is located; and (2) an LR distribution injector I name a ‘surface-distribution injector’, used primarily to distribute liquid refrigerant over, and to wet, one or more extended areas of the internal surface of the refrigerant passages of the evaporator in which the surface-distribution injector is located.
A surface-injection injector and a surface-distribution injector can be used merely to prevent the surface wetted by them becoming a hot spot by ensuring the film heat-transfer coefficient of that surface is approximately equal to the film heat-transfer coefficient it would have if it were immersed in liquid refrigerant where pool boiling prevails. Alternatively, a surface-injection injector, or a surface-distribution injector, may be used for ‘evaporative spray cooling’, or more briefly ‘spray cooling’, over a specified internal-surface area of an evaporator's refrigerant passages. In the case of a surface-distribution injector, the specified area may be only a small fraction of the internal-surface area of the evaporator's refrigerant passages over which the surface-distribution injector distributes liquid refrigerant; or the specified area may be equal to that internal-surface area.
I have used the term ‘evaporative spray cooling’ to denote techniques of liquid-refrigerant injection which achieve much higher heat-transfer coefficients than those achievable with pool boiling. Evaporative spray cooling, in the sense just defined, is discussed in a paper by Donald E. Tilton, J. H. Ambrose, and Louis C. Chow, ‘Closed-System, High-Flux Evaporative Spray Cooling’, 1989, SAE Technical Series 892316. The just-cited paper describes evaporative spray-cooling tests, with water at 100° C., in which the heat-transfer coefficients achieved were typically 1 Mw/M2 with (TW−TRS) equal to about 6° C., and typically equal to 8 to 10 Mw/m2 with (TW−TRS) equal to about 30° C. These results were obtained with orifices having a diameter between 0.51 mm and 0.76 mm; pressure differentials across the orifices of 1.4 to 7.5 bar; and distances of 1 cm, or of 1.5 cm, between the orifices and a test-surface area of about 1 cm2 at right angles to the axis of those orifices. Spray cooling can also be used to displace refrigerant vapor at an evaporator-wall location which tends to trap refrigerant vapor and is blanketed by it, thereby causing hot spots.
ii. LR Distribution Injectors
The design, location, and number, of LR distribution injectors used for cooling (the walls of) the refrigerant passages of unitary NP evaporators, or of component evaporators of NP evaporators, depend on the particular device being cooled by the airtight configuration to which the unitary evaporator or the component evaporators belong; and often also depend on the part of the device being cooled by the airtight configuration. For example, in the case of a piston engine, the design, location, and number, of LR distribution injectors will depend not only on whether the engine is a spark-ignition engine, a direct-injection compression-ignition engine, or an indirect-injection compression-ignition engine; but will also depend on the detailed design of the particular part of each of the three general types of engine just cited; and on whether, in the case of surface-distribution injectors, spray cooling is to be achieved in addition to surface distribution. I next discuss five examples of LR distribution injectors.
The first example uses a set of one or more region-distribution injectors merely to distribute liquid refrigerant around a piston-engine's cylinder liners. The set of one or more region-distribution injectors may be located at the crankcase end of the cylinder liners and have orifices through which exiting liquid-refrigerant jets point toward the cylinder head; or may be located at the cylinder-head end of the cylinder liners and have orifices through which exiting liquid-refrigerant jets point toward the crankcase. In the former case the cylinder-block liquid-refrigerant inlet 2′ will be located at the crankcase end of the cylinder liners and, in the latter case, inlet 2′ will be located at the cylinder-head end of the cylinder liners. In either case, inlet 2′ will have no fewer ports than the number of distribution-injector subsets not fluidly interconnected.
A plan view of the set of one or more region-distribution injectors mentioned in the immediately-preceding minor paragraph, in the case where a piston engine has two cylinders, and in the case where the set of one or more region-distribution injectors has only a single injector, is shown (1) in
In
The plan view, corresponding to the plan view shown in
The invention includes the case where longitudinal ribs are used in the annular space between the cylinder liners and the cylinder-block outer perimeter to keep refrigerant vapor distributed evenly around the cylinder-liner perimeters even when the angle φ is not zero degrees. The last-cited ribs can be made of thermal y-conducting material in thermal contact with the liners, thereby also acting as fins used to increase the rate at which heat is transferred from the liners to the refrigerant in refrigerant passages 504.
The second example, see FIG. 68, shows two cross-sections, in the same plane, of a set of surface-distribution injectors used to spray-cool housing 720 of valve stem 721 of exhaust valve 722 of a large piston-engine. The set of injectors could in principle consist of a single injector with a continuously-changing cross-section around the axis of valve guide 723. The set of injectors have on the left-hand side of valve stem 721 a cross-sectional area designated by numeral 724, and on the right-hand side of valve stem 721 a cross-sectional area designated by numeral 725. In the case where several injectors are used, they would be fluidly interconnected so that non-evaporated liquid refrigerant exiting injector orifices 726 exits at 727 (only a few orifices are designated by numeral 726). Liquid refrigerant enters the set of distribution injectors at 2V and refrigerant vapor, generated by liquid refrigerant after exiting orifices 726, exits at 3V.
The third example uses a set of surface-distribution injectors which form an annulus inside the cylinder-block coolant passages near the cylinder head of a large piston engine.
The fourth and fifth examples use a set of one or more surface-distribution injectors to spray-cool the critical areas of the cylinder-head coolant passages of a piston engine, the remaining areas of the cylinder-head coolant passages being cooled by wet refrigerant vapor generated by jets, exiting the injectors' orifices, when they impinge on those critical areas. The location and orientation of surface-distribution injectors for the purpose just cited can be discussed only in the context of a specific cylinder-head design. In the particular case where the surface-distribution injectors have cylindrical cross-sections with a straight axis, their axes could be in one or more perpendicular, parallel, or oblique, planes with respect to the axes of a bank of cylinders.
In the fourth example, the engine is a spark-ignition engine with two cylinders, two overhead camshafts (not shown), and four valves per cylinder; and the surface-distribution injectors are essentially horizontal and at right angles to the engine's crankshaft (not shown).
In the fifth example, see
iii. LR Pulsed Injection
A set of LR injectors, and particularly a set of (LR) surface-distribution injectors, used for cooling continuously the (walls of the) refrigerant passages of NP evaporators in general, and of NP evaporators used to cool piston engines in particular, often requires a much larger liquid-refrigerant mass-flow rate than that required for correct evaporator overfeed. (Correct overfeed in some applications may be zero.) The last non-parenthetical statement is especially true in the case of surface-distribution injectors used for spray cooling. The term ‘cooling continuously’, employed in that statement, is used to denote that the flow rate of (liquid-refrigerant) jets exiting the orifices of a set of LR injectors is continuous. I shall refer to the process of continuously cooling the (walls of the) refrigerant passages of an NP evaporator with LR injectors as ‘liquid-refrigerant continuous injection’, or more briefly ‘LR continuous injection’. A set of LR injectors may be the set of one or more injectors inside a unitary evaporator, or inside a set of one or more component evaporators of a split evaporator.
LR continuous injection is often impracticable because it often requires unacceptably-large EO or DR pumps, and an unacceptably-large separating device. I have therefore devised techniques for implementing ‘liquid-refrigerant pulsed injection’, or more briefly ‘LR pulsed injection’, which relies on the thermal capacity of the refrigerant-passage walls of a unitary evaporator, or of a component evaporator of a split evaporator, to prevent the temperature of those walls differing during LR-injector jet pulses and LR-injector jet interpulse periods by an unacceptable amount.
LR pulsed injection with a pulse-train duty ratio of 0.1 should be practicable in most applications, and in particular in most piston-engine cooling applications; and a pulse-train duty ratio as small as 0.01 should be practicable in several applications. A duty ratio of 0.1, in the case of a small piston engine with a maximum speed of 100 revolutions per second, could for example be achieved at that speed with a pulse train having a pulse duration of 10 milliseconds and an interpulse duration of 90 milliseconds. Such a pulse train would require the parts of the engine cooled by jets with that pulse train to have a thermal capacity large enough for the changes in engine-wall temperature during the pulse period (100 milliseconds) to be small enough (say ±3° C.), at the highest heat-flux value, to be acceptable.
I note that in certain applications it may be desirable to use in the same evaporator, or in the same component evaporator, both LR continuous and LR pulsed injection. An example where both continuous and pulsed injection may be desirable is a cylinder-head component evaporator. For instance, pulsed injection, with a given interpulse period, may be acceptable for cooling the component evaporator's one or more cylinder-head combustion-chamber walls, but may not be acceptable for cooling other component-evaporator refrigerant-passage walls, such as the guides of the stems of exhaust-gas valves, because the temperature change, with that interpulse period, may be unacceptably high at the one or more refrigerant-side surfaces of those other walls.
I also note that in certain applications it might be desirable, practical, and affordable, to have different pulse trains for different cylinders of the same engine. The pulses of a pulse train, for each cylinder, would in the last-cited case be controlled to coincide approximately with the highest heat-flux periods at the gas-side surface of the combustion chamber of each cylinder. The information for synchronizing evaporator LR injection pulses with those highest heat-flux periods is available from an engine's management system.
To illustrate the advantages of LR pulsed injection I use, for specificity only, the example where (1) the walls of the refrigerant passages to be cooled are the cylinder-head coolant passages of a four-cylinder piston engine having a maximum total cooling load of 46.5 kw; (2) the maximum cooling load of the cylinder-head coolant passages is 0.7 of the total cooling load; (3) the higher heat-flux regions of the cylinder-head coolant passages are to be spray-cooled by surface-distribution injectors; (4) the refrigerant used as the engine's coolant is a 50% aqueous ethylene glycol solution; and (5) the refrigerant's pressure, in the cylinder-head coolant passages, is 1.013 bar at the maximum cooling load. The total condensate volumetric-flow rate in the example just given is typically 0.022 liters/sec and the corresponding cylinder-head liquid-refrigerant volumetric-flow rate is about 0.015 liters/sec.
I assume the quality QEV,H of the refrigerant vapor exiting the cylinder-head evaporator or component evaporators must not exceed 0.2 to ensure the non-sprayed parts of the evaporator walls do not become hot spots. To achieve a quality qEV,H of 0.2, the overfeed ratio rEOH of those evaporators must be 4 which corresponds to a coolant-flow rate {dot over (m)}EH of 0.075 liters/sec. Each of the orifices used in the SAE paper cited earlier in this section V,H,5,c consume between 4 and 6.6 gph, namely between 0.0042 and 0.0069 liters/sec. It follows that the total number of those orifices in the cylinder-head evaporators, with an overfeed ratio of 4 and LR continuous injection, ranges between 18 (≈0.075÷0.0042) and 11 (≈0.075÷0.0069) orifices per cylinder head, namely is typically equal to 3 or 4 orifices per cylinder which is obviously too small a number to spray-cool all the surfaces subjected to high heat fluxes. Whereas the number of those orifices per cylinder, with an overfeed ratio of 4 and LR pulsed injection with a liquid-refrigerant duty ratio of 0.1, would typically be equal to 35. I note that an overfeed ratio of 4 would still only require a DR pump about one-twentieth the capacity of the circulation pump of a single-phase engine-cooling system with the same cooling capacity. I also note that a duty ratio of less than 0.1 should often be achievable.
The advantages of LR pulsed injection compared with LR continuous injection are not limited to smaller EO or DR pumps. The advantages of the former type of injection compared to the latter type of injection allows the use of a smaller separator, and also a lighter, less complex, and less expensive, separator. The reasons for the statement made in the immediately-preceding sentence are given next.
The value of qEV,H required to prevent hot spots occurring at a particular evaporator refrigerant-passage location, decreases—where spray cooling is not used—as the heat flux at that location increases. Usually, the heat fluxes at the internal surfaces of evaporator cylinder-head coolant passages vary, at the maximum cooling load, between a lower limit of between 0.2 and 0.3 Mw/m2 and an upper limit of between 0.7 and 1.0 Mw/m2.
Assume, for illustrative purposes only, that, with no spray cooling, locations (of those internal surfaces) with a heat flux of 0.4 Mw/m2 require a value of qEV,H not exceeding 0.2 for them not to become hot spots, and that locations with the maximum heat flux, say 0.75 Mw/m2, require a value of qEV,H not exceeding 0.05 for them not to become heat spots. And now assume that locations with a heat flux of over 0.4 Mw/m2 are spray-cooled. It follows that spray cooling, with the assumptions made, increases the maximum permissible value of qEV,H from 0.05 to 0.2 thereby greatly reducing the size, complexity, and cost of a separator required to deliver, at the maximum cooling load, refrigerant vapor of a given quality (say a quality of 0.98).
6. Combinations with Overflow P Evaporator
I choose as an example, see
Engine 500 is an in-line engine which I assume, for specificity only, has 4 cylinders. The location in elevation of refrigerant inlet 82″ (which may have one or more ports) assumes that (1) refrigerant passages 504 and 505 are fluidly interconnected, and that (2) refrigerant passages 504 are sized and configured to allow sewer flow to occur.
Subcooler 51h is a part of a heating and cooling unit (not shown) which has one or more dampers for isolating—in known ways—subcooler 51h from the cabin to which it supplies heat, and for preventing—whenever desired—ram air, or airflow induced by the heating and cooling unit's blower, flowing past the refrigerant passages (not shown) of subcooler 51h. (The heating and cooling unit may also include means which control the flow induced by that blower so that subcooler 51h rejects heat to the ambient air. See, for example, U.S. Pat. No. 5,036,803 for the case where the engine-cooling system is a single-phase heat-transfer system.)
Engine 500 drives DR pump 46 and LT pump 404B. Whether liquid refrigerant flows from reservoir 401 toward port 407 or vice-versa depends on the size of the aperture of proportional bidirectional (two-way) LT valve 435 which is used in part as a recirculation-control valve for pump 404B. When valve 435 is fully open, the entire refrigerant configuration is at ambient atmospheric pressure minus the relatively insignificant pressure resulting from the force exerted by corrugated wall 403. Valve 435 is controlled in mode 2 so that pR tends to pRD, and in mode 3 so that the level LRD of liquid-vapor interface surface 647 stays close to a preselected value LRD,D. This can be achieved by using (1) a single proportional liquid-level transducer 113, as shown in
A system of the invention having the airtight configuration shown in
In applications where engine 500 in
Where engine 500 in
Valve 752 is controlled by signal C′RSV2 so that when the transverse tilt θ2 of engine 500 becomes greater than a first preselected value, valve 752 connects port 753 to port 755; and so that, when the transverse tilt of engine 500 becomes smaller than a second preselected value less than the first preselected value, valve 752 connects port 754 to port 755. A measure of the current value of θ2 can be obtained from signal θ′2 generated, for example, by inclinometer 549 shown in FIG. 43K. The level LP of surface 123 shown in
Where desirable, manifold segment 94′″-753 can also be extended into refrigerant passages 505, and subcooler 51h can, like subcooler 51 in
Where refrigerant passages 504 are not suitable for sewer flow, evaporator refrigerant inlet 82″ in
Where such a by-pass is used, see refrigerant-circuit segment 83′-757 in
Cylinder-block component evaporators may be either overflow component evaporators having cylinder-block liquid-refrigerant overflow outlet 94′, connected at point 759 to overflow refrigerant-circuit segment 94″-96-750, as shown in
The extensions of overflow manifolds into refrigerant passages 505 may, see
An overflow P evaporator can also obviously be used with an IG auxiliary configuration instead of an ancillary configuration.
Where an IG configuration, instead of an ancillary configuration, is used with an overflow P evaporator, a mixing-control mode may be required. In this case an electric motor can be used to drive pump 46 in
In mode 2*, whenever p*R falls below p*RD, liquid refrigerant exits the refrigerant principal configuration at port 470B and enters the principal configuration at port 470H through bubble-lift R&IG-circuit segment 470B-470-470H supplied with inert gas at inert-gas inlet 470. (Inert gas exits the refrigerant principal configuration at outlet 471.) Inlet 470 is located high enough above the bottom of the U-tube shown in
When the engine stops running, a software clock starts running for a preselected first time interval during which (normally-open) valve 485 is controlled by signal C′LTV1 so that p*R tends to p*RDo, where p*RDo represents a range of preselected acceptable values which may be fixed or which may be a function of the ambient temperature TA. When the clock stops running and the current value of TR falls below TR,MIN, valve 485 is opened, the system's CCU is de-energized, and the system's control mode changes to mode 1*0A.
7. Mixed Evaporators
a. Preliminary Remarks
In the case of piston engines, with intake ports and/or exhaust ports above the engines' combustion chambers, and with twin overhead camshafts, I shall distinguish between the lower deck and the upper deck of the engines' one or more cylinder heads. I make the last-cited distinction not only for four-stroke engines, but also for two-stroke engines having such ports and camshafts. An example of a two-stroke engine with either the intake ports or the exhaust ports above the engine's combustion chambers, and with twin overhead camshafts, is a uniflow scavenging two-stroke engine. (See for example Gordon P. Blair, ‘Two-Stroke Engines’, 1990, Society of Automotive Engineers, see page 14, FIG. 1.5.) I use the term ‘cylinder-head lower deck’, or more briefly ‘lower deck’, to denote the part of the cylinder head between the cylinder-head combustion-chamber wall and the bottom of the intake and/or exhaust-valve springs; and the term ‘cylinder-head upper deck’, or more briefly ‘upper deck’, to denote the part of the cylinder head above the bottom of the intake and/or exhaust-valve springs. The lower deck of a cylinder head, as defined herein, includes the intake ports where the intake ports are located in the engine's cylinder head, and/or includes the exhaust ports where the exhaust ports are located in the engine's cylinder head.
The P evaporators in general, and the overflow P evaporators in particular, described thus far in this DESCRIPTION usually require, in the case of several types of piston engines, a higher cylinder head than the cylinder head of an engine (of the same type) employing single-phase cooling. This is particularly true for engines with twin overhead camshafts. Whether and by how much the height of the one or more cylinder heads of the last-cited engines is greater where a cooling system of the invention with a P evaporator is used, instead of a single-phase cooling system, depends—for a given engine displacement—(1) on how large a portion of the external surfaces of the exhaust-valve stems and ports must be kept immersed in liquid refrigerant while the engines' one or more cylinder heads are hot; and (2) on whether the cylinder heads' upper deck can accomodate refrigerant-vapor outlet ports. I next elaborate on the statement made in the immediately-preceding sentence using as an example an in-line engine with twin overhead camshafts and cross-flow intake and exhaust ports.
b. Description of Mixed Evaporators
One of the principal purposes of systems of the invention for cooling a piston engine of a vehicle is for those systems not to require the sizes of the cylinder-block and cylinder-head castings of the engine to be larger than the sizes of those castings if the engine were cooled by a single-phase cooling system. Whereas the last-cited purpose is usually achievable by systems of the invention having NP evaporators with LR injectors, it may often not be achievable by systems of the invention having cylinder-head component P evaporators even in the case of in-line engines subjected to small tilts. However, for certain applications an NP evaporator with LR injectors may be less cost effective than a third kind of evaporator I name ‘mixed evaporator’, or more briefly ‘M evaporator’, which combines certain features of P evaporators and NP evaporators. The applications for which M evaporators may be more cost effective than NP evaporators include cylinder-head component evaporators; and, in general, evaporators where (1) a high proportion of the internal surface of their refrigerant passages is subjected to heat fluxes high enough over a large-enough area to require the evaporator to have several surface-distribution injectors, and where (2) a substantial proportion of that area is located near the bottom of their refrigerant passages. The reason for M evaporators being sometimes more cost effective than NP evaporators, under the conditions recited in the immediately-preceding sentence, is that immersing certain high heat-flux surfaces in liquid refrigerant may be less expensive than using surface-distribution injectors to direct liquid-refrigerant jets onto those surfaces.
M evaporators are by definition ‘evaporators which cool the walls of their refrigerant passages subjected to high heat fluxes in part by immersing those walls in liquid refrigerant and in part by liquid-refrigerant jets exiting LR injectors'. The refrigerant-passage walls of M evaporators subjected to low heat fluxes are cooled by refrigerant vapor which is usually wet. The boundary between high and low heat fluxes at evaporator-wall internal surfaces depends on many factors, including the kind of refrigerant used, the refrigerant's pressure, and the shape of an evaporator's refrigerant passages. But usually surfaces subjected to heat not exceeding 0.25 Mw/m2 can be cooled by refrigerant vapor with reasonable velocities and vapor qualities provided those surfaces include no vapor-trapping locations; and surfaces subjected to heat fluxes exceeding 1 Mw/m2 cannot usually be cooled by refrigerant vapor with reasonable velocities and qualities, particularly where those surfaces include vapor-trapping locations.
Liquid-refrigerant injection by the LR injectors of an M evaporator may be continuous or pulsed, and the LR injectors may be local injectors or LR distribution injectors. Also the LR injectors of an M evaporator can, like the LR injectors of an NP evaporator, be longitudinal injectors or transverse injectors.
In a mixed evaporator the area of surface 123 may, as in
M evaporators, like NP evaporators, can have transverse injectors or longitudinal injectors, with cross-sections having any shape, and moreover the shape of the cross-section of a particular injector may change as a function of its location along the injector's axis. Also, M evaporators, like P evaporators, can be overflow evaporators or non-overflow evaporators, where the term ‘non-overflow evaporator’ refers to an evaporator whose liquid-vapor interface-surface level LP is determined by a transducer which provides a measure of that level and where CR pump 10 is controlled so that the current value of LP tends to, or stays close to, a desired preselected value. I note however that, whereas the value of LP in an M evaporator with no weirs is determined by the height of the ports of liquid-refrigerant overflow outlet 94, the value of LP in an M evaporator with weirs is usually determined by the height of those weirs.
A cylinder-head non-overflow M evaporator with no interconnecting ports 538 must be supplied with a drain line for returning excess liquid refrigerant in the evaporator to the refrigerant-principal-circuit segment downstream from the refrigerant passages of the unitary condenser, or of a component condenser of the split condenser, used in the same principal configuration as the evaporator. The drain line, in the case of the M evaporator shown in
8. Remote Control of Liquid-refrigerant Pulsed Injection
Each of the injectors of LR-injector sets 531″a and 531″b in
Condenser 508h is part of a cabin-heating and cooling unit (not shown) which has one or more dampers for isolating—in known ways—condenser 508h from the cabin to which it supplies heat, and for preventing—whenever desired—ram air, or airflow induced by the heating and cooling unit's blower, flowing past the refrigerant passages (not shown) of condenser 508h.
Liquid refrigerant generated in condenser refrigerant passages 399 of condenser 508, liquid refrigerant generated in the refrigerant passages of condenser 508h, and non-evaporated liquid refrigerant exiting component separators 42*a and 42*b respectively at 45*a and 45*b, is returned by gravity to condenser liquid header 509. This, in the case of a group H refrigerant, helps ensure the concentration of the refrigerant's component with the higher freezing temperature in header 509 is high enough for liquid refrigerant, trapped in header 509 while the principal configuration shown in
In addition to liquid-refrigerant return paths 6h-808a-808b-805-806-807, 45*a-808a-808b-805-806-807, and 45*b-808b-805-806-807; drain lines 645a-809a and 645a-809a and 645b809b are used to ensure only a minimal amount of liquid refrigerant is trapped in refrigerant passages 504a and 504b when the principal configuration shown in
Refrigerant and inert-gas line 6-810 is a line with a large-enough cross-sectional area (1) to allow liquid refrigerant to be transferred from condenser refrigerant outlet 6 to dual-return receiver liquid-refrigerant inlet 810, and (2) to allow inert gas to be transferred from outlet 6 to inlet 810 and from inlet 810 to outlet 6.
DR pump 46 includes pulley-and-clutch 621 for driving the shaft of pump 46 by engine 500; and electric motor 814 for driving the shaft of pump 46 through electric-motor pulley 815 and belt 816. The clutch of pulley-and-clutch 621 is normally not engaged, and is engaged only while motor 814 drives the shaft of pump 46. (Driving the shaft of motor 814 by belt 816 while engine 500 is running is usually acceptable, and therefore usually no additional clutch is needed to isolate the shaft of electric motor 814 while engine 500 is driving the shaft of pump 46.) DR pump 46 supplies pressure regulator 817 with liquid refrigerant at inlet 818. Excess liquid refrigerant supplied to regulator 817 exits at 819 and is returned to dual-return receiver 640 at a second liquid-refrigerant inlet designated by numeral 811. Liquid refrigerant, supplied to refrigerant-control valves 801B and 801H at respectively 802B and 802H, exits regulator 817 at outlet 820 at a pressure pj whose current value is maintained, by pressure regulator 817, above the current value of the refrigerant pressure at inlet 818 by a desired preselected amount (ΔJRp)D. The value of (ΔJRp )D is usually fixed. However, the invention includes using a pressure regulator which is controlled (see
Buffer 821 is used to store liquid during interpulse periods in variable-volume jet liquid-storage reservoir 822, and spring 823 (of buffer 821) is used to ensure liquid refrigerant is supplied (during jet pulses) to injectors 800′a, 800′b, 800″a, and 800″b, at a pressure close to (pR+ΔJRp), with the assistance of pressure-equalization line 849-850. Liquid refrigerant enters and exits reservoir 822 through inlet-outlet 824.
A system of the invention, having the R&IG configuration shown in
In mode 0*0A, no system-controlled elements are controlled. In mode *0B, (1) valves 485 and 486 are controlled by signals C′GTV1 and C′GTV2 so that p*R tends to p*RDo in for example the way described in the second minor paragraph of the seventh major paragraph of section V,H,6; (2) fan 510 does not run; (3) valve 801B is closed; and (4) valve 801H is controlled by signal C′IH so that the current value of TW rises as a preselected rate as a function of the current value of TW.
In mode 1*B (mode 1*A is not used), (1) valves 485 and 486 are controlled so that p*R tends to pRDo*; (2)fan 510 runs; (3) valve 801B is closed; and (4) valve 801H is controlled by signal C′IH so that the liquid-refrigerant (mean) flow-rate delivered by it is almost equal to the predetermined flow rate at which pump 46 can induce liquid-refrigerant flow while it is driven by electric motor 814.
In mode 2*, (1) valves 485 and 486 are controlled by signals C′GTV1 and C′GTV2 so that TW tends to TWD in for example the way described in the last-cited minor paragraph of section V,H,6, for making p*R tend to p*RD; (2) fan 510 does not run; and (3) valves 801B and 801H are controlled by signals C′IH and C′IB in one of the ways described in section V,H,5,b for maintaining the current value of respectively the overfeed ratios rEO,B and rEO,H close to their desired preselected values. I note that, because of interconnecting ports 538a and 538b, the value of rEO,B affects the value of rEO,H, but this should usually be only a second-order effect. If no ports 538a and 538b existed and refrigerant vapor outlets 3′a and 3′b were used (as for example in FIG. 63C), the values of rEO,Ha and rEO,Hb would be unaffected by the values of rEO,Ba and rEO,Bb, where rEO,Ha and rEO,Hb are the overfeed ratios of the cylinder-head component evaporators, and where rEO,Ba and rEO,Bb are the overfeed ratios of the cylinder-block component evaporators.
In mode 3*, (1) valves 485 and 486 are control led by signals C′GTV1 and C′GTV2 so that p*GR stays close to P*GR,MAX in for example the way described in the last-cited minor paragraph of section V,H,6; (2) fan 510 is control led by signal C′CF so that TW tends to TWD; and (3) valves 801B and 801H are controlled in the same way as in mode 2*.
Typical transitions are those recited in section V,G,2,b,iv (less the transition rules between mode 1*A and modes 0*0A, 0*0B, 1*B, 2*, and 3*).
The invention includes, see
The invention also includes using, see
The invention further includes adding, as shown in
A perusal of the subgroup IIFF principal configuration shown in
9. Separating Devices and Oil Heaters and Coolers
The location of a separating device depends, in the case of a piston engine, (1) on the location of the evaporator refrigerant-vapor outlet ports, which in turn depend on the type of piston-engine being cooled; (2) on the orientation of the engine with respect to the condenser, particularly where the condenser is an air-cooled condenser; and (3) on the location and shape of the available space for the separating device in the engine compartment. In the case where the engine has several banks of cylinders, each bank of cylinders may have its own component separating device which may be located at the side, at the end, or at the top, of a bank of cylinders. The first of the last-cited three locations is usually preferred with engines having—like most passenger-car engines envisioned by me—transverse refrigerant-vapor outlet ports. The second of the last-cited three locations is usually preferred only with certain engines, such as perhaps engines with a single overhead camshaft and uni-sided intake and exhaust ports, where a longitudinal vapor header is practicable. The third of the last-cited three locations is preferred with few engines and is unacceptable with any engine where, as in most passenger cars with in-line engines, no room is available above a bank of cylinders. (In the case of an engine having twin overhead camshafts, refrigerant vapor could be transferred to a separating device by narrow rectangular ducts between the two camshafts and between, as applicable, an engine's spark plugs or fuel injectors.)
Separating devices can have any shape and can use any known means for separating the liquid phase of a fluid from its vapor phase; and, in particular, any known means used in the steam-generating and refrigeration industries to accomplish the last-cited purpose.
I shall describe separating devices by using as examples separating assemblies. (Many separators can be derived from the separating assemblies described in this section V,H,9 merely by combining a separating assembly with a vessel, located below the assembly and fluidly interconnected with it, into a single unit.) I choose as examples of separating assemblies shapes which are unusual in the steam-heating and refrigeration industries, but which may be appropriate where (1) the engine has transverse refrigerant-vapor outlet ports, and where (2) the space available for a separating device is long—albeit possibly segmented in part—in a direction parallel to an engine's crankshaft (axis), and is short in a direction normal to the plane containing the engine's cylinder-bore axes.
The invention includes, where desirable, means for heating an engine's (lubricating) oil with the refrigerant of an airtight configuration used to cool the engine; and in particular, means for heating the engine's oil with the refrigerant's vapor. An inexpensive way of doing this, in the case where a separating device having a separating assembly similar to that shown in
The invention also includes, where desirable and practicable, means for cooling an engine's (lubricating) oil with the refrigerant of an airtight configuration; and, in particular, for cooling the engine's oil with the refrigerant vapor of an airtight configuration. An inexpensive way of doing this, in the particular case where a separating device having a separating assembly similar to that shown in
The invention further includes means for heating and cooling an engine's (lubricating) oil with the refrigerant of an airtight configuration by using the selfsame heat exchanger.
Under certain operating conditions the current value of the quality qEV of the refrigerant vapor entering a separating assembly with a heat exchanger used to cool engine oil, or any other fluid, may be high enough to allow the heat exchanger to superheat refrigerant vapor exiting the separating assembly. To prevent this occurring, the invention includes means (1) for obtaining a measure of the temperature TRV of refrigerant vapor after it exits a separating device with an oil-cooling heat exchanger; (2) for obtaining a measure of the refrigerant saturated-vapor temperature TRS at a point upstream from the separating assembly; (3) for comparing the current values of TRV and TRS; and (4) for increasing, whenever the current value of TRV exceeds the current value of TRS, the rate at which liquid refrigerant is supplied to an evaporator (belonging to an airtight configuration having a separating device which includes an oil-cooling panel) above the rate at which liquid refrigerant would be supplied to the evaporator if the current value of TRV did not exceed the current value of TRS. Acceptable measures of the value of TR5 include (1) in the particular case of a P evaporator, or an M evaporator—where available—the temperature of the liquid refrigerant in the evaporator; and (2) in general the value of TRS computed from p*R in the case of type A combinations, and from p*R in type C combinations under conditions where p*R is known to provide an acceptable measure of pR. I next describe an example of the technique just outlined in b this minor paragraph in more detail using the R&IG configuration shown in
In the case of (1) the R&IG configuration shown in
I assume for specificity only that the R&IG configuration shown in
10. Special Technique for Determining Liquid Level
A special technique for determining the level of liquid refrigerant in a refrigerant-circuit segment of an airtight configuration—and, in particular, in a receiver, separator, P evaporator, or M evaporator—is often preferable to alternative techniques for determining that level; and, in particular, to techniques employing float transducers.
The special technique mentioned in the immediately-preceding minor paragraph employs a differential-pressure transducer which in effect provides a measure of the weight of the column of liquid refrigerant present in a refrigerant-circuit segment beginning at a first point, hereinafter referred to in this section V,H,10 as ‘the upper point’, above the preselected highest level of the column, and ending at a second lower point, hereinafter referred to in this section V,H,10 as ‘the lower point’, at or below the preselected lowest level of the column. The last-cited measure can be obtained by two different methods. In the first of the two methods, the transducer's low-pressure port is connected to the upper point, the transducer's high-pressure port is connected to the lower point, and the refrigerant line connecting the transducer's low-pressure port to the upper point contains only refrigerant vapor. And, in the second of the two methods, the transducer's low-pressure port is connected to the lower point, the transducer's high-pressure port is connected to the upper point, and the last-cited refrigerant line contains only liquid refrigerant. With the former method, the transducer generates a signal representing a direct measure of the weight of the liquid column whose level is to be determined. And, with the latter method, the transducer generates a signal representing a measure of the absolute value of the difference between that weight and the weight of the liquid column in the refrigerant line connecting the high-pressure port to the upper point, thereby providing an indirect measure of the weight of the liquid column whose level is to be determined. Errors in determining this level, arising from changes in liquid-refrigerant density, can be corrected by measuring refrigerant pressure with an absolute-pressure transducer and adjusting, in the CCU, the measure provided by the liquid-level transducer. Errors arising from neglecting refrigerant-vapor weight can be corrected by iteration. And errors arising from changes in refrigerant-vapor density can—like errors in liquid-refrigerant density—be corrected by measuring refrigerant pressure. In most applications envisioned for airtight configurations, none of the last-cited three corrections is necessary.
I shall hereinafter refer to a differential-pressure transducer used as a liquid-level transducer as a ‘differential-pressure liquid-level transducer’, or more briefly as a ‘PD liquid-level transducer’.
A PD liquid-level transducer using the first method described, in the immediately-preceding major paragraph, in this section V,H,10, can be employed to provide a measure of the level of any one of the many refrigerant liquid-vapor interface surfaces shown in the FIGURES of this DESCRIPTION provided (1) the transducer's low-pressure port is connected correctly to the pertinent refrigerant line at the upper point mentioned earlier in this section V,H,10; and provided (2) the refrigerant line connecting the low-pressure port to the upper point is heated sufficiently, while the principal configuration of the airtight configuration with which the transducer is associated is active, to ensure that line contains no liquid refrigerant.
Examples of the correct connection mentioned under (1) in the immediately-preceding minor paragraph are given in
A PD liquid-level transducer using the second method described earlier in this section V,H,10 can be employed to provide a measure of the level of the refrigerant liquid-vapor interface surfaces shown in the FIGURES, only where (1) the transducer's high-pressure port is connected correctly to the pertinent refrigerant line at the upper point mentioned earlier in this section V,H,10; (2) the void fraction at the first point is substantially less than unity while the principal configuration of the airtight configuration with which the transducer is associated is active; and (3) the void fraction at the upper point is zero while the principal configuration is inactive. Examples of the correct connection mentioned under (1) in this minor paragraph are given in
11. Charging Techniques for Airtight Configurations
a. Preliminary Remarks
The one or more surfaces of a component of an airtight configuration intended to be in direct contact with the configuration's refrigerant and/or inert gas should usually be cleaned before the configuration is assembled. The cleaning method used depends on the one or more materials from which the last-cited one or more surfaces are made, and on the kind of refrigerant to which they will be exposed. In the case of certain metals such as aluminum and iron the invention envisions that the processes used to clean them may include steam-cleaning.
Air should be removed from the refrigerant enclosure of a refrigerant configuration before the refrigerant configuration is charged with refrigerant. Air should also be removed from the R&IG enclosure of a type C combination where the inert gas of the type C combination is initially not air. Any applicable known techniques may be used to remove the air from the two last-cited enclosures, including removing the air from them with a vacuum pump, or flushing the air out of them with an inert gas.
b. Type A Combinations
I choose the case where a type A combination is used to cool a piston engine. However, the outline of the typical technique described next also applies to type A combinations for most other applications.
For specificity, I discuss the last-cited technique in the context of the refrigerant configuration shown in
Where air in the refrigerant configuration shown in
After a successful pressure test, (1) liquid refrigerant is inserted at 828 and inert gas exits as 826 until liquid refrigerant starts exiting at 826, (2) whilst the internal volume of reservoir 401 is maintained at a first minimal preselected value (say equal to 10% of the reservoir's maximum internal volume), liquid refrigerant is inserted at 826 until liquid refrigerant exits at 828, and (3) engine 500 in
c. Type C Combinations with Complete Minimum-pressure Maintenance
To discuss charging techniques for type C combinations, I distinguish between the case where the inert gas used with a type C combination is air and the case where the inert gas used with a type C combination is not air. And, in the latter case, I distinguish between the case where the R&IG enclosure of a type C combination can be evacuated and the case where that enclosure cannot be evacuated.
In the case where the inert gas employed is air, I insert a preselected mass of refrigerant into the R&IG enclosure of a type C combination and allow displaced air in the R&IG enclosure to escape through an appropriately located flush valve. I then add air to, or remove air from, the R&IG enclosure until the total pressure inside the enclosure is equal to a predetermined charging value for the enclosure's current temperature. In cases where, at the ambient atmospheric temperature, the vapor pressure of the refrigerant employed is not substantially higher than the ambient atmospheric pressure, it may be desirable or even necessary either (1) to heat the refrigerant being inserted into the R&IG enclosure, or (2) to connect a vacuum pump to the last-cited flush valve and to use the pump to lower the total pressure inside the enclosure.
In the case where the inert gas employed is not air and the R&IG enclosure used can be evacuated, it is usually preferable to remove air from the R&IG configuration by evacuating it instead of flushing air out of it. In the case where an R&IG configuration is evacuated, a preselected mass of refrigerant is inserted into the R&IG enclosure after the enclosure has been evacuated, and then inert gas is added until the total pressure inside the enclosure is equal to a predetermined charging value for the enclosure's current temperature.
In the case where the inert gas employed is not air and the R&IG enclosure used cannot be evacuated, air is flushed out of the enclosure, with the inert gas to be employed, before inserting a preselected mass of refrigerant into the enclosure. Inert gas is then added to, or removed from, the R&IG enclosure until the total pressure inside the enclosure is equal to a predetermined value for the enclosure's current temperature.
In all of the foregoing three cases the R&IG enclosure is tested under pressure for leaks, with an appropriate gas, before refrigerant is inserted into the enclosure.
12. Orientation of Cylinders Cooled by Non-pool Evaporators
P evaporators and M evaporators severely limit the orientation of the cylinders of a piston engine cooled by them. This is true even where, at considerable additional cost, the level of the liquid-vapor refrigerant in each cylinder is controlled independently. (See, for example, U.S. Pat. No. 4,584,971 (Neitz et al) 29 Apr. 1986.) By contrast, NP evaporators in no way limit the orientation of those cylinders provided their refrigerant passages are configured appropriately and equipped with appropriately-located refrigerant inlet and refrigerant outlet ports.
1. Preliminary Remarks
I have so far discussed complete minimum-pressure maintenance, self regulation, and refrigerant-controlled heat release, or more briefly RC heat release, only in the context of (internal-combustion) piston-engine cooling and intercooling systems. Furthermore, I have restricted the piston-engine cooling and intercooling applications discussed to those where complete minimum-pressure maintenance and self regulation are always required, and where RC heat release is usually also required. However, from my teachings in sections V,F and V,G, it should be clear to those skilled in the art how type A, or type C, combinations can be used in piston-engine cooling and intercooling applications where only complete minimum-pressure maintenance and self regulation, or where only RC heat release and self regulation, are required.
2. Other Cooling and Intercooling Systems
It should be obvious, from the last-cited teachings, how a type A, or a type C, combination can be used to cool the stationary parts of motors, other than (internal-combustion) piston engines, such as internal-combustion rotary engines, gas turbines, and electric motors. It should also be obvious, from the last-cited teachings, how a type A, or a type C, combination can, where applicable, be used for intercooling motors other than piston-engines; for example for intercooling internal-combustion rotary engines or for intercooling gas turbines. It should further be obvious from those teachings how a type A, or a type C, combination can be used to cool electronic equipment such as computer chips, infrared arrays, and superconductors, and to cool the product of an industrial process. I therefore, in the examples given next in this section V,I,2, merely show typical interconnections between the principal configuration of an airtight configuration of the invention and several different kinds of devices other than piston engines.
An electric motor, an electric generator, a computer, or another heat-generating equipment, is sometimes located in an enclosure into which air cannot enter to cool the heat-generating equipment. In such cases, a system of the invention with an air-cooled condenser can be used to cool that equipment; and, where the equipment is installed on an automotive vehicle including an electric motor driving the vehicle, ram air generated by the vehicle's motion can be used to assist in cooling the equipment. Where the automotive vehicle is a boat or a ship, a condenser cooled by (usually treated) sea water can often be employed instead of an air-cooled condenser.
I note that the refrigerant employed depends on the temperature at which the components of circuit boards 892 are to be maintained. If those components include low-temperature superconductors, an appropriate refrigerant would be helium; if they include high-temperature superconductors, an appropriate refrigerant would be nitrogen; and if they include neither of the last-cited two superconductors, an appropriate refrigerant would often be a fluorinert coolant.
Numeral 900 designates the gas turbine's expander, numeral 901 designates the turbine's first-stage compressor, and numeral 902 designates the turbine's second-stage compressor. Air exiting compressor 902 at 903 is supplied to expander 900 at 904 after being heated by combustor 905.
The cooling system employs a liquid metal as its refrigerant; includes a CCU (not shown); and has a class IIIFN principal configuration, and a type IIR or a type IIIR ancillary configuration designated by the numeral 909. (A type IIR or a type IIIR ancillary configuration is usually preferred where a type A combination employs a liquid metal as its refrigerant.) The refrigerant passages of an NP evaporator are formed inside the stator of expander 900. The NP evaporator has a refrigerant inlet designated by numeral 2 and a refrigerant outlet designated by numeral 3. Freeze protection where required is achieved in a way similar to that described in section V, I,3,c,ii.
The intercooling system includes a CCU (not shown), intercooler air-cooled condenser 508i, intercooler type 2 separator 42i, intercooler DR pump 46i, intercooler fan 510i, intercooler fixed-volume LR reservoir 424i, and intercooler LT pump 404i. The intercooling system also includes block 906i representing an assembly which includes, for example, intercooler intake-air section 560i and intercooler evaporator 561i shown for instance in
3. Heating and Heat-recovery Systems
a. Preliminary Remarks
I shall use a heating, or a heat-recovery system, to illustrate techniques of the invention for achieving (1) partial minimum-pressure maintenance in the case of a type A or a type C combination, and (2) freeze protection and refrigerant-controlled heat absorption, or more briefly RC heat absorption, in the case of a type A combination.
Heating and heat-recovery systems differ fundamentally from cooling systems only in that, in the case of the former systems, the thermal capacity of their principal heat sink is finite; whereas, in the case of the latter systems, the thermal capacity of their principal heat sink is quasi-infinite. It follows that the airtight configurations and control techniques disclosed in sections V,F to V,H can mutatis mutandis also be used, in heating and heat-recovery applications, to achieve complete minimum-pressure maintenance and self regulation with a type A, or with a type C, combination, and RC heat release with a type A combination. It also follows that my teachings given next in sections V,I,3,b to V,I,3,e can be used to achieve, in cooling and intercooling applications, partial minimum-pressure maintenance with a type A, or with a type C, combination, and RC heat absorption with a type A combination. I shall therefore not describe (1) complete minimum-pressure maintenance, self regulation, and RC heat release, in heating and heat-recovery systems; and (2) partial minimum-pressure maintenance, freeze protection, and RC heat absorption, in cooling and intercooling systems.
b. Type A Combinations with Partial Minimum-pressure Maintenance.
i. Preliminary Remarks
Type A combinations with a partial minimum-pressure-maintenance capability are, for example, particularly cost effective where
ii. System for Generating Steam with Recovered Radiant Heat
The specific example chosen is a system—which I shall hereinafter refer to in this section V,I,3,b,ii, as ‘the system’—for recovering radiant energy and for utilizing the recovered radiant energy to generate saturated steam in the temperature range between say 145° C. and 220° C. (Examples of radiant heat are solar radiant energy, and the radiant energy emitted by steel slabs and blooms in a steel-making plant.) But the partial minimum-pressure-maintenance technique discussed next would usually be affordable with any other system having non-airtight components in only principal-configuration refrigerant-circuit segments completely filled with liquid refrigerant while the system is active and is in its self-regulation mode, and while it is inactive. A similar technique may also be affordable with a system having non-airtight components in principal-configuration refrigerant-circuits filled only partially with liquid, or even containing no liquid, while the system is inactive—provided the total internal volume of those segments is small enough for the system's LR reservoir and LT pump to be affordable.
I assume the system is installed in a heated building, and that therefore a suitable refrigerant is water. (In the case where the radiant energy is solar radiant energy, the refrigerant passages of the system's solar collector, and the refrigerant lines associated with the solar collector, would be located and sloped so that no liquid refrigerant remained in them after the system is de-activated. (See U.S. Pat. No. 4,358,929 (Molivadas), 16 Nov. 1982.)
Typical water saturated-vapor temperatures for generating steam between 145° C. and 220° C. lie, at the design maximum heat-transfer rate, in the range between 175° C. and 250° C. Refrigerant circuits using water with saturated-vapor temperatures in the range between 175° C. and 250° C. usually have steel pipes with welded-steel joints, and therefore their piping should—with a large margin of safety—be immune to air ingestion, while inactive, at ambient temperatures found inside heated buildings. (The vapor pressure of water at 10° C. exceeds 0.01 bar.) However, the foregoing circuits may include the refrigerant passages of components such as refrigerant pumps or refrigerant valves which may, as in the example discussed next, be unavailable or unaffordable where required to be airtight while the system is inactive.
In
The system's non-airtight components are DR pump 46, (liquid-refrigerant) flow-rate transducers 141 and 143, and service valves 926, 927, and 928. The refrigerant-circuit segment with the non-airtight components can be isolated, while the system is inactive, with (glandless) bidirectional liquid-isolating valve 929 and unidirectional liquid-isolating valves 930 and 931. The refrigerant principal circuit (of the principal configuration) also includes a refrigerant absolute-pressure transducer 932 which generates a signal pRis′ providing a measure of the refrigerant pressure pRis in the liquid-refrigerant circuit segment isolated by valves 929, 930, and 931, while the system is inactive. DR pump 46 is controlled as a function of the flow rates FDR and FEO obtained (by the system's CCU) from signals F′DR and F′EO, respectively, generated by flow-rate transducers 141 and 143 respectively. Techniques for controlling pump 46, as a function of FDR and FEO, so that self-regulation conditions (A) to (D) are satisfied, have already been disclosed in this DESCRIPTION. The ancillary configuration includes (glandless) refrigerant-isolating valve 933. While the system is active, valve 929 is open, and valve 933 is closed. (Valve 933 isolates LR reservoir 401 from the high refrigerant operating pressures in the principal configuration, thereby allowing a less expensive reservoir to be used.)
Cold water enters fluid passages 281 after passing through three-way cold-water valve 304 having water inlet 935 and water outlets 936 and 937. Valve 304 is used to bypass cold water around fluid passages 281. Fuel-fired steam boiler 940 is used to supplement, as required, heat supplied by the system. (Boiler 940 may be a fire-tube or a water-tube boiler for the lower part of the range of saturated-vapor temperatures given in section V,I,3,b,ii, but would be a water-tube boiler for the upper part of the range of saturated-vapor temperatures given in the last-cited section.) Techniques similar to those described in section V,Q of my co-pending U.S. patent application Ser. No. 400,738, filed 30 Aug. 1989, can for example be used to ensure boiler 940 provides the supplementary heat necessary to ensure steam is supplied at the required mass-flow rate and pressure to the utilizing equipment or process (not shown) while the system is (1) supplying no heat, (2) supplying only preheated water, or (3) supplying steam at an inadequate temperature, or at an inadequate rate. (The interconnections shown in
For specificity, I first consider the case where evaporator refrigerant passages 102a to 102f are located low enough for them to contain liquid-refrigerant at start-up. In this case, the following start-up and shut-down procedures can be used. (Line LL′ indicates the level of liquid-refrigerant in the principal configuration while it is inactive.)
When the radiant heat source is turned on, valve 929 is opened, valve 933 is closed, and pump 46 is started, as soon as the refrigerant pressure pRis exceeds pRDis by a first preselected positive amount, where pRDis is the preselected desired value of pRis while the system is inactive.
When the radiant heat source is turned off, pump 46 continues to run, valve 829 stays open, and valve 933 stays closed, while pRis stays at or above pRDis plus a second preselected positive amount smaller than the first preselected positive amount. When pRis falls below pRDis plus the second preselected positive amount, pump 46 stops running, valve 929 closes, and valve 933 opens. Thereafter, while the radiant source of heat stays turned off, air-transfer pump 420 is controlled so that (the value of) pRis tends toward pRDis
Signals F′DR, F′EO, and pRis′, generated by transducers 141, 143, and 932, respectively, are supplied to the system's CCU (not shown). And signals C′DR, C′LIV1, C′LTV3, C′AT, and C′WB, used to control pump 46, valve 929, valve 933, pump 420, and valve 304, respectively, are generated by the system's CCU.
I note that, if valves 929, 930, and 931 were leakproof, reservoir 401 would be minute because it would in essence only need to accomodate differences, in liquid refrigerant volume in the isolated principal-configuration circuit segment, caused by changes in temperature within the temperature range of interest. In practice, however, valves 929, 930, and 931 may have a slow leakage rate which would have to be offset by liquid refrigerant stored in reservoir 401, and pump 420 would have to be controlled to maintain pRis at the preselected value of pRDis.
Very similar techniques to those described in the immediately-preceding major paragraph can also be used where passages 102 contain no liquid refrigerant at start-up—provided the radiant heat-source intensity, during start-up, is low enough for passages 102 to be exposed to that intensity while they contain no liquid refrigerant. Where the condition just cited is not satisfied, additional means and control techniques are required to ensure evaporator 924 is not damaged.
c. Type A Combinations with Freeze Protection
i. Preliminary Remarks
Freeze protection, in the sense described under (a) to (e) in section III,E, can be used without heating the LR reservoir of a type A combination where the thermal equilibrium temperature of the LR reservoir with its surroundings is always high enough to prevent the combination's refrigerant freezing. This is, for example, the case where the refrigerant is water and the LR reservoir is located in a heated building. However, certain important refrigerants such as liquid metals have freezing temperatures much higher than the space inside heated buildings. Where such refrigerants are used, the LR reservoir must be heated and insulated so that it is located in a space above the freezing temperature of the refrigerant. Examples of liquid-metal refrigerants are potassium, sodium, and lithium, which have respectively freezing temperatures of 63.7° C., 97.8° C., and 179° C. Such refrigerants are collectively thermodynamically-suitable fluids for (liquid-vapor) two-phase heat-transfer systems in roughly the saturated-vapor temperature range between 600° C. and 1700° C., and are therefore thermodynamically suitable for ultra-high-temperature heat-transfer applications such as, for example, the utilization of heat of waste gases, in the range between 900° C. and 1200° C., leaving soaking pits and reheating furnaces in steel plants; the utilization of heat collected by high-gain solar collectors, which currently operate at temperatures up to 1500° C.; and the utilization of the heat of gas-turbine exhaust gases (which often exceed 600° C.).
ii. System for Running a Gas Turbine with Heat from Waste Gases
The specific freeze-protection example discussed is a system for recovering heat from the waste gases of a reheating furnace in a steelmaking plant and for utilizing the recovered heat to run a gas turbine. The heat-recovery system shown in
In
Compressed air exits, at 914, single-stage turbine compressor 915 driven by gas-turbine expander 900 and enters condenser fluid passages 281 of condenser 913. Heat released by the heat-recovery system's refrigerant in passages 399 is absorbed by compressed air flowing through passages 281. Heated compressed air leaving passages 281 is supplied to inlet 904 of expander 900 after passing through combustor 905. Whenever gas turbine 917 is required to run while furnace 910 is not operating, or while its exhaust gas is not supplying heat at a high-enough rate to run turbine 917, combustor 905 is used respectively to provide the heat required, or to supplement the heat supplied by the heat-recovery system to the turbine's compressed air. (Means for controlling a supplementary source of heat are well known and therefore not shown.) I next discuss only freeze-protection techniques.
While the principal configuration of the heat-recovery system is active LT valve 933 is open.
When the principal configuration is deactivated, the heat-recovery system's CCU (not shown) applies a signal C′LTV3 which opens valve 933, and a signal C′AT which causes air pump 420 to run until the internal volume VLR of reservoir 401 reaches its maximum value VLR,MAX. The maximum value of VLR,MAX is chosen no smaller than the largest possible volume of the heat-recovery system's liquid refrigerant charge over the range of liquid refrigerant temperatures of interest. As soon as VLR is equal to VLR,MAX, the heat-recovery system's CCU closes valve 918 to stop liquid refrigerant flowing back into the principal configuration through port 407.
Temperature transducer 919 is used to generate a signal T′LR which provides a measure of the refrigerant temperature TLR in the reservoir. The value of the temperature TLR is maintained by heating elements 920 above the refrigerant's freezing temperature. Numeral 921 designates insulation around cylinder 419. Elements 920 may be electrical heating elements, or may be passages through which flows a fluid having a higher temperature than the refrigerant's freezing temperature.
d. Type A Combinations with Refrigerant-controlled Heat Absorption
i. Preliminary Remarks
RC heat absorption is suitable for systems of the invention having a heat source whose temperature is lower than the maximum-permissible temperature of their refrigerant and of their evaporator refrigerant passages. Examples of such a heat source are (1) the coolant of an internal-combustion piston or rotary engine having a single-phase or two-phase cooling system; (2) the flue gas of a boiler; or (3) the heat-transfer fluid of a water boiler or of a steam boiler. Examples of the systems with the heat sources cited in the immediately-preceding sentence are subsystems for heating buildings and their water supplies, for heating ships and their water supplies, or for supplying heat to low-temperature industrial systems. Such subsystems would typically employ water as their refrigerant and be either (1) low-pressure subsystems operating at (absolute) pressures up to about 2 bar, or (2) subatmospheric-pressure subsystems operating at pressures up to about 0.9 bar. In the latter case, the subsystem's component condensers could have refrigerant passages formed by using the techniques described in the last minor paragraph of section V,b,15.
ii. System for Heating Compartmentalized Spaces in a Building or in a Ship
The system shown in
In
To achieve heat-absorption control (1) pump 404(A) is controlled by signal C′LT(A) so that the current value of the level LD(A) of liquid-vapor interface surface 521(A), derived from signal L′D(A), tends to value LDD(A) which may be a single preselected value, or a range of preselected values, within a preselected lower limit and a preselected upper limit; and (2) pump 46(A) is controlled by signal C′DR(A) so that the current value of the refrigerant pressure pR(A), derived from signal p′R(A), tends to a desired preselected value which varies in a pre-prescribed way as a function of one or more parameters characterizing the environment of the building, or the ship, in which the refrigerant configuration shown in
Whenever the rate at which condenser 237(A) releases heat changes because the value of pR(A) is changed, or because of the actions caused by the thermostat in a compartment of the building, or of the ship in which that thermostat is located, the amount of liquid refrigerant in the component condensers in the compartment changes thereby changing the range of the amounts of liquid refrigerant in the principal configuration for which self regulation can be achieved. The refrigerant configuration and control techniques described in this major paragraph automatically maintain the amount of liquid refrigerant in the principal configuration, within the range for which self regulation can be achieved, by changing the amount of liquid refrigerant in variable-volume LR reservoir 401(A).
e. Type C Combinations with Partial Minimum-pressure Maintenance
i. Preliminary Remarks
Many fossil-fuel-fired industrial heating systems often have their non-airtight components—such as pumps with mechanical seals, and valves and gauges with glands—located only in the vicinity of their boiler. In such cases, a type C combination, employing a refrigerant whose pressure falls below ambient atmospheric pressure while the combination's principal configuration is inactive, needs only partial, and not complete, minimum-pressure maintenance. I next discuss a specific example of a type C combination with partial minimum-pressure maintenance.
ii. System for Supplying Heat to an Industrial Process
The specific example chosen is a low saturated-vapor temperature heating system employing a fuel-fired NP evaporator and used to provide heat to a low-temperature industrial process, say an electroplating process. The refrigerant employed is water and the system may be a low-pressure system or a subatmospheric-pressure system. (In the case of an electroplating plant, the system could be a subatmospheric-pressure system.)
In
Assume the desired value p*RDis of pRis is 0.75, as might be the case in a subatmospheric-pressure system operating typically at 0.85 bar. Then sufficient inert gas must be stored in fixed-volume IG reservoir 453 to ensure the pressure p*Ris does not fall below 0.75 bar at the design minimum ambient temperature which is say 10° C. Let VGR, the internal volume of reservoir 453, be one-twentieth of the volume VGPP of the principal configuration which must be filled with inert gas to achieve partial minimum-pressure maintenance. Then the system must be charged with a sufficient mass of inert gas to allow the volume (VGR+VGPP) to be maintained at a pressure of at least 0.75 bar at 10° C. Assume VGR is required not to exceed 5% of the value of VGPP. Then, while the system's principal configuration is active and all the inert gas in the system is stored in reservoir 453, the pressure at 10° C. in the reservoir would be 0.75 bar times
namely 15.75 bar. However while the system is operating at its design maximum temperature, t e temperature in reservoir 453 will be much higher even if the ambient temperature is only 10° C. Assume the maximum temperature which might at times be reached by TGR is 80° C. Then the pressure in reservoir 453 would increase from 15.75 bar to 15.75 bar times 1.25(=353/283), namely to 19.6 bar. Consequently, to meet the foregoing 5% requirement, reservoir 453 would have to be designed so that it can withstand a maximum pressure of about 20 bar. Thus, for example, a 2.5 liter reservoir capable of withstanding 20 bar would be large enough in the example discussed to store a sufficient mass of inert gas to maintain 50 liters of inert gas in the principal configuration at 0.75 bar.
The system, with the R&IG configuration shown in
Before start-up, bidirectional isolating-valve 952 is closed and bidirectional GT pump 443 is controlled so that p*R tends to a preselected value p*RDis of p*R. The system is then, by definition, in its partial-minimum-pressure-maintenance mode.
At start-up, burners 180 are set to, say, their minimum delivery rate. Thereafter, as soon as the value of pR* exceeds p*RDis by a first preselected value, burners 180, valve 952, and pump 443, are controlled by the system's CCU (not shown) in a pre-prescribed way so as to keep the value of p*R within preselected limits. (The pre-prescribed way is application dependent.) As soon as the liquid level in condensate receiver 7 starts rising (because refrigerant is condensing), pumps 10 and 27 start running, and pump 443 continues running until the value of p*GR reaches p*GR,MAX. Thereafter 443 is controlled so as to keep the current value of p*GR close to p*GR,MAX, namely so as to keep the system in mode 3*. (The system has, except during transients, no other control mode while its principal configuration is active.)
To shut down, burners 180, valve 952, and pump 443, are controlled in a pre-prescribed way so as to maintain the value of p*R within the pre-prescribed limits. As soon as the value of p*R falls below a preselected value, valve 952 is closed. At this time, burners 180 are turned off if they have not already been turned off, and pump 443 is controlled so that p*R tends to p*RDis ; namely the system returns to its partial-minimum-pressure-maintenance mode.
Type B combinations can—like type A combinations—be endowed, where applicable, with one or more of the eight properties named complete minimum-pressure maintenance, partial minimum-pressure maintenance, freeze protection, self regulation, refrigerant-controlled heat release, gas-control led heat release, refrigerant-control led heat absorption, and evaporator liquid-refrigerant injection; and are suitable for several heat-transfer applications.
Type B combinations are usually employed where (1) it is more cost-effective to achieve complete minimum-pressure maintenance, partial minimum-pressure maintenance, or refrigerant-controlled heat release, with an inert gas instead of with liquid refrigerant; and where (2) freeze protection in the sense described under (a) to (e) in section III,E is required.
Type B combinations have, in addition to a principal configuration, an ancillary configuration and an inert-gas configuration. Type B combinations can in principle have any class of principal configuration, or any type of specialized principal configuration, employed by type A, or by type C, combinations. Type B combinations can, in principal, also have any one of the type IR to type VIR configurations, and any one of the type IG to type VG configurations, described earlier in this DESCRIPTION. Operating methods which can be used with type B combinations should be obvious in view of the operating methods of type A and type C combinations disclosed earlier in this DESCRIPTION. The techniques for charging type C combinations described in section V,H,11,c can mutatis mutandis also be used with type B combinations.
1. Type C Combinations with Complete Minimum-pressure Maintenance
a. Preliminary Remarks
Interconnections between a principal configuration and an IG or an IGP configuration of the same type C combination have so far only been shown for (1) a gas-cooled condenser, and in particular an air-cooled condenser, represented by a condenser having nominal y horizontal vapor and liquid headers; and for (2) a liquid-cooled condenser, and in particular a water-cooled condenser, represented by a block (black box) having inside it symbols
where one of those two symbols represents one or more refrigerant passages and where the other of those two symbols represents one or more fluid ways. (The modifier ‘nominally horizontal’ is used to indicate that the vapor and liquid headers are essentially horizontal while the platform on which they are located is horizontal.) Also, the location of the foregoing interconnections has been shown so far only for cases where a condenser is an entity which is separate and distinct from other components of a type C combination. In practice, a condenser—and particularly a gas-cooled condenser—of a type C combination is in certain important applications combined in a single physical structure with one or more other components of the combination.
Locating the foregoing interconnections correctly is essential in ensuring the optimal operation of an IG configuration. And the correct location of those interconnections can often not be shown without distinguishing between various kinds of condensers, and between various combinations of a particular kind of condenser with one or more other components of the type C combination to which the condenser belongs. I therefore in section V,K,1,b show the correct locations of the subject interconnections in several cases.
The location of port 440 of one-port IG or IGP configurations, and of port 471 of two-port IG or IGP configurations, is governed by the same considerations as those governing the location of air vents in non-airtight two-phase heat-transfer systems; and, in particular, in conventional steam heating systems. The selection of the subject location, together with the use of associated baffles, is discussed in detail in published documents; for example, in the three papers in the “Proceedings of the NATO Advanced Studies Institute on Thermal-Hydraulic Fundamentals and Design of Two-Phase Flow Heat Exchangers”, July 6-17, published by Kluwer Academic Publishers, ISBN 90-247-3693-5. The foregoing three papers are titled “Condensation with Non-Condensables and in Multicomponent Mixtures”, Michael K. Jensen; “Condensers and their Design”, D. Butterworth; and “Numerical Methods for the Analysis of Flow and Heat Transfer in a Shell-and-Tube Heat Exchanger with Shell-Side Condensation”, M Cumo.
I note that one of the reasons for using a two-port IG or IGP configuration, instead of a one-port IG or IGP configuration, is that the preferred location of port 440 for the purpose of removing inert gas from a principal configuration is unsuitable for the purpose of supplying inert gas to the principal configuration. I also note that cases exist where port 440, port 470, or port 471, may consist of several subports. One of those cases occurs where a distributed venting technique (discussed, for example, on pages 313 and 314 of the publication cited above in this section V,K,1,a) is used to remove inert gas from a principal configuration. That distributed venting technique is usually affordable only with shell-side condensation, or with tube-side condensation in condensers having only a few refrigerant passages.
In the examples given in section V,K,1,b, I show the location of port 471 of a two-port IG or IGP configuration. Where a one-port IG or IGP configuration is used, port 440 of the one-port configuration would be placed at the same location as port 471 of the two-port configuration.
In, for example,
In several FIGURES illustrating R&IG configurations, no trap has been shown to avoid cluttering those FIGURES. In practice a trap would be used with most R&IG or R&IGP configurations, and particularly with R&IG configurations having an IG configuration with a gas pump.
b. Examples of Interconnections
Air-cooled condensers with nominally horizontal headers may have inclined, or even horizontal, refrigerant passages. However, horizontal refrigerant passages should not be used in applications where those passages may be subjected to ambient temperatures at which liquid refrigerant in them may freeze. Where liquid refrigerant in air-cooled condenser refrigerant passages cannot freeze, air-cooled condensers with nominally vertical headers and horizontal refrigerant passages can be used as well as air-cooled condensers with nominally horizontal headers and refrigerant passages. Liquid refrigerant in condenser refrigerant passages may not freeze, even where their heat sink is very cold, for one of several reasons. For example, where a group H refrigerant (see section V,F,2,d) is employed, and the evaporator overfeed ratio is high, horizontal condenser refrigerant passages can often be employed. The reason for this is that, under the just-cited conditions, the concentration in those passages of the vapor of the refrigerant's component with the lowest freezing point will be nearly equal to the concentration of that component in its liquid phase.
2. Type C Combinations with Partial Minimum-pressure Maintenance
In certain type C combinations with partial minimum-pressure maintenance, it may be desirable to provide an inert gas outlet in the part of the combination's principal configuration outside the configuration's isolated segment into which inert gas is inserted. The reason for this is that the valves used to isolate that isolated segment may leak and allow some inert gas to exit that segment. For example valve 952 in
1. Preliminary Remarks
The control rules discussed in sections V,G and V,H, for the one or more controllable elements of an IG configuration, are as applicable p*R tends to p*DRo or to p*RD; p*GR stays close to p*GR3 or to pGR,MAX; TW tends to TWD; p*Ri tends to p*RDoi or to p*RDi; or TWi tends to TWDi. Also, the transition rules discussed in sections V,G and V,H, from mode 2*, or from mode 2*0, to mode 3* are one of the following six sets of rules:
p*GR=p*GR,3 and p*R>p*RD+Δp*R1 (28)
T*RSB=T*RSθ and p*R>p*RD+Δp*R1 (29)
VGR=VGR,MAX and p*R>p*RD+Δp*R1 (30)
p*GR=pGR,MAX and p*R>p*RD+Δp*R1 (31)
p*GR=pGR,MAX and TW>TWD+ΔTW1 (32)
p*GR=pGR,MAX and TIi>TIDi+ΔTI1 (33)
However, the foregoing control and transition rules may in certain applications be inadequate and should either be supplemented with additional rules, or replaced by alternative rules. Supplementary and alternative rules are discussed in section V,L,2.
I use the term ‘mixture purity’ to denote the mole fraction XG of inert gas in an inert-gas. and refrigerant-vapor mixture. In the case of an ideal gas—such as nitrogen up to 10 bar and at or above temperatures of about 290K—we have
where p*G is the total pressure of the mixture inside the trap, assumed constant throughout the trap and equal to the total pressure p*Ro of the mixture at, as applicable, port 440 or port 471; and where pG and pR are respectively the partial pressures of the inert gas and the refrigerant vapor in the mixture.
The maximum achievable mixture purity at the exit of a trap depends on the temperature of the one or more cold fluids used to remove heat from the mixture flowing through the trap. These cold fluids are usually the ambient air, and/or a cold-water supply or the sea. However, in some applications, the temperature of the naturally-available cold fluid or fluids may not be low enough to achieve the desired mixture purity at the trap's outlet. In such cases a refrigerated cold fluid may be used. For example, in the case of a trap belonging to an airtight configuration installed in a land vehicle, it may be desirable to use the ambient air as the trap's first component cold fluid, and to use the liquid phase of the refrigerant of the vehicle's air-conditioning system as the trap's second component cold fluid. Also, for example, in the case of an airtight configuration installed on the ground, it may be desirable to use water from the local water supply as the trap's first component cold fluid, and to use the liquid phase of the refrigerant of an air-conditioning system, or of a refrigeration system, as the trap's second component cold fluid. In either case the liquid refrigerant of the air-conditioning system, or of the refrigeration system, would be used to cool the mixture in the trap after it has been cooled by the first component cold fluid. I note that even in the former of the two last-cited examples the maximum cooling rate required to be provided by the second component cold fluid is typical y merely of the order of 100 watts where the maximum cooling rate of the airtight configuration is 50 kW.
To understand the importance of using a trap in certain applications, consider the case where (1) the refrigerant of an airtight configuration is water; (2) the temperature and the total pressure of the inert-gas and refrigerant-vapor mixture in the vicinity of, as applicable, port 440 or port 471 is 90° C.; and (3) the temperature of the mixture at the configuration's trap outlet is 45° C. Then, using relation (34), the mixture purity xGθ in the vicinity of port 440 or port 471 and at the trap's inlet is
namely the trap has increased the mixture purity from 0.31148 to 0.90532.
In airtight configurations having an IGP configuration or an IG configuration with no GT pump the increase in mixture purity achieved in the last-cited example by using a trap means that the internal volume of an IG reservoir required to accommodate a given mass of inert gas is reduced by a factor exceeding 2.9 (<0.90532/0.31148). I note that this reduction in volume could be achieved in principle by cooling the mixture in the IG reservoir instead of in a trap (as shown, for example, in
In airtight configurations with an IG configuration having a bidirectional GT pump or a unidirectional GT pump causing the mixture to flow from the airtight configuration's refrigerant circuits toward the airtight configuration's IG reservoir, an increase in mixture purity by a factor of 2.9 increases the rate at which a GT pump with a given inherent capacity pumps inert gas by a factor of 2.9, thereby reducing the pump's required inherent capacity by a factor of 2.9. Also increasing mixture purity allows, for a given pump compression ratio, the pump to be cooled to a lower temperature without refrigerant vapor condensing in the pump.
2. Control Techniques for Complete Minimum-pressure Maintenance and GAS-Controlled Heat Release
a. Discussion of Control and Transition Rules
Any set of control rules, and any set of associated transition rules, must together ensure that the condition cited next is satisfied: the current value of the total pressure p*R (in a principal configuration) just above the first (liquid-vapor) interface surface downstream, as applicable, from port 440 or from port 471 must be high enough, with respect to the current value of the refrigerant-vapor pressure pR at that interface surface, for 1o refrigerant-vapor bubbles to exit the surface; namely in symbols
p*Rθ>pRθ or equivalently T*RSθ>TRSθ (35)
where I have assumed that the current values of p*Rθ, pRθ, T*RSθ, and TRSθ, in the vicinity of, as applicable, port 440 or port 471, provide a sufficiently accurate measure of the current values of respectively p*R, pR, T*RS, and TRS, at the first liquid-vapor interface surface downstream from port 440 or from port 471. Consequently, where condition (35) is not satisfied automatically by the control and transition rules given in sections V,G and V,H, the appropriate controllable element of the IG configuration used should preferably be controlled so that it is satisfied, at least under steady-state operating conditions. This last statement is true for all IG configurations, but is especially true for IG configurations having a gas pump.
The control modes where condition (35) may sometimes not be satisfied are most likely to be control modes 2*, 2*0, and 3*. I therefore have devised alternative control rules for controlling the appropriate controllable element of an IG configuration in mode 2* or 2*0, and alternative control rules for controlling that element during mode 3*.
The expression ‘tends to’ used in describing control rules in modes 2* and 2*0, for the appropriate controllable elements of an IG configuration, can be expressed algebraically, as appropriate, by
FGD=KGP1 (p*R−p*RD) or FG=KGT (TW−TWD), (36), (37)
where FGD is the desired current value of volumetric flow rate FG of the inert-gas and refrigerant mixture entering an IG configuration; where KGP1 and KGT are preselected quantities which may each have a fixed value, or a value which changes in a pre-prescribed way as a function of one or more preselected characterizing parameters other than those used in relation (43) given below; and where the superscript ‘i’ would be added to the symbols appearing in relation (37) in the case of an intercooler. The alternative rules in modes 2 and 2*0 are identical to rules (36) and (37) while
Δp*Rθ>Δp*R0θ where Δp*Rθ≡p*RθpRθ (38), (39)
and where Δp*R0θ is a small fixed preselected value of Δp*Rθ. However, when
Δp*Rθ≦Δp*R0θ, (40)
the following two rules replace rules (36) and (37):
FGD=KGP1 (p*R−p*RD)·KMθ or FG=KGT (TW−TWD)·KMθ (41), (42)
where
where FG0 is a preselected value of FG chosen small enough to ensure the value of PRo does not exceed a preselected value. The quantity FG0 may be a fixed finite positive value; may be zero; or may change in a pre-prescribed manner as a function of a preselected characterizing parameter, for example of the parameter Δp*Rθ, and/or of its derivative with respect to time.
The expression ‘stays close to’ used in describing control rules, for the appropriate controllable element of an IG configuration in mode 3*, can be expressed algebraically, as applicable, by
FGD=KGV1 (VGR,3−VGR), FGD=KGV2 (VGR,MAX−VGR), or FGD=KGP2 (pGR,MAX−pGR), (44), (45), (46)
when respectively
(pGR,3−pGR)>ΔpGR,MAX, (VGR,MAX−VGR)>ΔVGR,MAX, (47a), (48a)
and
(pGR,MAX−pGR)>ΔpGR,MAX, (49a)
where KGV1, KGV2, and KGP2, are preselected quantities which may have a fixed value, or a value which varies in a pre-prescribed way; and by
FGD=0 (50)
when respectively
(pGR,3−pGR)≦ΔpGR,MAX; (VGR,MAX−VGR)≦ΔVGR,MAX; (47b), (48b)
and
(pGR,MAX−pGR)≦ΔpGR,MAX, (49b)
where ΔpGR,MAX and ΔVGR,MAX are small positive quantities, and where the superscript ‘i’ would be added to the symbols appearing in relations (44) to (49a) in the case of an intercooler.
A first set of alternative control rules in mode 3* are identical to those expressed in relations (44) to (49a) except for adding, as appropriate, the multiplier KMθ or KMθ to relations (44) to (46). A second set of alternative control rules replaces relations (44) to (46) by control rule
FGD=KMθ (51)
when
(p*Rθ−pRθ)>Δp*R,MAXθ (52a)
and by control rule (50) when
(p*Rθ−pRθ)≦Δp*Rθ, (52b)
where Δp*R,MAXθ is the maximum value of (p*Rθ−pRθ) for which the effectiveness of, as applicable, air-cooled condenser 508 or water-cooled condenser 594—or any other condenser of the principal F configuration of a system of the invention—is not degraded to an unacceptable degree by the presence of inert gas in its refrigerant passages.
A first set of alternative transition rules to those given in relations (28) to (33) is to merely delete the first condition in each of those relations. A second set of alternative transition rules is to replace the first condition in each of relations (28) to (33) by, as applicable, condition (52a) or (52b).
b. Implementation of Alternative and Supplementary Control and Transition Rules
The instrumentation of conditions (38) and (40), and of expression (43), can be implemented by using a proportional absolute-pressure transducer to obtain a measure of p*Rθ and a proportional temperature transducer to obtain a measure of TRSθ. The value of pRθ, corresponding to the value of TRSθ, can be obtained from published tables or graphs where the refrigerant employed is an azeotropic-like fluid, and from published tables and a prediction of the concentrations of the components of the refrigerant-vapor—in the neighborhood of, as applicable, port 440 or port 471—in the case where the refrigerant employed is a non-azeotropic fluid. The last-cited concentrations can usually be predicted to a sufficiently high accuracy from (1) the concentrations of the components of the non-azeotropic refrigerant with which the airtight configuration has been charged; (2) the value of pRθ; (3) the value of the evaporator overfeed rEO; and (4) the fraction of rEO supplied to a point of the principal configuration's principal circuits upstream from the first refrigerant liquid-vapor interface surface downstream from port 440 or port 471. For example, where an airtight configuration has been charged with an aqueous ethylene glycol solution having a glycol concentration c, the mean value of the liquid glycol concentration {overscore (c)}E in the configuration's evaporator, while the principal configuration is active, can be determined as a function of the evaporation pressure; of rEO; and, as applicable, of rM or rMA as defined in respectively relations (15) and (19). And, in turn, the glycol concentration cθ in the inert-gas and refrigerant-vapor mixture in the vicinity of port 440 or port 471 can be determined as a function of p*Rθ, usually assumed equal to the evaporation pressure; and as a function of rEO and of the fraction of rEO supplied to the refrigerant principal-configuration point cited earlier in the present minor paragraph under (4).
An example of the locations of a proportional absolute-pressure transducer providing a measure of p*Rθ, and of a proportional temperature transducer providing a measure of TRSθ, are the locations of respectively transducers 617 and 616 in FIG. 57A. (The temperature TRθ obtained from signal TRθ′ is equal to p*Rθ.) An alternative location for transducer 616 is shown in
Sometimes it may be practical and preferable to obtain measures of the total pressure p*Go and of the partial refrigerant pressure pR0 at a trap's outlet. In this case I use, instead of conditions (38) and (40), respectively conditions
Δp*Go>Δp*G0o and Δp*Ro≦ΔpR0o; (53), (54)
and, instead of expression (43), the expression
where, by the assumption I made immediately following relation (34),
Δp*Go≡pGo−pRo=p*G−pRo, (56)
and where Δp*G0o is the minimum permissible value of Δp*Go. In
Alternatively, I control the appropriate controllable element of an IG configuration so that the current value of FG tends to the predicted current value FGp computed by the CCU (of a system of the invention) on the basis of stored information on the performance of that controllable element as a function of the values of appropriate characterizing parameters obtained from appropriate transducers. For example, in the case where the IG configuration is a type IG or a type IVG configuration and where the controllable element is a variable-speed, electrically-driven, gas pump, the pump's inherent capacity can be predicted as a function of the value of its speed; of the temperature and total pressure of the particular inert-gas and refrigerant-vapor mixture entering the pump; and of the gas pump's compression ratio. And in the case where the IG configuration is a type IG, a type IIG, or a type IIIG configuration, the value FGp can be derived from the rate of change of the volume of the configuration's variable-volume IG reservoir. (In principle, the performance of a controllable element depends, for a given inert gas and a given refrigerant, on the ratio of the partial pressure of the refrigerant vapor in the inert-gas and refrigerant-vapor mixture entering a gas pump and entering an IG reservoir; but that partial pressure is usually only a small fraction of the total pressure of the mixture entering the gas pump and the IG reservoir, and therefore the effect of the presence of the refrigerant vapor in the mixture can usually be neglected. If that effect cannot be neglected it can be taken into account.
In some applications it is not necessary to measure the actual current value of FG, or to predict the current value of FG, and to use a servo which controls the appropriate controllable element of an IG configuration so that the current value of FG tends either to FGa or to FGp. Instead that controllable element is merely controlled so that the current value of FG tends to the expression given on the right-hand side of, as applicable, relation (41), (42), or (43). GT-pump recirculation valve 775 in
1. Cylinder-head and Cylinder-block Injection
2. Dry-up Prevention
Dry-up prevention mode 1*B is described in sections V,G,2,b,iv and V,H,8. In the former section electrically-driven SC pump 63h (see
I next describe a dry-up prevention technique which uses inert gas, stored in the IG reservoir of an R&IG or an R&IGP configuration, to force a portion of the liquid refrigerant into the configuration's evaporator refrigerant passages after the one or more refrigerant pumps of the R&IG or of the R&IGP configuration stop running. The last-cited technique can be used with a P evaporator, and with an NP evaporator having no liquid-refrigerant injectors. However, I choose to describe that technique for the case where the evaporator is an NP evaporator with liquid-refrigerant injectors, and where the airtight configuration is the IG configuration shown in FIG. 112A. In
In the case where an R&IG configuration stores at times, in its IG reservoir, inert gas at a pressure too high to cause liquid refrigerant to exit the liquid-refrigerant injectors' orifices at a low-enough rate, means must be provided for preventing inert gas being supplied to distributor 789 until the pressure in the inert gas reservoir falls below a preselected upper limit. This means may, for example, include using transducer 605, having inlet 770, to generate a signal p*′GR providing a measure of p*GR and opening valve 792 and closing valve 791 only after the motor stops running and the current value of p*GR falls below the preselected upper limit. The last-cited means may also include, as applicable, using or adding transducer 603 generating signal p*′GR providing a measure of p*R, and opening valve 792 and closing valve 791 only after pump 46 stops running and the current value of (p*GRp*R) falls below the preselected upper limit.
In certain cases, and particularly where an IG configuration has a GT pump used (in part) to store inert gas in the configuration's IG reservoir at a much higher pressure than the highest operating total pressure in the refrigerant circuits associated with the IG configuration, it may be desirable, or even necessary, to use, instead of on-off valve 792, a throttling valve, controlled by signal C′GTV5, to control the pressure at which inert gas is supplied to distributor 789 in
3. Spray Cooling
I have in section V,H,5,c,i used the term ‘evaporative spray cooling’ to denote techniques of liquid-refrigerant injection which achieve much higher heat-transfer coefficients than those achievable with pool boiling. I now distinguish between (1) ‘(evaporative) continuous spray cooling’ where the liquid-refrigerant jet, exiting the orifice of an LR injector, forms a continuous stream of liquid, and (2) ‘(evaporative) droplet spray cooling’ where the liquid-refrigerant jet exiting that orifice forms a stream of separate and discrete droplets. Test results using continuous spray cooling are given, for example, in a paper by M. Mondegand and T. Inoue titled ‘Critical Heat Flux in Saturated Forced-Convection Boiling on a Heated Disk, with Multiple Impinging Jets’, ASME co Transactions, Vol. 113, August 1991. And test results using droplet spray cooling are discussed, for example, in the paper by Tilton, Ambrose, and Chow, cited in section V,H,5,a of this DESCRIPTION, and in the paper by S. G. Yao and K. J. Choi titled ‘Heat Transfer Experiments of Mono-Dispersed Vertically Impacting Sprays’. Continuous spray cooling or droplet spray cooling can be used with ‘LR continuous injection’ or with ‘LR pulsed injection’. (See section V,H,5,c,iii for a definition of the last two terms.) An important difference between droplet spray cooling and LR pulsed injection, where they are used together, is that the period of the waveform of the droplet generator is much shorter than the shortest pulse produced by the valve controlling pulsed injection. For example, the period of the waveform generated by the droplet generator employed by Yao and Choi in the last-cited paper has a period in the range between 1.00 and 0.33 msec, whereas the shortest pulse produced by the valve controlling pulsed injection in the same airtight configuration would usually not be less than 10 msec. The waveform generated by the droplet generator could have a much higher frequency, say up to 20 kHz, than the 3000 Hz frequency corresponding to the 0.33 msec period. (The appropriate frequency depends on the desired droplet size along the direction of the liquid jet and on the jet's velocity.) The shape of the waveform generated may be square, sinusoidal, or may have another shape; but the waveform would usually be approximately symmetrical about its mean value, and would have a large enough amplitude and a low enough mean value for no liquid refrigerant to exit an LR injector's orifice when the waveform assumes its minimum value. (In electrical terminology the pulse train generated by the droplet generator and the valve controlling the pulse train would be a pulse-modulated carrier with 100% modulation.)
Any known device can be used as a droplet generator, including a diaphragm whose motion is controlled by a piezoelectric or a magnetostrictive device.
Liquid-refrigerant injection in general, and evaporative spray cooling in particular, are candidates for any application where the evaporator refrigerant passages of an airtight configuration, or of an evacuated configuration, experience high heat fluxes. For example, evaporators using evaporative spray cooling and water as the refrigerant would, for a given flame and combustion gas and for a given steam generation rate, be much smaller than conventional water-tube boilers because the critical heat flux of spray cooling is five to ten times higher than that of conventional forced-convection evaporative cooling. Where liquid-refrigerant injection is used instead of conventional water-tube boilers, at least a part of at least some of the boilers' water tubes would be replaced by an inner tube with orifices and an outer tube with no orifices. The inner tube is an LR injector and the space between the inner and outer tube constitutes an evaporator refrigerant passage. The inner and outer tubes need not have a common axis, and not need even be circular at low internal pressures.
1. Type C Combinations with Complete Minimum-pressure Maintenance
In the first case of the three cases recited in section V,H,11,c, I distinguish between applications where (1) air inside an R&IG enclosure does not react chemically with the refrigerant inside the enclosure, or with the internal surfaces of the walls of the enclosure, and where (2) the oxygen in the air reacts with that refrigerant, or with those surfaces, and is depleted without causing a significantly adverse effect. (If oxygen were depleted while causing a significant adverse effect, air would not be an inert gas in the sense the term ‘inert gas’ is defined in definition 72 in section Il,A,2.) In the applications cited under (2) in this minor paragraph, the predetermined charging value of the total pressure inside an R&IG enclosure is established by taking into account the reduction in total pressure at a given temperature, caused by depletion of the oxygen originally contained in the air inside the enclosure.
In the third case of the three cases recited in section V,H,11,c, I distinguish between applications where (1) a small fraction, say a few percent, of the mass of the non-condensable gas inside an R&IG enclosure, after the enclosure has been charged with refrigerant and inert gas, may be air; and where (2) it is desirable or essential to minimize the mass of air remaining inside the enclosure after it has been charged with refrigerant and inert gas. In the latter applications, the following flushing method steps may be used: (i) with, where applicable, all valve passages of the R&IG enclosure open, inert gas is inserted into the enclosure until the total pressure inside the enclosure reaches the maximum permissible value; (ii) inert gas and air inside the enclosure are allowed to diffuse throughout the enclosure until the inert gas and the air inside the enclosure form a quasi-homogeneous mixture; and (iii) the mixture is purged until its total pressure inside the enclosure is at, or slightly above, ambient atmospheric pressure. Steps (i) to (iii) are repeated until the mass of air, and in particular of oxygen, remaining inside the R&IG enclosure is less than or equal to a preselected upper limit. The technique or the method used to respectively measure or predict the mass of air remaining inside an R&IG enclosure depends on the inert gas used and on the design of the airtight configuration to which the enclosure belongs. Where applicable, permissible, and possible, the airtight configuration's refrigerant pumps and GT pumps are run during step (ii) to assist the diffusion process by forced convection. In the case of the R&IG configuration shown in
An airtight configuration may be fixed to the ground and have essentially only one environment, or may be installed on a moving platform having changing environments. I shall refer to the effective temperature of an airtight system's current environment as the ‘environmental temperature’. The term ‘effective temperature’ takes into account, in addition to ambient temperature, radiant energy absorbed by the airtight configuration and radiant energy released by the airtight configuration to remote, including celestial, bodies. The temperature of, as applicable, the refrigerant enclosure, or the R&IG enclosure, of an airtight configuration is equal to the environmental temperature when the enclosure is in thermal equilibrium with its current environment.
An airtight configuration of a type C combination can almost always achieve complete minimum-pressure maintenance, over the entire range of its environmental temperatures, if the (total) internal pressure throughout the airtight configuration's R&IG enclosure can be maintained at or above a preselected minimum-pressure-maintenance value when the enclosure's temperature is equal to the design lowest environmental temperature. An airtight configuration is usually not charged at the design lowest environmental temperature. I have therefore devised a method for determining the internal pressure at which the configuration should be charged for minimum-pressure maintenance to be achieved while keeping the value of that internal pressure as low as possible at the design highest environmental temperature. To this end I use the relation
where TC and T0 are respectively the value of the configuration's charging temperature and of the configuration's design lowest environmental temperature; where VL,C is the value of the liquid refrigerant volume at TC; where VT is the total volume of the enclosure, assumed constant except for thermal expansion; where VL,0 is the value of the volume VL of the liquid refrigerant in the enclosure at T0; where p*c and p*0 are the configuration's internal pressure at respectively TC and T0; where ΔEVT is the increase in the value of VT, caused by thermal expansion, as the enclosure's temperature increases from T0 to TC; where MG is the inert-gas mass with which the enclosure is charged (at TC); and where ΔMG is the mass of inert gas which comes out of solution as the temperature of the enclosure increases from T0 to TC. The value of ΔMG can be obtained for a specific refrigerant, and a specific inert gas, as a function of temperature, total pressure, and concentration, from published tables. The definition of, and the sign preceding, ΔMG in relation (57) assumes the refrigerant with which the airtight configuration is charged includes nitrogen in solution. If that refrigerant does not include nitrogen in solution, the definition of ΔMG would be the mass of inert gas in solution at T0, and the negative sign preceding ΔMG would be replaced by a positive sign.
Relation (57) applies to an airtight configuration having a fixed-volume IG reservoir. It also applies to an airtight configuration having a variable-volume IG reservoir which is constrained so that its internal volume VGR has its minimum possible value ΔVGR,MIN while the configuration is being charged, and while the configuration's enclosure is at T0—provided VT,0 includes VGR,MIN. If a variable-volume IG reservoir of an airtight configuration is not thus constrained and the value of VT increases from VT,0 at T0 to VT,C at TC, the second factor in relation (57) is replaced by the factor
2. Type C Combinations with Partial Minimum-pressure Maintenance
In the case of partial minimum-pressure maintenance, relation (57), or relation (58) with the second factor replaced by Kv, can in essence be used for the isolated part of the R&IG enclosure (of a type C combination) into which inert gas is inserted.
GT pumps may, depending on the application, be multi-stage compressors with or without intercooling, or may be single-stage compressors. In either case, the effective capacity of a GT pump may be non-zero for time intervals which are substantially shorter than the time intervals during which the effective capacity of a GT pump is zero.
The effective capacity of a GT pump may be zero either because it is not running, or because a valve is used to cause the pump's effective capacity to be zero while the pump is running. A GT pump, in a correctly designed circuit, consumes no power while it is not running; and consumes, while it is running with a given inherent capacity and a zero effective capacity, only a small fraction of the power the GT pump would consume at the same inherent capacity if its effective capacity were equal to its inherent capacity. An example of a correctly designed circuit in the latter case is shown in
In cases where the time intervals during which a GT pump has a non-zero effective capacity are short compared to the time intervals during which the pump has a zero effective capacity, it is often advantageous to place a cylinder of a GT pump in direct physical and thermal contact with an IG reservoir supplied with the inert-gas and refrigerant-vapor mixture exiting the pump. This last statement assumes the walls of the pump and the reservoir are made of thermally-conducting material. The specific advantages obtained by the last-cited direct contact—which inherently combines the thermal capacities of the cylinder and the reservoir—depend on several design parameters, including the relative magnitudes of the masses of the cylinder and the reservoir; and, in the case where the cylinder and the reservoir are air-cooled, on the relative magnitudes of the external surfaces (including extended surfaces) of the cylinder and the reservoir. Generally speaking, however, the foregoing specific advantages include reducing, at a given pump inlet pressure, the reservoir's internal pressure and the pump's compression ratio; and/or reducing the temperature of the inert-gas and refrigerant-vapor mixture exiting the pump. Furthermore, in cases where the pump is driven by a motor which runs even while the pump's effective capacity is zero, the foregoing specific advantages also include reducing the temperature of the pump while it runs with zero effective capacity. The achievable reduction in the last-cited temperature can be substantial in several cases, including in cases where the external surface of the reservoir is considerably larger than the external surface of the cylinder of the GT pump in direct physical and thermal contact with the reservoir.
To maximize the reduction in an IG reservoir's internal pressure achievable by placing a GT pump and an IG reservoir in direct physical and thermal contact, it is necessary to maximize the rate at which the inert-gas and refrigerant-vapor mixture, inside the reservoir, transfers heat by convection to the reservoir's walls. T0 this end, I use known techniques for extending the internal surface of the reservoir's walls, and for increasing the rate at which the mixture circulates by natural convection inside the reservoir. The foregoing known techniques include using one or more low-pitch spirals, made of thermally-conducting material, inside and in direct physical and thermal contact with the inner surface of the reservoir's walls. The surface of the spirals extends the internal surface of the reservoir's walls, and the spirals are located and configured so that the velocity of the inert-gas and refrigerant-vapor mixture with respect to the internal surface of the reservoir's structure is higher than that velocity would have been in the absence of those spirals.
An example of a configuration where the cylinder of a GT pump is placed in direct physical and thermal contact with an IG reservoir is shown in
Pressure-equalization lines have been so far shown in this DESCRIPTION in
In certain special applications it may be desirable for an airtight configuration, or an evacuated configuration, to include a refrigerant diverting valve and a refrigerant line for by-passing refrigerant around the configuration's condenser.
It is well know that the rate {dot over (Q)}abs at which the heat-transfer fluid of a prior-art ‘heat-transfer system’, as defined in section 1, absorbs heat from a combustion gas obtained by burning a fuel can be estimated from the fuel's mass-flow rate {dot over (m)}F, or almost equivalently from the fuel's volumetric flow rate FF. It is also well known that the accuracy of the estimate of {dot over (Q)}abs can be increased in heating systems, and in engine-cooling systems, by also using the air-fuel ratio {dot over (m)}F/{dot over (m)}A, where {dot over (m)}A is the mass-flow rate of the air used to burn the fuel. It is also well known (see for example ‘Internal Combustion Engine Fundamentals’ by John B. Heywood, McGraw 1988, section 12.7.2 and the references cited therein) that the accuracy of the subject estimate can be further increased in the case of engine-cooling systems by using the current value of additional parameters such as engine intake temperature and, as applicable, spark timing or fuel-injection timing.
I assert that the facts cited in the immediately-preceding minor paragraph apply also to airtight systems (of the invention) used for heating and/or cooling, and that the invention includes using those facts for estimating the rate at which airtight systems absorb heat from a combustion gas. I note that the last-cited rate can be used to estimate the resulting evaporation rate of an airtight system by using
{dot over (m)}θ=({dot over (Q)}abs−cpi{dot over (m)}CΔSb1T−cpi{dot over (m)}EΔSb2T−cpg{dot over (m)}CΔSbT) (59)
where {dot over (Q)}abs is the rate at which heat is absorbed by the refrigerant from the combustion gas; where {dot over (m)}C and {dot over (m)}E are the mass-flow rates of the refrigerant through respectively the condenser and the evaporator refrigerant passages; where ΔSb1T and ΔSb2T are the amounts (expressed in degrees Celsius) by which respectively the flow rates {dot over (m)}C and {dot over (m)}E are subcooled; where cpi and cpg are the specific heats of the refrigerant in respectively its liquid and vapor phases; and where ΔShT is the amount by which the flow rate {dot over (m)}C is superheated.
For examples of industrial applicability see section III,C.
The present application is a continuation-in-part of my PCT patent application Ser. No. 92/01654, filed Mar. 11, 1992, titled AIRTIGHT TWO-PHASE HEAT-TRANSFER SYSTEMS, and of my application Ser. No. 07/400,738, filed Aug. 30, 1989, now U.S. Pat. No. 5,333,677 titled EVACUATED TWO-PHASE HEAT-TRANSFER SYSTEMS; the cited PCT application being a continuation-in part of my application Ser. No. 07/400,738, filed Aug. 30, 1989, now U.S. Pat. No. 5,333,677 titled EVACUATED TWO-PHASE HEAT-TRANSFER SYSTEMS and of my then-pending application Ser. No. 07/696,853, filed May 7, 1991, now abandoned titled TWO-PHASE HEAT-TRANSFER SYSTEMS; the last-cited patent application being a continuation-in-part of my application Ser. No. 07/400,738, filed Aug. 30, 1989, now U.S. Pat. No. 5,333,677 titled EVACUATED TWO-PHASE HEAT-TRANSFER SYSTEMS and of my then-pending application Ser. No. 06/815,642, filed Jan. 2, 1986, now abandoned titled TWO-PHASE HEAT-TRANSFER SYSTEMS; the first of the two last-cited patent applications being a continuation-in-part of the second of the two last-cited patent applications; and the second of the two last-cited patent applications being a continuation-in-part of five following then-pending applications: (a) Ser. No. 06/374,707, filed May 4, 1982, now abandoned titled TWO-PHASE HEAT-TRANSFER SYSTEMS,(b) Ser. No. 06/362,148, filed Mar. 26, 1982, now abandoned titled VAPOR-GENERATING SYSTEMS,(c) Ser. No. 06/361,784, filed Mar. 25, 1982, now abandoned titled POWER SYSTEMS,(d) Ser. No. 06/355,520, filed Mar. 5, 1982, now abandoned priority date Jul. 7, 1980 (namely filing date of PCT/US80/0089), titled SOLAR TWO-PHASE HEAT-TRANSFER SYSTEMS, and(e) Ser. No. 06/235,980, filed Feb. 19, 1981, now abandoned titled FORCED REFRIGERANT CIRCULATION SOLAR HEATING SYSTEMS. The first four of the five last-cited applications were continuations-in-part or divisional applications of then co-pending applications (1) Ser. No. 252,206, filed Apr. 8, 1981, titled FORCED REFRIGERANT-CIRCULATION SOLAR HEATING SYSTEMS,(2) Ser. No. 252,205, filed Apr. 8, 1981, titled TWO-PHASE SOLAR HEATING SYSTEMS,(3) Ser. No. 144,275, filed Apr. 28, 1980, titled SOLAR POWER SYSTEM, now U.S. Pat. No. 4,358,929, and(4) Ser. No. 902,950, filed May 5, 1978, titled SOLAR HEATING SYSTEM, now U.S. Pat. No. 4,340,030. The last two patent applications were continuations-in-part of then Ser. No. 457,271, filed Apr. 2, 1974, titled HEATING AND COOLING SYSTEMS, now U.S. Pat. No. 4,211,207; and application Ser. No. 235,980, filed Feb. 19, 1981, was a divisional application of then-pending application Ser. No. 902,950, and was filed for the purpose of provoking an interference with Bottum U.S. Pat. No. 4,220,138, filed Jan. 24, 1978.
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| Number | Date | Country |
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| Number | Date | Country | |
|---|---|---|---|
| Parent | PCTUS92/01654 | Mar 1992 | US |
| Child | 08112204 | US | |
| Parent | 07696853 | May 1991 | US |
| Child | PCTUS92/01654 | US | |
| Parent | 07400738 | Aug 1989 | US |
| Child | 07696853 | US | |
| Parent | 06815642 | Jan 1986 | US |
| Child | 07400738 | US | |
| Parent | 06374707 | May 1982 | US |
| Child | 06815642 | US | |
| Parent | 06362148 | Mar 1982 | US |
| Child | 06374707 | US | |
| Parent | 06361784 | Mar 1982 | US |
| Child | 06362148 | US | |
| Parent | 06355520 | Mar 1982 | US |
| Child | 06361784 | US | |
| Parent | 06235980 | Feb 1981 | US |
| Child | 06355520 | US |
