EP0055712A1

EP0055712A1 - Solar two-phase, heat-transfer systems

Info

Publication number: EP0055712A1
Application number: EP81901386A
Authority: EP
Inventors: Stephen Molivadas
Original assignee: Individual
Current assignee: Individual
Priority date: 1980-07-07
Filing date: 1980-07-07
Publication date: 1982-07-14
Also published as: AU7223881A; WO1982000191A1

Abstract

Systemes ameliores de transfert de chaleur a deux phases permettant d'absorber la chaleur de l'energie solaire radiante, et de transferer cette chaleur vers une substance a chauffer. Le fluide de circulation est appele refrigerant. Les systemes de l'invention possedent la propriete decisive d'autoregulation. Le terme autoregulation est utilise pour decrire la capacite d'un systeme de transfert de chaleur a deux phases d'assurer automatiquement qu'au moins les quatres conditions de fonctionnement interne citees ci-apres soient remplies pour toutes les conditions exterieures pour lesquelles ce systeme est concu: 1) le refrigerant (2) entrant dans l'absorbeur collecteur (1) existe exclusivement dans sa phase liquide et possede un debit suffisamment important pour maintenir la quantite de chaleur de la portion evaporee du refrigerant (3) sortant de l'absorbeur collecteur (1) en-deca d'une limite superieure positive predeterminee; 2) la vapeur du refrigerant (9) entrant dans le condenseur (8) est maintenue a l'etat sec; 3) la quantite de refrigerant liquide d'appoint dans le condenseur (8) est suffisamment petite pour maintenir la surface de condensation effective du condenseur au-dessus d'une limite superieure positive predeterminee; et 4) la valeur absolue de la difference entre la temperature de la vapeur saturee (3) sortant de l'absorbeur collecteur (1) et la temperature de la vapeur saturee (9) du refrigerant entrant dans le condenseur (8) est maintenue en-deca d'une limite superieure predeterminee.Improved two-phase heat transfer systems for absorbing heat from radiant solar energy and transferring that heat to a substance to be heated. The circulation fluid is called refrigerant. The systems of the invention have the decisive property of self-regulation. The term self-regulation is used to describe the ability of a two-phase heat transfer system to automatically ensure that at least the four internal operating conditions listed below are met for all of the external conditions for which this system is designed: 1) the refrigerant (2) entering the collecting absorber (1) exists exclusively in its liquid phase and has a sufficiently large flow rate to maintain the quantity of heat of the evaporated portion of the refrigerant (3) leaving the absorber collector (1) below a predetermined positive upper limit; 2) the vapor of the refrigerant (9) entering the condenser (8) is kept dry; 3) the amount of liquid back-up coolant in the condenser (8) is small enough to keep the effective condensing area of the condenser above a predetermined positive upper limit; and 4) the absolute value of the difference between the temperature of the saturated vapor (3) leaving the collecting absorber (1) and the temperature of the saturated vapor (9) of the refrigerant entering the condenser (8) is maintained in -deca of a predetermined upper limit.

Description

D E S C R I P T I O N

SOLAR TWO-PHASE, HEAT-TRANSϊΕR SYSTEMS

1. TE CHN I CAL F I EL D

The invention disclosed in this patent application relates to improved two-phase heat-transfer systems for absorbing heat from solar radiant energy, and for transferring and releasing this heat to a substance to be heated.

The term "two-phase heat-transfer system" is used here to denote a system that employs a circulating fluid to absorb heat from a heat source, at least in part by evaporation, at an essentially constant temperature, and to transfer this heat to a substance to be heated by releasing it, at least in part by condensation, at an essentially constant lower temperature. The heat thus absorbed or released is called "latent heat". II. B A C K GRO UN D ART

Two-phase heat-transfer systems can, with available fluids, solve many of the known problems of solar heating systems employing a single- phase heat-transfer system; namely, a system that employs a circulating fluid which absorbs heat from a heat source, and transfers this heat and releases it to a substance to be heated, without changing phase throughout a heat-transfer cycle. In fact, many expired and unexpired patents have already been granted for such two-phase heat-transfer systems which use solar radiant energy as the heat source. However, these prior-art systems fail to exploit the full potential of latent-heat transfer, in part, because they do not possess the crucial property of "self-regulation".

By contrast, all the two-phase, heat-transfer systems disclosed in this document do possess this crucial property which constitutes the link which unites the various species and sub-species disclosed into a single inventive concept. Furthermore, almost all envisioned practical embodiments of the systems disclosed possess also either the property of "inherent over-pressure protection", or the property of "combined overpressure and over-temperature protection". Without one of these two additional crucial properties, the property of "self-regulation" would have a much smaller beneficial impact on the technology of two-phase heat-transfer systems for solar heating applications.

The foregoing properties, and the properties of "freeze protection" and "sub-atmospheric protection" which are also possessed by certain embodiments of the invention are defined and discussed in PART III of the DESCRIPTION.

Surrounding U.S. patents of the invention include Newton 3,260 064, Snelling 3,390,673 and Traylor 2,434,086.

Related patent applications filed with the U.S. Patent Office by this applicant are applications Serial No. 457,271 and Serial No, 902, 950, filed 2 April 1974 and 4 May 1978, respectively, III. DI S CLO S URE O F I NVE N T I O N

A. PURPOSE AND SCOPE OF INVENTION

The purpose of the invention is to improve prior-art two-phase heat-transfer systems for solar heating systems by devising techniques for achieving the properties mentioned in PART II of this DESCRIPTION.

B. DEFINITION AND DISCUSSION OF TERMS

Certain terms used in this DISCLOSURE OF INVENTION shall have the following meaning for the purpose of this patent application: (a) A "refrigerant" is defined as the heat-transfer fluid of a two-phase heat-transfer. Thus the term "refrigerant" is used to denote the function of a fluid, and not its nature; and specifically to denote the function of absorbing heat from a heat source, at least in part by evaporation, and of releasing heat to a heat sink, at least in part by condensation. The term "heat sink" in the context of this definition of a refrigerant includes a sink of finite or infinite thermal capacity and, in particular, the substance to be heated, or being heated, by a two-phase heat—transfer system, (b) The term "absorber" includes any means employed to absorb heat from solar radiant energy and to transfer this heat by conduction or convection, or both, to the refrigerant flowing through one or more refrigerant passages which are an integral part of the absorber, or which are in thermal contact with it; it also includes any means employed to form one or more transparent passages in which the refrigerant itself absorbs heat "directly" from solar radiant energy, where the. term "directly" is not intended to exclude reflected or refracted solar radiant energy by the concentrator (optical system) of a solar focusing collector. Whenever the absorber of a number of solar collector modules are connected in series, the term "absorber inlet" refers to the inlet of the first absorber which the refrigerant of a heat-transfer system enters, and the term "absorber out-let" refers to the outlet of the last absorber from which this refrigerant exits. (c) the term "condenser" includes any means in which a refrigerant releases heat to the substance to be heated at least in part by condensation. This substance may be a solid, or a fluid, which absorbs sensible heat, latent heat, or both. C. DISCUSSION OF KEY DISTINCTIVE PROPERTIES OF VARIOUS EMBODIMENTS OF THE INVENTION 1. S e l f-Re g u l a t i o n a. General

Broadly speaking, the term "self-regulation" is used here to denote the capability of a two-phase, heat-transfer system to exploit the full potential of the latent-heat properties of the: refrigerant by

(a) adjusting automatically without using a (vapor) throttling valve — the flow of this refrigerant, and by

(b) correctly apportioning the liquid phase of this refrigerant between different parts of the. refrigerant circuits, within the system.

By contrast, prior-art systems are only capable of exploiting the latent-heat properties of their refrigerant over a fixed narrow range of external conditions determined by the system's design.

More precisely, the term self-regulation is used to denote the capability of a two-phase, heat-transfer system to ensure automatically that at least the four internal operating conditions cited below are satisfied for all external conditions under which this system is designed to operate: first, the refrigerant entering the collector absorber exists exclu sively in its liquid phase and has a mass flow rate large enough to maintain the amount of superheat of the evaporated portion of the refrigerant exiting the collector absorber below any preselected, positive, upper limit; second, the refrigerant vapor entering the condenser is maintained in a dry state; third, the amount of liquid refrigerant which backs up in the condenser is small enough to maintain the condenser's effective condensing surface above a preselected, positive, upper limit; and fourth, the absolute value of the difference between the saturated vapor temperature exiting the collector absorber and the saturated vapor-temperature of the refrigerant entering the condenser is maintained below a preselected, upper limit. I shall henceforth refer to a two-phase, heat-transfer system which possesses the property of self-regulation as a Rankine heat- transfer system because the thexmodynamic cycle, which I shall refer to as a Rankine heat-transfer cycle, traversed by its refrigerant: resemble the cycle which would be traversed by the working fluid of a Rankine power cycle if the engine were removed.

The property of self-regulation is particulary important in the case where a two-phase, heat-transfer system is used in a solar-heating application. The reasons for this fact are discussed next. First, in systems where the heat source is solar radiant energy at the earth's surface, the intensity of the heat source varies over a wid range of values which are not under the control of the system; whereas, in systems — which I shall refer to as conventional two-phase heat- transfer systems — where the heat source is a conventional fuel, or electrical energy, the uncontrolled variation of the intensity of the heat source varies over a much narrower range of values.

Second, in most solar-heating applications a low rate of heat collection, often less than one tenth of the maximum rate, can be utilized and therefore the useful range of uncontrolled evaporation rates over which self-regulation should be achieved in solar heat-transfer systems may often be over 10:1; whereas in conventional two-phase heat-transfer systems the range of uncontrolled evaporation rates is usually less than 1%:1.

Third, the net rate at which heat is absorbed by the refrigerant in a solar collector absorber, and hence also the evaporation rate of this refrigerant, depends not only on the intensity of solar radiation; but also on the evaporation temperature and the equivalent temperature of the collector's surrounding, which can both vary over a wide range of values. Fourth, in solar heat-transfer systems, the heat losses of the boiler, namely of the solar-collector absorber, to its surroundings are many times greater than those of the boiler of conventional heat-transfer systems -to.its surroundings; and furthermore these heat losses increase much more rapidly in the former systems than in the latter systems for a given increment of the difference in temperature between the boiler and its surroundings. (The reason for this fact is that the solar-collector absorber of a heat-transfer system cannot be thermally insulated from its surroundings to the same degree as the boiler of a conventional heat-transfer system without, at the same time, suffering an unacceptable decrease in the amount of solar radiant energy intercepted by the absorber surface in direct thermal contact with the refrigerant fluid.) Fifth, the solar-collector absorber of solar heat-transfer systems is exposed directly to the "equivalent temperature" of the local outdoor surroundings; whereas the boiler of conventional heat-transfer systems is usually located in a building and is therefore at most exposed only indirectly to this equivalent temperature, namely through a building's structure. Therefore, the solar-collector absorber's surroundings can be much colder than those of the boiler, and consequen tly, even if this absorber were thermally insulated from its surroundings to the same degree as the boiler, its heat losses to these surroundings in cold weather would be much greater than those of the boiler.

It follows from the first three facts that the inability of a two-phase heat-transfer system to adjust the flow rate of its refrigerant, and to apportion the liquid phase of this refrigerant between different parts of the system's refrigerant circuits, over a wide range of evaporation rates and temperatures, results in a much greater degradation in system thermal performance when the system uses the sun's radiation instead of a conventional source of energy as its heat source. It also follows from the fourth and fifth facts that the evaporation temperature should, where the sun's radiation is used as the heat source, be kept as low as possible, and should therefore exceed the temperature of the substance being heated by the minimum amount practicable. Consequently, the condensation temperature of the refrigerant should at all times exceed its evaporation temperature by the smallest amount practicable which requires that the adjustment of refrigerant flow rate be accom plished without using a (vapor) throttling valve. b. Self-Regulation with no Independent Subcooling Control

In many solar heating applications the substance being heated is, at any given time, at a nearly uniform temperature. (A prominent example of such an application is the case where the heat-transfer system is used to heat the water in the storage tank of a domestic or service-water heating system.) In most such applications, no significant amount of refrigerant subcooling is required, and in this case, the third internal operating condition for self-regulation becomes "the amount of liquid refrigerant actually present in the condenser, in excess of the amount which would be present if the condensate at the condenser's refrigerant exit were not subcooled, in essence, is zero". The qualification, "in essence" is used, to allow this modified third internal operating condition to include the case where the condensate at the condenser's exit is subcooled by about one or two degrees Celsius. (This amount of subcooling at the condenser refrigerant exit may often be desirable or even required to avoid condensate pump cavitation.) In the case where no significant subcooling is required, and the quality of the refrigerant vapor is one (namely it is both dry and not superheated), self-regulation may also be defined as the capability of a two-phase heat-transfer system to ensure automatically, for all external conditions under which the system is designed to operate, that in essence

(a) only latent heat is absorbed or released by the refrigerant; and that in essence

(b) the saturated temperature of the refrigerant at the inlet of the condenser is equal to the saturated vapor temperature of the refrig erant at the exit of the absorber.

In this alternative definition of self-regulation, the first three internal operating conditions for self-regulation have been replaced by the first of the two foregoing conditions, and the fourth internal operating condition for self-regulation has been replaced by the second of these two conditions. This alternative definition of self-regulation, although less explicit or precise, may better convey the meaning and purpose of self-regulation in the particular case where no significant amount of superheat is required at the absorber outlet, and no significant amount of subcooling is required at the condenser's refrigerant outlet.

In some cases, — which include even solar-heating applications where the substance being heated is at a nearly uniform temperature — a significant amount of subcooling, at the condenser's refrigerant outlet may be desirable. A significant amount of subcooling, defined as over say 5°C, can as explained later be provided, while also achieving self-regulation, with the same techniques used to achieve self-regulation when no significant amount of subcooling is required (or desired). The alternative definition of self-regulation given earlier can be broadened to cover the case where a small amount of superheat is required at the absorber outlet, and a significant amount of subcooling is required at the condenser's refrigerant exit, by re-writing the modified third condition as follows: "the maximum possible amount of latent heat is absorbed or released by the refrigerant within the limits imposed by the amount of superheat desired at the absorber's outlet, and by the amount of sub-cooling required at the condenser's refrigerant outlet". c. Self-Regulation with an Independent Subcooling Control In some solar heating applications the substance being heated may be a fluid which is, at any given time, being both preheated over a wide range of temperatures in its liquid phase and vaporized at a tempera ture near the upper end of this range. (An example of such an application is the case where a heat-transfer system is used to preheat and vaporize the working fluid of a heat engine.) In such cases efficient heat transfer from the refrigerant to the working fluid requires the refrigerant to be subcooled by a large amount, namely by an amount where the heat released by the refrigerant to the working fluid during subcooling may be as great as one third (or even more) of the heat released by the refrigerant to the working fluid during condensation. In such cases, it is nearly always more cost effective to employ a physically distinct means, referred to here as a subcooler, in which the refrigerant is subcooled and the working fluid is preheated, from the means, referred to here as a condenser, in which the refrigerant is con densed and the working fluid is vaporized.

A large amount of subcooling, whether or not a separate subcooler is used, can also be provided, while also achieving self-regulation, with the same techniques used to achieve self-regulation when no significant amount of subcooling is required or desired. However, where a large amount of subcooling is required, it may be desirable to change automatically the ratio of the amount of heat released by the refrigerant while it is being subcooled to the amount of the heat released by the refrigerant while it is being condensed. (For example, in the case of a heat engine it may be desirable to change this ratio as a function of the engine's load or condensing temperature.) In this case a separate subcooling control is used, as explained later, to control automatically the amount of refrigerant subcooling. 2. Ov e r - p re s s u r e P r o t e c t i o n

In discussing over-pressure protection, freeze protection, and sub-atmospheric protection, I distinguish Forced Refrigerant circulation species, or more briefly FRC species, from Natural Refrigerant Circulartion species, or more briefly NRC species. The former use at least one refrigerant pump and are most often used where the condenser refrigerant outlet is below the absorber inlet; whereas the latter are only used where the condenser refrigerant outlet is above the absorber inlet. The equilibrium temperature of the absorber surface of a solar collector while it is intercepting solar radiation, and while no heat is removed from it by the refrigerant, is called the collector stagnation temperature. This temperature is, in general, much higher than the maximum design operating temperature.

As long as both phases of a fluid are present in the passages of a solar collector, the pressure of the vapor of the fluid under steady-state conditions, will be equal to the saturated pressure corresponding to that temperature. A collector absorber panel designed to withstand pressures corresponding to collector stagnation temperatures would be unnecessarily heavy and expensive. (For example, the saturated pressure of R-114 is 439 psig at a stagnation temperature of only 290°F; whereas the maximum operating pressure of R-114 to heat water up to 160°F and 175°F is only about 100 psig and 125 psig respectively.)

The term "over-pressure protection" denotes that the refrigerant is prevented from exceeding significantly its maximum design operating pressure under any stagnation condition. In most FRC species installations this protection is achieved by the very nature of the refrigerant circuit configuration itself. Namely, the circuit configuration allows the entire volume of liquid refrigerant to be stored below the solar-collector absorber whenever the heat-transfer system is inactive, and permits all liquid refrigerant to be removed from the collector absorber upon system de-activation solely under the influence of gravity and by the process of vapor condensation in, and migration to, the part of the refrigerant circuits below the absorber.. I refer to this type of overpressure protection as "inherent over-pressure protection" In NRC species installations and in some FRC species installations, liquid refrigerant is stored entirely, or in part, above the lowest level of the absorber. In this case a valve is used to prevent liquid refrigerant entering the absorber whenever the refrigerant pressure rises above a preselected upper limit. The valve is operated in a "fail-safe manner" unaffected by electrical mains or water mains failure, or by a malfunction of the heat-transfer system's electrical or pneumatic controls. I refer to this type of over-pressure protection as "fail-safe over— pressure protection".

3. O v e r -P r e s s u r e a n d O v e r- T e mp e r a t u r e P r o t e c t i o n

The maximum stagnation temperature which can be attained by the absorber of a collector may be too high either for the refrigerant or for the absorber itself. In this case over-temperature, as well as over-pressure, protection is provided by rejecting unwanted heat to an external heat sink which will usually be water or the outdoor ambient air.

4. F re e z e P r o t e c t i o n In some applications, the most suitable refrigerant may freeze at some of the equivalent temperatures of the collector's outdoor surroundings; and furthermore this refrigerant may be of the type which expands as it changes from its liquid to its solid phase. The most common example of such a refrigerant is water, which is probably the preferred refrigerant for most heating applications in the range between 125°C and 200°C. The present invention provides means for protecting FRC species systems against damage caused by such a refrigerant when it freezes. 5. S ub -A t mo s p h e r i c P r o t e c t i o n

Some refrigerants, such as R-114, may boil at a temperature below the minimum temperature of the enclosure in which a part of a two-phase heat-transfer system is enclosed, but above some of the equivalent temperatures of the outdoor surroundings of the system's collector. In this case, sub-atmospheric protection can be provided in the sense that the refrigerant pressure in all parts of the system, including the collector absorber, can be maintained above atmospheric pressure even when the equivalent temperature of the collector's surroundings falls below the refrigerant's boiling temperature at atmospheric pressure. D. GENERAL SPECIES OF THE INVENTION

1. S p e c i e s o f t h e Inv en t i o n

The various specific embodiments of the invention can be grouped into two general species (a) Forced Refrigerant-Circulation Species, and (b) Natural Refrigerant-Circulation Species.

Forced Refrigerant-Circulation species, or more briefly FRC species, employ at least one pump to circulate the refrigerant; whereas Natural Refrigerant-Circulation species, or more briefly NRC species, rely on the net static head resulting from the combined action of the heat absorbed from solar radiation and the local gravitational field to circulate the refrigerant.

Mostly, but not always, FRC species systems are preferred for large installations (say, with a total collector aperture greater than 100m²), and NRC species systems are preferred for small installations, (say, with a collector aperture below 25m²) .

There is no universal best mode for implementing the property of self-regulation described in Part III of this application. As will become apparent in Part V of this application, there are — depending on the application — four best modes or classes of systems for implementing this property with FRC species, and two best modes or classes of systems for implementing this property with NRC species.

After these six classes of systems have been discussed, techniques for over-pressure protection, combined over-pressure and over-temperature protection, freeze protection, and sub-atmospheric protection are discussed. Some of these techniques apply to all six classes and to all refrigerants, some apply to only some of these classes or to refrigerants with special properties.

IV. BRIEF DESCRIPTION OF DRAWINGS

Figure 1 illustrates in diagramatic form a typical embodiment of the invention.

Figure 2 shows details of Figure 1 between refrigerant inlet 9 of condenser 8 and outlet 16 of condenser pump 14 applicable to a several alternative embodiments of the invention.

Figure 3 illustrates a first alternative embodiment of the invention.

Figure 4 illustrates a second alternative embodiment of the invention.

Figure 5 illustrates a third alternative embodiment of the invention.

Figure 6 illustrates in diagramstic form a technique for obtaining an alternative signal for computing the desired effective capacity of condensate pump 14 in the case of the third alternative embodiment of the invention.

Figure 7 illustrates a fourth alternative form of the invention. Figure 8 illustrates a sixth alternative embodiment of the invention. Figure 9 illustrates in diagramatic form a technique for controlling the ratio of sensible heat to latent heat released in the case of the third alternative embodiment of the invention.

Figure 10 illustrates an alternative technique for controlling this ratio. Figure 11 illustrates in diagramatic form a technique of overpressure and over-temperature protection applicable to several embodiments of the invention.

Figure 12 illustrates in diagramatic form a technique for sub-atmospheric protection in the case of the first embodiment of the invention.

Figure 13 illustrates an alternative technique for freeze protection in the case of the first embodiment of the invention. V. B E S T M O D E F O R C A R R Y I N G O U T T H E I N V E N T I O N A. GENERAL REMARKS

As mentioned in Part III of this PCT application, the various specific embodiments of the invention can be grouped into two general species

(a) Forced Refrigerant-Circulation (FRC) species, and

(b) Natural Refrigerant-Circulation (NRC) species.

All FRC species must satisfy, for all external conditions under which a particular system is designed to operate, at least one additional internal operating condition; namely fifth, the available Net Positive Suction Head, or more briefly NPSH, available to a pump used in the system is maintained above the required NPSH for this pump. The best general mode or manner for implementing the five prerequisite conditions for achieving self-regulation in FRC species embodi ments depends on the particular solar heating application for which these embodiments are used. I distinguish and describe four general classes of FRC species embodiments. The preferred class depends on the application for which it is to be used.

By contrast, there is, with one exception, only one best general mode or manner for implementing the four ρrerequisite conditions for achieving self-regulaiton in NRC species embodiments. This exception occurs in the case where it is important that the enveloperof the space occupied by the refrigerant-circuit configuration used meets stringent constraints discussed later. I therefore distinguish and describe two general classes of NRC species embodiments, and describe the preferred basic refrigerant-circuit configuration for each of these two classes. B. DEFINITION AND DISCUSSION OF TERMS Certain terms used hereinafter shall have the following meaning for the purpose of this patent application:

(a) A "principal refrigerant circuit of an FRC species system is defined as the (closed) refrigerant circuit which includes

(1) the absorber inlet and outlet (2) the condenser refrigerant inlet and refrigerant outlet, and (3) the condensate pump inlet and outlet.

(b) A "principal refrigerant circuit" of an NRC species system is defined as the (closed) refrigerant circuit which includes (1) the absorber inlet and outlet, and

(2) the condenser refrigerant inlet and refrigerant outlet.

(c) The term "internal volume of the absorber" denotes the total internal volume of the refrigerant passages in the absorber in which the refrigerant absorbs heat directly from solar radiant energy or indirectly through the walls of these passages.

(d) The term "preselected limit", whether it refers to a lower or an upper limit does not necessarily mean that this limit has a fixed value. It may have a variable value determined by a function store in the CCU whose arguments are variables, such as evaporation rate and evaporation temperature, which are measured and supplied to the Central Control Unit (CCU) .

(e) The term "quiescent conditions" is used to refer to any set of conditions under which no radiant energy is being absorbed by the ref rigerant of a two-phase heat-transfer system, and all the, liquid- vapor interfaces of the refrigerant in this system are at the same temperature.

(f) The term "separator" is used to denote any device which can be used to separate wet vapor into its liquid and vapor phases and deliver dry vapor at one outlet and liquid at another outlet. The term

"separator" includes, in particular, a common header to a number of absorber refrigerant passages which has a large enough diameter for a horizontal liquid-vapor interface to form within this header, and from which nearly dry vapor can be supplied from an outlet connected to the vapor space within this header. The foregoing refrigerant passages may end

(1) at the surface of this header, or may

(2) enter this header and have their outlets above the (horizontal) liquid surface in this header. In the former case these same passages, if their diameter is large enough, may in some configurations be used to remove liquid by sewer flow. In the latter case a separate liquid outlet must be provided, below the liquid surface in the header, from which liquid accumulating in the header can be removed. C. FRC SPECIES

1. Gen e r a l Re mark s

The first three classes of FRC species were devised for applications where no requirement exists for controlling the ratio of the amount of sensible heat released by subcooling to the amount of latent heat released by condensation from the FRC system's refrigerant to the substance being heated. Among these three classes, the first two were devised primarily for applications where the evaporated refrigerant exiting the absorber is not required or desired to be superheated; whereas the third class was devised for applications where a small amount of superheat, which can be controlled within narrow limits, is desired.

The fourth class of FRC species was devised primarily for applications where a requirement exists for controlling the ratio of sensible heat released by the FRC system's refrigerant during subcooling to latent heat released by this refrigerant during condensation. Furthermore, this fourth class of FRC species may, where this ratio is not controlled, also be used either to control the quality of the refrigerant vapor exiting the absorber, or to ensure that quality one vapor is supplied to the condenser, over the system's entire operating range, even when a condensate pump with no capacity control is used.

Each of the foregoing four classes of FRC species may use a subcooler which is physically distinct from the condenser. Employing a separate subcooler in no way affects the basic operating principles of any one of these classes. (A separate subcooler is usually preferred where the ratio of sensible heat to latent heat released to the substance being heated is significant, say over 10%.)

In applications where the heat-exchange surface of the substance being heated is at a nearly spatially uniform temperature (say within ±10°C), no significant amount of subcooling (say exceeding 2°C) is desired or used. Examples of such applications are heating a substance to change its phase from a solid to a liquid, or from a liquid to a vapor; or heating a liquid or a gas, stored in a reservoir at a nearly uniform temperature, by circulating it through the FRC system's condenser and heating it in that condenser by raising the liquid or gas temperature by, typically, 5° to 15°C.

The principles of operations of the four basic classes of FRC species are described next for the case where the amount of subcooling is insignificant, and can therefore be neglected. The effect of subcooling on the operation of these systems is discussed later. 2. C l a s s A _{F N} S y s t e m s a. General Remarks. A Class A_FN system is usually preferred, to a Class A_FF system where:

(a) the desired quality of the refrigerant vapor exiting the absorber is high, say over 0,8 and need not be controlled precisely over the entire range of operating conditions, and where (b) the pressure drop through the absorber is small, say less than 10kPa. Class A_FN systems have the following distinctive features compared to the three classes of FRC species: they employ

(a) a natural-circulation, absorber, auxiliary refrigerant circuit, which is used primarily to prevent the refrigerant vapor supplied to the condenser from being wet even when the quality of the refriger ant vapor exiting the absorber is low; and

(b) a condensate pump used exclusively to pump condensed vapor, consequ ently the mass flow rate through the pump and the condenser are equal under steady-state conditions, b. General Operating Principles

The basic refrigerant-circuit configuration of a class A_FN system is shown in Figure 1.

Refrigerant, exclusively in its liquid phase, enters absorber 1 at 2, and vapor exits at 3, and enters separator 4 at 5. In this separator, the refrigerant evaporated in absorber 1 is separated from the non-evaporated refrigerant in a way which allows only dry refrigerant vapor to exit at vapor outlet 6, and only non-evaporated refrigerant to exit at liquid outlet 7. The dry refrigerant vapor exiting at 6 successively enters condenser 8 at 9, exits, mostly in condensed form at 10, and enters receiver 11 at 12. Only the condensed refrigerant exits receiver 11 at 13 and enters condensate pump 14 at 15. (The fact that no pump motor or pump engine is shown explicitly is not meant to imply that this motor or engine is necessarily an integral part of the pump assembly.) The refrigerant exits this pump at 16 and enters the refrigerant line connecting separator liquid outlet 7 to absorber inlet 2 at a mergence point 17 located between 7 and 2. I shall refer to this refrigerant circuit identified by 2-3-5-6-9-10-12-13-15-16-17-2 as the principal refrigerant circuit, and the refrigerant circuit identified by 2-3-5-7—17-2 as the natural—circulation, absorber, auxiliary refrigerant circuit,

A correctly configured (natural-circulation, absorber,) auxiliary refrigerant circuit can prevent the evaporated refrigerant exiting the absorber from-being superheated and the refrigerant vapor exiting the separator-vaporoutlet from being wet, thereby satisfying the first and second internal operating conditions. To this end, the volume of liquid refrigerant in this refrigerant circuit must be maintained within a lower and an upper limit for all values of the evaporation rate and the evaporation temperature over which the heat-transfer system is designed to operate. Each of these two limits varies with the evaporation rate and the evaporation temperature in a way which depends on the precise configuration of the auxiliary refrigerant-circuit. (The requirements which must be met by a correctly configured natural-circulation, absorber, auxiliary refrigerant circuit are discussed under the heading "Class A_NN Systems".) In addition to the volume of liquid refrigerant in the auxiliary refrigerant circuit being maintained within the foregoing lower and upper limits, the volume of liquid refrigerant in liquid line segment 9-15 should be maintained below an upper limit low enough for the third internal operating condition to be satisfied, and the volume of liquid refrigerant upstream of the condensate pump should be maintained above a lower limit high enough for the fifth internal operating to be satisfied. Each of this second, set of lower and upper limits also varies with the rate and the temperature of evaporation. It follows that the effective capacity of the condensate pump — to achieve self-regulation — must be controlled so that the volume of liquid refrigerant is maintained between the first set of limits in the auxiliary refrigerant circuit and between the second set of limits in the refrigerant circuit segment 9-15, namely between the condenser.refrigerant and the condensate pump inlet. A prerequisite for this to be possible is that the total volume of liquid refrigerant present in the refrigerant circuits of the heat-transfer system be

(a) large enough to satisfy the minimum liquid volume requirements of the auxiliary refrigerant circuit and the refrigerant-circuit segment 9-15,

(b) small enough not to exceed the maximum permissible liquid volume in both the auxiliary refrigerant circuit and the foregoing refriger ant-circuit segment, and

(c) sufficient to fill under all operating conditions the refrigerant- circuit segment between the inlet of the condensate pump and mergence point 17 exclusively with liquid refrigerant, Consequently, the refrigerant charge, or more precisely the amount of refrigerant mass m^ with which the heat-transfer system is charged should be chosen so that the corresponding liquid refrigerant volume satisfies all three of the foregoing volumetric conditions. c. Liquid-Level Control Technique with Fixed Limits Consider the case where the immediately preceeding three volumetric conditions can be met by maintaining the refrigerant liquid (surface) level in the refrigerant-circuit segment 9-15 within a fixed, preselected, lower limit LRI and a fixed, preselected, upper limit L_R2, neither of which need necessarily be in receiver 11, although the lower limit will visually always be in this receiver. The limit L_R1 is selec ted so that whenever the refrigerant liquid level in line segment 9-15 is above L_R1 , the available NPSH to the condensate pump is greater than the required NPSH by that pump; and limit L_R2 is selected so that whenever the refrigerant liquid level in this line segment is below L_{R2 ,} the liquid refrigerant which backs up in condenser 8 is small enough to maintain the condenser's effective condensing surface above a preselected, positive, upper limit. (Where deliberate subcooling in the condenser is not required, the limit L_R2 will be below the condenser refrigerant outlet 10.) In this case, the Central Control Unit, here- inafter referred to as the CCU, should control the effective capacity of the condensate pump so that the liquid refrigerant level is maintained within the foregoing two limits. (The numeral 18 indicates a typical location of the refrigerant liquid-surface level.) To this end, the CCU is supplied with a measure of the difference between the levels of limits L_R1 and L_R2 and a measure of the height of level 18, referred to as L_R with respect to a specified reference level, which may be one of the foregoing two limits. The measure of L_R used may, in general, be a continuous function, a multi-valued function, or in the limit a three-level function with one level indicating that L_R is below L_R1 , a second level indicating that L_R is between L_R1 and L_R2 , and a third level indicating L_R is above L_R2 The CCU, based on the value of the measure of L_R supplied to it, generates a signal that controls the effective capacity of the condensate pump in a way that tends during system activation,

(a) to maintain L_R between L_R1 and L_R2 if L_R is already between these two limits, and

(b) to return the level L_R between these two limits whenever, for any reason, L_R lies outside these limits,

(The liquid refrigerant surface level L& will not, in general, lie between L_R1 and L_R2 at the time the system is activated,) In general, proportional control of condensate pump capacity is preferred. To this end, a sensor must' be used from which the CCU can compute a signal which is a continuous function of the level L_R and increases with it, and the effective capacity of the condensate pump must be capable of being varied continuously.

The principles of operation of a class A_FN system will now be discussed in greater detail, without loss of generality, by assuming the sensor is a differential pressure transducer (effectively weighing a liquid column) and the effective capacity of the condensate pump is varied by changing its speed. Clearly, any other type of proportional sensor and any other means for varying the effective capacity of the condensate pump could be used to implement the same operating principles Details of the principal refrigerant circuit segment between points 9 and 16 in Figure 1 are shown in Figure 2, where the height of the refrigerant level 18 (referred to as L_R ) in receiver 11 is determined by using differential pressure transducer 19. In Figure 2, the same numbers indicate the same devices, points, or liquid level, as applicable, as those designated by these numbers in Figure 1, and the components of the system not shown in Figure 2 are identical to those in Figure 1. In the case of a through-type receiver, such as the one shown in Figure 1 or 2, this pressure transducer, for reasons explained later, is connected between outlet 13 of receiver 12, and a point 20 located at the lowest point of the circumference of the horizontal segment of line 10-12. Transducer 19 provides a signal L'_R , which represents a measure of L_R . This signal is supplied to the CCU as indicated by the symhol 5. The information represented by this signal is used in the CCU to generate a control N'_pc which increases the speed N'_pc of pump 14 as liquid level 18 rises, and conversely decreases the speed of pump 14 as this level falls. Symbol 5 indicates that the signal is supplied by the CCU. The particular criteria used to activate and de-activate the heat- transfer system depend on the particular substance to be heated, and the particular purpose for which it is to be heated. To keep the description of operating principles general, I shall refer to the control signal generated whenever these criteria are met as the activation signal. Similarly, I shall refer to the control signal generated whenever the de-activation criteria are met simply as the de-activation signal. (Activation and de-activation criteria include the following events: the temperature of the absorber, or of the refrigerant vapor at the absorber outlet rises above the temperature at a given point of the substance to be heated, by a preselected amount., and the temperature of the absorber, or of the refrigerant at this same outlet, falls below the temperature at the same point of the substance to be heated plus a smaller preselected amount.) The set of rules for controlling the condensate pump are: Condensate pump 14 is started whenever

(a) the activation signal is "on", and

(b) liquid level 18 in receiver 11 rises above level L_R1 by a presele ted amount; and it is stopped whenever

(a) the activation signal turns "off", or

(b) liquid level 18 falls below level L_RI.

The pump continues to run while the activation signal is "on" and liquid level 18 remains above level L_R1. While the heat-transfer system is operating, the speed of condensate pump 14 is, in effect, controlled by level 18 of the liquid refrigerant in the receiver. The pump runs at its maximum speed when liquid level 18 exceeds upper limit L_R2 , and stops when this level falls below L_R1, Between these two liquid levels, the pump, as mentioned earlier, runs at a speed which increases with increasing liquid level, and decreases with decreasing liquid level By my earlier assumption, as long as level 18 remains between L_R1 and L_R2,

(a) the auxiliary refrigerant circuit supplies, at the vapor outlet 6 of separator 4, quality one refrigerant vapor, (b) the effective condensing surface of condenser 8 remains above a preselected value, and (c) the available NPSH to the condensate pump is above the required NPSH by that pump. Thus, provided the vapor passages between the absorber outlet and the condenser refrigerant inlet are large enough for the fourth internal operating conditions to be satisfied, all five internal operating conditions required for self-regulation in FRC species systems are satis fied with the foregoing capacity control.

In any self-regulating two-phase, heat-transfer system, namely in any Rankine heat—transfer system the following relation always holds under steady—state conditions;

where m_re, m_rc, and m_rR, are the rates at which liquid refrigerant is being evaporated in the absorber, evaporated refrigerant is being condensed in the condenser, and condensed refrigerant is exiting the rec eiver, respectively. In an A—system, the condensate pump will return liquid refrigerant to the (natural-circulation, absorber) auxiliary refrigerant circuit at the same rate m at which condensed refrigerant exits the receiver, and therefore, using (1),

wher is the refrigerant mass flow rate delivered by the conden sate pump to the auxiliary refrigerant circuit, namely is the effective capacity of this pump. With the level control described earlier, where k_L is a factor which may be a non-linear function of (L_R - L_R1), and may be a function of other variables. For example, in the case here a differential pressure transducer is used to measure (L_R - L_R1), the signal supplied by the pressure transducer is proportional to (L_R - L_R1), where is the density of the liquid refrigerant,

Since is a function of the refrigerant saturated vapor temperature , the level L_R, varies with this temperature for a given value of Often this functional dependency may be acceptable, If not, it can be eliminated by the CCU if is measured and supplied to the

CCU, and the functional relatio between and , for the particu lar refrigerant used, is stored in the CCU, and the right-hand side of equation (3) multiplied by the inverse of this functional relation. The internal volumes of the separator and receiver can be chosen so that considerable lags in the value of in response to changes in the value of can be allowed without violating the conditions that must be satisfied for self-regulation, (For example, the lag for to reach 67% of its new steady-atate value after a step-function change in m can, with, economically practicable separator and receiver volume be greater than 10 seconds for a heat-transfer system whose collector aperture is 100m² and whose refrigerant is R—114 — namely dichloro- tetraflύoroethane — without violating the conditions for self-regula tion.) Furthermore, a bias error in instrumentation causes only the value of L_R, for which m_rp is equal to a given value of m_re , to change, and does not cause an error in the value of m_rpc. Also this change in level, for a reasonable bias error can be kept below a significant frac tion of the difference (L_R2-L_R1). Finally, scale errors in instrumen tation only affect the response time of changes in to changes in which — as already stated — are not critica It follows that the instrumentation for controlling the speed, and in general the effec tive capacity, of the condensate pump of an A_FN Class system need not be accurate. d. Liquid-Level Technique with Variable Lower Limits

In general, it is possible to configure the refrigerant circuits of a Class A_FN, system, and to select the refrigerant charge m_rT used, so that all three of the volumetric conditions designated by (a), (b), and (c), on pages 15 and 16 are satisfied. However, in some cases this may require a configuration which is not cost-effective, or is impracticable because of spatial constraints of a particular installation. These cases are most likely to occur where the system is required to operate over a wide range of evaporation rates and at low enough saturated vapor temperatures for the ratio of the specific vapor volume to the specific liquid volume of the refrigerant to be "large". A large ratio in the present context may be any number between 25 and 500 depending on a number of factors, and in particular on the--minimum permissible quality of- the refrigerant vapor exiting the absorber at the maximum evaporation rate. Considerations which affect the choice of minimum permissible vapor quality at the absorber outlet are

(a) the increase of condensate—pump flow power with decrease in vapor quality at this outlet,

(b) the decrease in refrigerant average evaporation (boiling) film heat-transfer coefficient in the absorber when this quality drops below a lower limit, which depends on a number of factors, including the mass flow rate per unit cross-sectional area and the particular refrigerant used. More specifically, the volume of liquid refrigerant in a given absorber, with a fixed internal volume, required to maintain the refrigerant vapor exiting the absorber in a given state — namely in a dry state with a fixed amount of superheat including zero, or in a wet state with a fixed quality — is a maximum at the lowest evaporation rate for which a Rankine heat-transfer system is designed to operate. To maintain the refrigerant exiting the absorber in the same state as the evaporation rate increases above this lowest evaporation rate, the volume of liquid refrigerant in the absorber must decrease, for a given increase in evaporation rate, by an amount which increases as the evaporation temperature, namely the saturated vapor temperature of the refrigerant in the absorber, decreases. The liquid volume thus displaced from the absorber, as the evaporation rate increases, must be accommodated in other parts of the auxiliary refrigerant circuit and, in particular, in the separator. If this is impracticable because, for example, the required separator volume to accomplish this is unacceptably large, the portion of the liquid volume displaced from the absorber that cannot be stored in the other parts of the auxiliary refrigerant circuit, without violating the second or fourth internal operating condition, or both must be accommodated outside the auxiliary refrigerant circuit. (The second internal operating condition may be violated because the. liquid level in the separator may be high enough for entrained liquid to exit the separator. The fourth internal operating condition may be violated because this liquid level is also high enough for the vapor space left between the separator inlet and the separator outlet to be small enough to cause a vapor pressure drop which results in an unacceptably large drop in saturated vapor pressure between the foregoing inlet and outlet.) The liquid-level control technique described earlier, does, however not allow the liquid refrigerant that cannot be accommodated in the auxriliary refrigerant circuit, without violating the conditions for self-regulation, to be stored in the receiver. Furthermore, the increase in liquid refrigerant present in the condenser, as the evaporation rate increases from its minimum to its maximum design value, is often small compared to the amount of liquid refrigerant that must be removed from the auxiliary refrigerant circuit to preserve self-regulation as the evaporation rate increases over this range.

Therefore, to preserve self-regulation, the size of the receiver is increased, and a modified condenser capacity-control technique is used, which allows the amount of liquid refrigerant that can be stored in the receiver to be changed; and, in particular, to be increased as the evaporation rate increases.

Briefly, this modified control technique consists of varying the effective capacity of the condensate pump in accordance with.the relation where the lower levelllimit is changed between L_R1 and (L_R1 + ΔL_R1) and where the value of ΔL_R1 is computed by the CCU as a function of the evaporation rate, or the evaporation temperature, or both. The effect of this modified control law is to cause the liquid level L_R, for a given condensate-pump effective capacity, to vary with operating conditions, and thus cause the amount of liquid refrigerant stored in the receiver also to vary with these conditions. Therefore, by appropriate choice of the function the change in the amount of liquid refrigerant stored in the receiver can be made to match the net change in the amount of liquidl-refrigerant that should be present in the auxiliary refrigerant circuit and the condenser to preserve self-regulation.

The details of the foregoing liquid-level control technique with variable lower limit are discussed under the heading "Details of Condensate-Pump, Liquid-Level, Control Technique". 3. C l a s s A_FF S y s t em s a. General Remarks

The fundamental difference between class A_FF systems and class A_FN systems is that the former employ a pump, which I shall refer to as an "overfeed pump", to produce the desired net static head in the absorber auxiliary refrigerant circuit, whereas the latter employ no such pump. In the refrigerant circuit configuration of an A_FF system, the separator can be located at any height in relation to the absorber or the condenser: whereas, in class A_FN systems, the height at which the separator cr can be located is constrained by the fact that this separator must be above the absorber inlet. b. General Operating Principles

The basic refrigerant circuit of class A_FF systems is shown in Figure 3. This configuration differs from the configuration shown in Figure 1 only by the fact that outlet 7 of separator 4 — instead of being connected directly to inlet 2 of absorber 1 — is connected to overfeed ipump.21 at inlet 22, and that outlet 23 of overfeed pump 21 is connected to refrigerant line 16-2 at a mergence point 24 located anywhere on this line between absorber inlet 2 and condensate pump outlet 16, The condensate pump is controlled by any one of the two liquid-level techniques discussed under the heading "Class A_FN Systems".

The effective capacity of the overfeed pump is — except in the unusual case where it is kept constant, or nearly constant, (by for example, using a positive, constant-displacement pump run at fixed speed) — is varied as a function of the effective capacity of the condensate pump. This function can be a continuous function or a multilevel function, The choice of function depends on the desired result. The effective capacity m_rpo of the overfeed pump can be controlled to keep the quality of the refrigerant vapor at the absorber outlet constant, within narrow limits, over a wide range of evaporation rates and evaporation temperatures. It can also be controlled to maximize the average boiling film heat-transfer co-efficient of the refrigerant in the absorber, by changing the value of the refrigerant vapor quality at the absorber outlet as a function of the evaporation rate, in cases where this film co-efficient is a sensitive function of this rate.

The particular technique used to control the effective capacity of the overfeed pump as a selected function of the effective capacity of the condensate pump depends greatly on the type of pump used; and also on other factors, including the tolerance within which the actual quality of refrigerant vapor at the absorber outlet is required to equal the desired quality at this outlet.

Consider, as one example, the case where both the condensate pump and the overfeed pump are identical

(a) positive displacement pumps whose volumetric capacity is to a high degree of accuracy only a function of their speeds, with

(b) electric motors whose speed is to a high degreee of accuracy, propor tional to the applied voltage.

(An example of a pump with these properties is a hydraulic diaphram pump with a permanent-magnet dc motor.) In this case, the desired quality of the refrigerant vapor at the absorber outlet, which is given by

can be obtained,, to a high degree of accuracy, merely by applying to the overfeed pump a dc voltage V_o (not shown) given by

where V_c (not shown) is the voltage applied to the condensate-pump motor.

Consider, as a second example, the case where the same pumps are used, but their motor speeds cannot be determined by the voltages applied to them. In this case, rotational speed transducers (not shown) could be used to determine their. speed,, and-the voltage applied to the overfeed pump motor could be controlled, to provide a desired refriger ant vapor quality at the absorber outlet, by adjusting this yoltage until the output signal of the rotational speed transducers of the two pumps indicated that the overfeed-pump was running at a speed given by

Consider, as a final example, the case where the effective capacities of the two pumps cannot be deduced from their speeds. In this case, liquid flow transducers 25 and 26 can be used to produce signals which represent measures and of the volumetric flow rates produced by the volumetric flow rates of the condensate pump and the overfeed pump, respectively. These signals are supplied to the CCU as indicated by the symbols 27 and 28. The CCU, as indicated by the symbol 29, supplies overfeed pump 21 with a control signal which controls its speed N so that the overfeed I ppuump tends to satisfy the relation

The quality may be a fixed number or a function of any variable —

' such as m_re or — stored in the CCU. A low level safety switch30 is used to provide the CCU with a signal L'_S1 which stops the over-feed pump if the level L_S of liquid refrigerant in the separator falls below minimum preselected level L_S1 which is chosen high enough to satisfy the overfeed pump's NPSH requirements. The symbol 31 indicates that the signal L'_S1 generated by this low level safety switch when L_S falls below L_S1 is available to the CCU to turn off the overfeed pump. 4. C l as s B S y s t erns a. General Remarks

The fundamental distinctive feature of Class B systems compared tό.:the other three classes of FRC species is that Class B systems have only one refrigerant circuit, and that the refrigerant vapor exiting the absorber must always be dry if the second internal operating condition is to be satisfied.

Class B systems are useful where a small amount of superheat in the refrigerant vapor exiting the absorber is desirable, and the pres-ur sure drop in the absorber is high. A small amount of superheat may, for example, be desirable to prevent condensation in the connecting tubing in installations where the condenser is located at a considerable distance from the absorber. A high pressure drop in the absorber can often occur where several concentrating collectors are connected in series. b.. General Operating Principles

The basic refrigerant—circuit configuration of a Class B system is shown in Figure 4.

Refrigerant, exclusively in the liquid phase, enters absorber 1 at 2. In a correctly operating system, only dry refrigerant vapor exits absorber 1 at 3. This dry vapor enters condenser 8 at 9 and exits, mostly in condensed form, at 10, and then enters receiver 11 at 12. Again, in a correctly operating system only condensed refrigerant exits receiver 11 at 13, and enters condensate pump 14 at 15. The refrigerant exits pump 14 at 16, and is returned to absorber 1 at 2. The absence of an absorber auxiliary refrigerant circuit has two important consequences:

(a) the liquid refrigerant displaced from the absorber, as the evaporation rate increases above Its minimum design value, must be accom modated entirely in the condenser-receiver segment of the principal refrigerant circuit, which in Class B systems is the only refrigerant circuit; and

(b) non-evaporated refrigerant exiting the absorber would -— in contrast to Class A_FN and Class A_FF systems — not only violate the second internal operating condition but would also cause any condensate-capacity control technique based exclusively on the level of the liquid surface in receiver 11 to malfunction. In theory, a condensate-pump, liquid-level control technique of the generalϊype described earlier could be made to control the effective capacity of the condensate pump correctly, but it is usually much simpler to use instead the control techniques described next. c. Superheat-Control Technique

This control technique consists in measuring the actual amount of superheat at some point between the absorber outlet and the condenser inlet and using a feedback control loop to control the effective conden sate pump capacity so that the actual amount of superheat tends toward the desired amount of superheat, The desired amount of superheat can be chosen to be small, perhaps as small as one degree Celsius, but cannotbe zero.

Any known method for measuring superheat may be used. One appropriate method for measuring superheat, when the Class B system operates over a wide range of saturation temperatures is described next.

A temperature transducer 32, preferrably immersed in the refrigerant, is used to measure the temperature of the refrigerant vapor at a point in line 3-9. The signal generated by this transducer is, as indicated by the symbol 33 supplied to the CCU. A pressure transducer 34 is located at the same point and used to measure the refrigerant pressure The signal generated by this transducer is also supplied to the CCU as indicated by the symbol 35. The functional relation for the particular refrigerant used is stored in the CCU, and from this relation, and from the signals and the CCU computes the super-heat The CCU compares this actual value of the super-heat with the desired value stored in the CCU and supplies, as indicated by symbol 51, the condensate pump with a control signal N'_p which tends to annul the difference

This condensate-pump control technique is self-starting because, before system activation, the value of is determined by the temp erature of the substance to be heated, and the system would not be activated — whatever the activation criterion or criteria used — unless

A low-level, safety control switch 37 is used to provide the CCU with a signal L'_R1 which stops the condensate pump if the level L_R of liquid refrigerant in the receiver falls below a minimum preselected level L_R1 , which is chosen high enough to satisfy the condensate pump's NPSH requirements. The symbol 38 Indicates that the sifnal L_R1, generated by this low-level safety switch Is supplied to the CCU.

5. C l a s s C_FN S y s t e m s a. General Remarks

Class C_FN systems fall into two general sub-classes: (a) Class C_FN systems with no independent subcooling control, and (b) Class C_FN systems with independent subcooling control.

The former sub-class, in turn, uses two condensate-pump capacity control techniques (a,l) fixed-capacity control, and (a,2) evaporation-rate capacity control, The key distinctive feature of all Class C_FN, systems, compared to the other three classes of FRC species systems, is that the condensate pump not only returns condensed refrigerant vapor to the absorber, but also liquid refrigerant which was not evaporated in the absorber, b. Refrigerant-Circuit Configuration with no Receiver The basic refrigerant-circuit configuration of a class C_FN system with no receiver is shown in Figure 5.

Refrigerant, exclusively in the liquid phase, enters absorber 1 at 2; and, in general, wet refrigerant exits at 3 and enters separator 4 at 5. In this separator which also functions as a receiver for excess liquid refrigerant, the evaporated and non-evaporated refrigerant are separated in a way which allows only dry refrigerant vapor to exit at. vapor outlet 6, and only non-evaporated refrigerant to exit at liquid outlet 7 The dry refrigerant vapor exiting at 6 enters condenser 8 at refrigerant inlet 9; exits mostly In the condensed state at 10, and merges, with the non-evaporated refrigerant exiting separator 4 at liquid outlet 7, at point 45. The condensed refrigerant from condenser 8 and the non-evaporated liquid refrigerant from separator 4 are supplied to condensate pump 14 at inlet 15, and returned by this pump through outlet 16 to absorber inlet 2. In this configuration the refrigerant outlet 10 of condenser 8 must be at a high enough level above the free liquid surface 40 for the static head between these two levels to be greater or at least equal to the pressure drop in the condenser at the maximum design evaporation rate. c. Fixed-Capacity Control The system charge m is selected so that — at the minimum design evaporation rate, and at the evaporation temperature, within the design range, for which total refrigerant liquid volume is a minimum — liquid level 40 in separator 4 is above the level L_S1 at which the low-level safety switch 30 supplies the CCU, as indicated by symbol 31, with turn-off signal L'_S1

The effective capacity of the condensate p.imp is chosen just large enough to prevent the amount of superheat of the refrigerant vapor exiting absorber 1 at 3 from exceeding a preselected upper limit at the maximum design evaporation rate, over the entire range of evaporation temperatures for which the system is designed to operate.

After system activation, using any desired activation criteria, the pump operates at constant volumetric capacity and the liquid over- feed flow rate is equal to the pump mass flow rate less the current evaporation mass rate. Therefore, the quality of the vapor at the outlet of the absorber varies with liquid density and evaporation rate mre and is given by where m_rp is the mass flow rate of liquid refrigerant from condensate pump outlet 16 to absorber inlet 2, and is the volumetric pump capacity, (volumetric flow rate).

At low evaporation rates, the quality of the vapor at the absorber outlet will be low and the film heat-transfer co-efficient in the absorber may be sub-optimum. This inefficiency in heat transfer and the inefficiency in pump power at low evaporation rates make fixed-capacity control attractive only for small low-cost systems, d. Variable-Capacity Control A more efficient implementation of the system is to vary the volumetric flow r of the condensate pump, and hence also its effective capacity o give a controlled amount of liquid overfeed in the absorber. The overfeed rate may be a selected function of the evaporation rate. In particular, the overfeed rate can be made a linear function of the evaporation rate, thereby obtaining a desired constant quality at the absorber outlet which is independent of the evapo ration rate.

Because the subcooling in the condenser will be small, both the condensed refrigerant, exiting the condenser at 10, and the non-evaporated liquid refrigerant, exiting the separator at 7, will have almost the same density Therefore, the mass flow rates m_rc and m_ro , of the condensed and overfeed liquid refrigerant may be measured, with sufficient accuracy, by flow-rate transducers 41 and 42 in the liquid 30 lines 10-45 and 7-45, respectively. The signals F' and F' generate by these transducers are supplied to the CCU as indicated by symbols

43 and 44, respectively. In the CC se signals are multiplied by . for the saturated temperature at which the system is opera- ting.

The vapor quality q_rA,o is given by

in the steady state, because the condensation rate m is equal to the evaporation rate m_re, an d because the total flow rate (m_re + m_ro) equal the punp mass flow rate . Therefore, if the refriger ant circulation pump is of a type for wKich the volumetric capacity can be easily predicted by the CCU, for example, a positive displacement pump with speed control, the pump can be controlled by a signal from the CCU so that

if is the desired quality, or so that if is the desired overfeed. Note that in this latter case the overfeed flow-rate tramsducer 42 is not required. However, if the volumetri capacity of the pump can not be easily predicted, a preferable control technique is to include transducer 42 in ths system and use closed-loop control of the condensate pump so that the actual capacity

F tends to the desired value hich is either , or ro , for control of vapor quality or contro of overfeed flow rate, respectively.

In all cases, the refrigerant pump is de-activated by the CCU in response to a signal L'_S1, from the low-level safety switch 31 if refrig erant level 40 in the separator falls below level L_S1. In a Class C_FN. system, this would occur only because the refrigerant charge used is too small.

A simplified method for controlling the circulating pump capacity is to use stepped control of capacity, for exampl&stepped speed drive or stepped volumetric capacity, in which the effective capacity of the pump is always sufficiently greater than the evaporation rate for the vapor quality to be less than one at the absorber outlet. For this cruder control method, the accuracy of the flow transducers 41 and 42 may be relaxed, so that less expensive transducers may be used. Alternatively, with a condensate pump whose capacity steps are predictable, the flow transducers may be replaced by transducers which give more indirect measurements of evaporation rate. One set of alternative transducers are shown in Figure 6: namely a differential, pressure transducer 47 and a temperature transducer 6 . The pressure transducer is connected to the refrigerant Inlet 9 and refrigerant outlet 10 and therefore produces a signal Δp'rc , which represents the refrigerant pressure drop in the condenser. The temperature transducer produces a signal , which represents the saturated vapor temperature of the refrigerant. These two signals are supplied to the CCU dicated by symbols 49 and 40. From these signals, the CCU computes m as a func r^e tion of Δp_rc and , obtained from tests and stored in the CCU. The effective capacity of the condensate pump can be controlled continuously or in steps — which in the limit may only consist of two steps — by a signal supplied by the CCU whose nature depends on the type of condensat pump used. This signal designated by in Figure 6 is, as indicated. by symbol 3 , supplied by the CCU.

D. NRC SPECIES 1 « C l a s s A_NN S y s t e m s a. General Remarks

The operating principles of FRC species systems were discussed with out specifying the configuration of the absorber and its refrigerant passages. By contrast, the operating principles of Class A_NN systems are discussed in detail for the general case where the absorber axis or or plane, as applicable, is "tilted," namely makes a finite (non-zero) angle with a local horizontal plane. The limiting case where the absorber is horizontal is discussed briefly after the operating principles of Class A_FN systems have been discussed- for the general case. The absorber passages of NRC species systems as well as FRC species systems may have any configuration that does not cause a vapor lock. For example, in the case of line-focus collectors, the absorber passages may consist of a single cylindrical tube, an annulus or segments of an annulus between two concentric tubes; In the case of a point-focus col lector, the absorber passages may consist of a tapered spiral tube; and in the case of flat-plate collectors, the absorber passages may consist of a "waffle" pattern of the type produced, for example, by the Olin Corporation ROLL BONDR process.

The absorber fluid passages, belonging to one or more physically distinct collector modules, may be connected in series or In parallel, or both, but are interconnected by one or more headers or manifolds, so that liquid refrigerant enters at a single point, referred to as the absorber inlet. Similarly, the absorber fluid passages are also interconnected so that the refrigerant exits at a single point, referred to as the absorber outlet, or interconnected by one or more separator modules so that the refrigerant from all the absorber fluid passages exits in one or more separate modules, referred to collectively as the separator, The vapor outlets of the separator modules are interconnected, in turn, to merge at a single point referred to as the separator vapor outlet; and the liquid outlets of these modules are interconnected to merge at a second single point referred to as the separator liquid outlet. The absorber outlet of tilted absorbers, namely absorbers whose axis or plane is not horizontal, must be above the absorber inlet in NRC species systems, as well as FRC species Class A_FN systems. In the case where there is no single absorber outlet, because the refrigerant exits directly into one or more separator modules, both the inlet and the vapor outlet of each separator module must be above the absorber inlet. This constraint on the relative levellof the absorber inlet on the one hand and the absorber outlet — or the separator module inlets and vapor outlets where a separator is used — on the. other hand, does not apply to FRC species Class A_FF, Class B_F, and Class C_FN systems. b. General Operating Principles

Th e usually preferred basic refrigerant-circuit configuration of a Class A_NN system is shown in Figure 7 . In this configuration the separator also fulfills the function of a receiver.

By contrast, in Class A_FN, Class A_FF, and Class B_F systems, a receiver is usually employed in addition to a separator.

Refrigerant, exclusively in its liquid phase, enters absorber 1 at 2. Dry or wet refrigerant vapor exits at 3, and enters separator 4 at 5. In this separator, the refrigerant evaporated in absorber 1 is sepa rated from the non-evaporated refrigerant in a way which allows only dry refrigerant vapor to exit at vapor outlet 6, and only non-evaporated refrigerant to exit at liquid outlet 7. The dry refrigerant exiting at 6 successively enters condenser 8 at 9, exits mostly in condensed form at 10, and enters the refrigerant line connecting separator liquid outlet 7 to absorber inlet 2 at a point of mergence 39 located between 7 and 2. I shall refer to the refrigerant circuit identified by 2-3-5-6-9-10-39-2 as the principal refrigerant circuit; and to the refrigerant circuit identified by 2-3-5-7-39-2 as the absorber auxiliary refrigerant circuit. Note that this auxiliary circuit is equivalent to the natural-circulat±on absorber auxiliary refrigerant circuit used in FRC species systems.

Before discussing the operating principles of Class A_NN, systems, I note that these principles are unaffected by the location of the point at which liquid line 10-39 is connected to liquid line 7-2, provided this point is not significantly below absorber inlet 2. Furthermore, liquid line 10-3 could have also been connected directly to separator 4 without affecting these operating principles. (This is also incidentally true for the natural-circulation, absorber auxiliary refrigerant circuits of Class A_FN systems.) However, the explanation of these principles using a refrigerant-circuit configuration in which the mergence point 39 is, as in Figure 7, , at the same level as the absorber inlet.2, may be easier to follow.

In understanding the basic spatial relationships, Figure 7is intendedto show qualitatively the relative vertical distances between outlet 3 of absorber 1, and the center-lines (not shown) of separator 4 and condenser 8. The line L_o-L_o represents the level of the liquid refrigerant surfaces in absorber 1, and in the liquid refrigerant lines between 7 and 39 and 10 and 11, when the system is quiescent and the liquid refrigerant is at a uniform temperature. As shown, separator 4 and condenser 8 are located at a level above that of level L_o, with the condenser outlet 10 higher than the separator outlet 6. If ports 5 and 6 of the separator or port 10 of the condenser were below level L_o, the separator completely, and the condenser partially, would be filled with liquid and could not perform their functions.

Considering the case where radiant heat is absorbed in the absorber, causing a refrigerant evaporation rate (m_re)₁, at a rate of (Q_A)₁ the amount of liquid refrigerant in the collector decreases because the absorber volume is partially filled with saturated vapor (of lower density) and the level of the liquid refrigerant in the lines 7-39 and 10- 39 rises to accomodate the liquid displaced from the absorber. When steady-state conditions have been reached, the mean absorber L_A1 liquid level (averaged over the vertical extent of the absorber) will settle at, say, level L_A1 and the level of the liquid surface inj line 7-11 at a level L_S1 such that the net static head (L_S1-L_A1) balances the friction induced pressure drop by the flow of liquid and vapor refrigerant in the auxiliary refrigerant circuit 3-5-7-39-2-3. While the system is operating in a steady state; the pressure at points 5 and 9 are essentially equal, and the liquid level in the refrigerant line 10-39 will settle at a level L_C1 such that the net static head (L_C1-L_S1) balances the friction-induced refrigerant-flow, pressure drops in the circuit 6- 9-10-39-7-6 which are mainly due to the pressure drop in the condenser 8. Similarly, if the radiant heat is absorbed by the absorber at a higher rate (Q_A)₂, causing an increased evaporation rate m_re2, the mean liquid level in the absorber falls further to say L_A2 and the liquid levels in lines 7-39- and 10-39 rise higher to the levels L_S2 and L_C2 respectively, to balance the increased flow pressure drops and to accom odate the increase in liquid displaced from the absorber.

Thus, the magnitude of the static head (L_S2-L_A2) , required to offset the pressure drop around the auxiliary circuit 5-7-39-2-3-5 for a given flow rate, is a- function of the (friction-induced) pressure drop in the absorber and increases with it. Hence, the minimum height of the separator vapor outlet 6 above outlet 3 of the absorber necessary to allow the required static head (L_s2-L_A2) to be provided — at the maximum design heat-absorption rate (Q_A) max — without flooding the separator, is decreased by using refrigerants and absorbers with smaller flow pressure drops. Similarly, the minimum height of the condenser liquid outlet 10 above the vapor outlet 6 of the separator, necessary to allow the required static head (L_C2-L_S2) to be provided — at maximim heat absorption rate — without the liquid backing up into the condenser outlet, is decreased by using condensers with smaller pressure drops.

It follows that self—regulation of the system over the design range of evaporation temperatures and evaporation rates is determined by the relative levels of the absorber inlet and outlet, the separator inlet, vapor outlet, and liquid outlet, and the condenser refrigerant inlet and outlet; the configuration of the refrigerant passageways, and the correct choice of refrigerant charge. For a particular system configuration and refrigerant,-the correct charge is determined by the refrigerant liquid levels L_A1, L_S1and L_C1 , for a heat-jtransfer fate Q_A1 equal to the minimum design operating rate, when the average temperature of the liquid refrigerant is equal to that operating temperature,^within the design range, which corresponds to a minimum liquid volume and a maximum volume of vapor, within the system for a given charge mass, (This operating temperature is determined analytically from the internal dimensions of the systemscomponents and the published tables of specific volume or density for the liquid and saturated vapor states of the particular refrigerant.) The liquid level L_A1 is that giving the liquid volume required in the ab absorber for the vapor at the absorber outlet to be in a given state, for example, at quality one. The liquid levels L_S1 and L_C1 are determined relative to L_A1 by the total pressure ilrops around the auxiliary and principal refrigerant circuits at this operating temperature and minimum heat transfer rate.

In practice, the correct charge (once determined) is inserted over a particular environmental temperature range, with the system quiescent at that temperature, and may be checked by the location offthe quiescent liquid surface line L_o - L_o .

At other evaporation rates or other temperatures in the design operating range, the volume of liquid in the system will be greater so that the mean liquid levels L_A1 L_S1 and L_C1 will, in general, be higher than before, though by differing amounts because of the different vapor and liquid densities at these temperatures, with the result that the vapor qualityώexiting the absorber at 3 will be less than before and thus the first internal operating condition will continue to be met. To minimize the variation of quality of the vapor exiting the absorber at minimum Q over the design temperature range, and to reduce the overall system height, it is convenient for the variation in liquid volume to be accommodated in the separator 4. This is achieved by permitting level L_S1 to enter the separator, and to configure the cross-sectional area of the region traversed by L_S1 so that,tas the refrigerant density and absorber pressure drop changes with temperature, the mean volume of liquid in the absorber remains near that required for the selected vapor quality(without superheat) at the absorber outlet. Depending on the configuration and the specific variation of refrigerant density with temperature, this may be possible for Increasing temperature, or for decreasing temperature relative e temperature for minimum liquid volume, but not for both. Let be the saturated vapor temperature at which the total liquid refrigerant volume is a minimum for a given refrigerant charge. Also let (L_o)min be the minimum liquid level in the absorber under quiescent conditions at this temperature for which the vapor quality exiting the absorber at the minimum design evaporation rate is one. It then follows from the preceeding discussion that at all other evaporation rates and temperatures, within the design range, the vapor exiting the absorber will not be superheated. The minimum height at which the liquid outlet of the separator can be located and fulfill its additional function of a receiver, without having excess liquid storage function, is the height determined by the level (L_o)min. The liquid volume which must be stored outside the absorber and the condenser is a function of the evaporation rate and the evaporation temperature. Let the maximum value of this volume be V_max, and the corresponding liquid level in the separator above (L_o)_max be L_s. This level can, in principle, be made arbitrarily small by choosing the horizontal cross-section of the separator arbitrarily large. And the static head obtained with this level can be sufficient to cause a high enough flow rate in the absorber at all evaporation rates and tempera tures for no superheat to appear at the absorber exit by choosing the cross-sectional area of the absorber refrigerant passages large enough.

The minimum height of the refrigerant outlet of the condenser above L_S for which the third internal.operating condition_. is satisfied is determined by„the .refrigerant pressure drop in the condenser. This pressure drop is a maximum for the highest design evaporation rate and the lowest design evaporation temperature. This minimum height of the condenser refrigerant outlet can again, in principle, by made arbitrarily small by vising a condenser with an arbitrarily low maximum refrigerant pressure drop.

It follows that the separator liquid outlet and condenser liquid outlet can be located at arbitrarily small levels above the level (L_o) max by selecting the large enough cross-sectional areas for the absorber and condenser refrigerant passages, and a large enough volume for the separator. Note that if the separator outlet of a given refrigerant circuit configuration is at the level (L_o)min at a given tilt angle, it will remain above this level for any larger tilt angle. Consequently, a given Class A_NN configuration can be used over a wide range of tilt angles . 2. C 1 a s s B_N S y s t e m s a. General Remarks

Class B_N systems differ from prior-art two-phase natural-circulation systems by the fact that they (a) use a receiver to accomodate increases in specific liquid refrig erant volume with temperature, and to prevent liquid refrigerant displaced from the absorber at high evaporation rates from backing up into the condenser through its refrigerant outlet, and may (b) use a receiver with a horizontal cross-section which changes as a function of height to help control variations in the state of the refrigerant vapor exiting the absorber with changing evaporation rate and temperature, b. General Operating Principles

The general operating principles of a Class B„ system are clearl the same as those of refrigerant circuit 2-3-5-7-9-10-37-2 of a Class A^ system, except that the principal and only refrigerant circuit of a Class B system has a receiver 11, as shown in Figure 8.

The principal advantage of a Class B system, compared to a Class A_NN system, arises from the fact that the length of the liquid line 10-12 can be zero, and that the vertical distance between the surface of the liquid refrigerant in the receiver and the condenser refrigerant outlet can also be zero without imposing any constraints on minimum permissable refrigerant pressure drop in the condenser; whereas a zero vertical distance between the surface of liquid refrigerant in the separator of a Class A_NN system and the refrigerant outlet of its condenser requires the refrigerant pressure drop in this condenser to be zero if liquid back-up in this receiver is to be prevented.

E. DETAILS OF CONDENSATE-PUMP, LIQUID-LEVEL CAPACITY-CONTROL TECHNIQUE 1. T y p e s o f P u m p - C a p a c i t y C o n t r o l

The objective of condensate-pump, capacity control is to ensure that the five internal operating conditions are satisfied under all external conditions in the design range. As discussed earlier, this objective is achieved by selecting the proper refrigerant charge for the system, imposing certain restraints on the system configuration, and by control ing the volume of liquid refrigerant present in the absorber and the absorber auxiliary refrigerant circuit. The most appropriate type, of FRC species condensate-pump and overfeed pump capacity control depends on system size and on the refrigerant used.

Four types of capacity control are: (1) pump speed control, (2) pump-capacity variation at fixed speed, (3) pump by-Pass or throttle control at fixed speed, and (4) pump on-off control. The fourth type of control is generally the least expensive, bμt is usually not desirable for many reasons, including the resulting frequent start and stop transients which affect adversly the reliability of the pump and its driving mechanism. The third type of capacity control is also usually undesirable because of a general reduction in pumping efficiency, but may be attrac tive in some small systems. The second type of control is mostly appli cable to large systems and may be combined with the first type of control to obtain continuous capacity control, . The first type of control is widely used and without loss of generality,will be assumed in describing in greater detail the basic principles of liquid-level condensatepump, capacity-control techniques. Clearly, these basic principles apply to any other type of pump-capacity control.

Variation of pump speed requires a variable speed drive. It is not relevant to the control principles described below whether the variation of speed is obtained by an engine, a variable speed ac or dc motor, a stepping motor, or a variable speed hydraulic or mechanical drive with a constant speed power source. Each method of speed variation has advantages for particular applications and designs of heat-transfer systems, but the basic control principles described next are the same for all methods of speed control. 2. D e t a i l s o f B a s i c P r i n c i p l e s

There are two main types of (Liquid)pumps,.. The first is a positive displacement type in which the volume of fluid displaced per revolution (or cycle) of the pump is relatively independent of pump speed and the hydraulic head or pressure rise generated. The second type is a centrifugal pump, in which the volume of fluid displaced per cycle may be very dependent on the pump speed and the hydraulic head to be overcome. The control principles are the same for both types of pumps, but the volu metric flow fxinction of the centrifugal-type pump must be compensated by the CCU if approximately linear control of volumetric capacity is desired. If a positive displacement pump is used, compensation of the flow function can usually be omitted.

Referring to Figure 2, the rotational speed of the condensate pump is given by the CCU. where Pp. is the desired pump volumetric flow rate (m³/sec and is the pump volumetric function in terms of speed N the pressure rise (Pascal) and the type and temperature of the liquid refrigerant being used. In particular, the pump volumetric flow function is given by

where ηνis the volumetrie efficiency and is the pump displacement volume to sufficient accuracy for most system applications, employing positive displacement pumps. Therefore,

In the steady st ate, the desired volumetric pump flow rate is equal to the condensate flow rate s the refrigerant liquid density.

The condensate flow rate may be measured directly by using a flow transducer, but this is not required for controlling the refrigerant pump capacity. The use of a sensor to determine the refrig erant liquid level 18 in the receiver 11, between the condenser and the condensate pump, is sufficient to control the pump volumetric rate. Figure 2 shows that a differential pressure transducer 19 is used to det ermine the liquid level 18, essentially by weighing two liquid columns, but other types of sensor could obviously be substituted. Note that pressure transducer 19 is connected between outlet 13 cr just below outlet 13, of receiver 11 and a point 20, preferrably In a horizontal leg of refrigerant line 10-12. Point 20 is located at the lowest point of the circumference of the horizontal leg of line 10-12 to ensure that the line 19-20 is kept full of liquid refrigerant under all operating and non-operating conditions, and thus assure that pressure transducer 19 provides at all times a reliable measure of the difference in weights, and therefore also of the heights of the liquid columns in line 19-20 above transducer 19 and the liquid column between the level of the liquid refrigerant surface 18 in receiver 11 and the level of transducer 19 (If the lower pressure side of transducer 19 were connected to the vapor space in receiver 11 — and thus neglecting vapor weight — were used to measure directly the weight of the liquid column between surface 18 and the level of transducer 19, a heating element would have been needed to stop vapor condensing in the line connecting the lower-pressure side of transducer 19 to the vapor space in receiver 11. And clearly, liquid refrigerant in this line would cause transducer 19 to produce an erroneous signal.)

wher is a transducer scale factor, L_R is the liquid surface height and L _R1 is a reference height, and is the liquid refriger ant density. The CCU controls the volumetric capacity of the condensate pump to be the sum of a term proportional to signal S_L and a bias signa S _B . Therefore,

or

The liquid level L_Rin the receiver is the result of the refrigerant flow into the receiver from the condenser and out of the receiver to the pump. Therefore, the derivative L _Rof the liquid level is given by

where A_Ris the cross section area of the receiver. When is controlled by equation 26 this gives a closed loop level control which is a stable first-order control system and has a steady-state solution the desired control law for the condensate pump volumetric capacity. The steady-state liquid level L _Ris given by or, and the corresponding pump speed N_p is given by The response of the system to a step transient, such as a cloud partially obscuring the solar radiation, is exponential with a response time-constant Is given by

The time constant 1 may be made Independent-of the liquid--temperature by compensating the control signal S_L for changes in liquid density as a function of temperature. That is, the CCU controls the proportional constant to be giving

and

From (31) we get and comparing (39) with (4) and remembering that in the steady-state

we obtain

and The bias signal S_B is generated in the CCU to control the steady- state liquid level L_R so that, for a given system spatial configuration and a selected charge of a particular refrigerant, the volume of liquid refrigerant in the absorber and the absorber auxiliary refrigerant cir cuit of Class A_FN and Class A_FF systems can be maintained at the optimum value over the design range of saturated vapor temperatures and evaporation rates. The changes in liquid density and vap nsity are compensated, for a particular system, by generating as a polyno mial function of in the CCU, which is provided with a measure of the vapor temperature , or equivalently the refrigerant liquid temperature when sub-cooling is small. Similarly, the bias signal S_Bas a function of the evaporation rate m _re can also be deter mined by the CCU If It is supplied with any acceptably accurate measure of m _re and the functional dependence of SB on m _re is also expressed as a polynomial (determined by earlier tests on the heat-transfer system) F. OPERATION OF FRC SPECIES AND NRC SPECIES SYSTEMS WITH SUBCOOLING 1. S ub co o l in g w i th no I n d ep en den t Co n t ro l

FRC and NRC species systems may be operated either with essentially zero, or with substantial, subcooling of the liquid refrigerant. A system may be operated with a moderate degree of subcooling by designing the condenser to hold a sufficient amount of condensed vapor within it and by increasing the system's refrigerant charge to supply this additional liquid volume. If a large amount of subcooling is required, it is preferable to Insert in FRC species systems. a subcooler in the refrigerant line between the condensate pump outlet 16 and the mergence point 17 or 24 in Class A_FN and Class A_FF systems, respectively, and absorber inlet 2 in Class B_F, or Class C_FN systems. In a Class A_NN system the subcooler would be inserted in the line between mergence point 39 and absorber inlet 2; and in a Class B_N system, between receiver outlet 13 and absorber inlet 2.

In FRC species Class A_FN, Class A_FF, and Class B_F systems, the mass flow rate of refrigerant through the subcooler,- if one is used, is the same as the mass flow rate through t he condenser (in the steady state). Therefore, the amount of subcooling Δ_sbTr , obtained with these three systems, is determined by the sensible heat delivered to the substance being heated, and more specifically by where Q _sb is the sensible heat transfer rate, is the specific heat of the liquid re , and m _r is the refrigerant mass flow rate. If the product for the refrigerant exceeds the specific heat, mass flow-rate product for the substance being heated,- the subcooling temperature dr op across the subcooler is less than the temperature rise of the substance being heated in the subcooler and sufficient sensible heat transfer capacity is available in the foregoing three classes of systems. The sensible heat transfer is not eontrolled independently during the operation of these three classes of systems, but- adjusts itself automatically to the sensible heat transfer load as long as sufficient sensible heat transfer capacity is available.

If the sensible heat transfer capacity through subcooling is not sufficient in a Class A_FN, Class A_FF, or Class B_F , system, a Class C_NN system should be used.

2. S ub c o o l i n g wi th I n d e p en d en t C o n t ro l

A Class C_FN system should be used for liquid refrigerant subcooling under conditions when the liquid subcooling capacity in a Class A_FN Class A_FF, or Class B_F system would be Inadequate; namely in cases where the sensible heat transfer load exceeds the sensible heat transfer capacity of the liquid refrigerant when the mass flow rate of liquid refrigerant through the subcooler~is equal to m_rc. A Class C_FN system provides a greater sensible-heat transfer capacity than the other three classes of FRC systems because the liquid refrigerant mass flow rate through the pump and the subcooler may be arbitrarily larger than the condensation rat

In this application, where the sensible heating load may vary, it is preferable to control the effective capacity of the condensate pump by a signal which is a function of the amount of subcooling achieved (or conversely, the amount of sensible heating or temperature rise of the substance being heated). The control technique for a Class C system is shown in Figure 9. Temperature sensors 52 and 53, measure the temperatures and at the refrigerant outlet 54 and fluid (sub stance to be heated) inlet 55 of subcooler 56, located between condensate pump outlet 16 and absorber inlet 2 and supply signals and to the CCU as denoted by symbols 57 and 58.

Alternatively, the tenφerature sensors 59 and 60 may be positioned as shown in Figure 10 at refrigerant inlet 61 and at the substance being heated outlet 62 of the subcooler 56, delivering temperature signals and _, ^as shown by symbols 63 and 64 to the CCU respectiv ely. In either case the capacity of pump 14 would be controlled by the CCU according to the deviation of the differences measured by the sub- cooler temperature sensors from a pred etermined value ΔT, which may be a function of condensation temperatur . That is, the CCU would give a closed loop capacity control by commanding a pump flow rate equal to

where G. is a fixed control gain constant, for the first or alternative arrangement, respectively.

G. COMPUTATION OF THE EVAPORATION RATE BY ESTIMATING THE REFRIGERANT

HEAT-ABSORPTION RATE

The stagnation temperature, namely the no-load radiation equilibrium temperature , of some types of collectors — such as flat-plate collectors — can be replicated sufficiently accurately by a small dummy collector module. In the case of a flat-plate collector, the dummy collector module uses the same transparent cover or covers, absorber panel, and frequency-selective or non-selective absorber coating, as the operating collector modules. However, the dummy collector has no refrigerant charge. The absorber panel is however insulated from the surrounding air and supporting structure in a manner which replicates the rate at which the absorber panel of the operating collector modules lose heat to their surroundings. The dimensions of the dummy collector module must be large enough for so-called end effects to be negligible. A typical dummy collector module for a flat-plate collector would have an area of about 0.1m². A temperature transducer is placed in thermal contact with the dummy collector-module absorber panel. The tilt angle of this panel is the same as that of the panels in the operating collector modules. I consider here only the case where the amount of superheat of the refrigerant in the absorber — which in this case consists of all the absorber panels of the flat-plate collector array — is negligible. In this case, the rate at which heat is absorbed by the refrigerant in the absorber is given by where is the evaporation temperature, namely the saturation temperature of the refrigerant "in the absorber, and where Ke Is a factor which can be determined by calibration tests, and which remains fairly constant over a wide range of evaporation temperatures, Also In this case the evaporation rate m _reis given by

so that where the latent heat of evaporation rate is tabulated for a given refrigerant as a function of the evaporatio n temperature. This fxinction can be stored in the CCU so that the CCU can compute m _reas a function of

A more accurate estimate of m _recan be obtained, particularly. whenever the range of the values of the operating temperature is ke as a function, determined by tests, of and , where is the ambient temperature of the air in collector's out door surroundings. Alternatively, Q_rA can be expressed in the form

where and are linear or quadratic functions of and respectively. In cases of collectors with a high heat loss rate to th urroundings (which are usually sensitive to wind speed) , a term — where is a factor determined by calibration and μ _w Is an appropriate component of the wind speed in the vicinity of the collector aperture — can be added to the right-hand side of equation (47) to take into account the effect of wind speed on Q_rA.

H. P RO T E CT I O N TE CHN I Q UE S

1. O ve r -P r e s s u r e P r o t e c t i o n

All two-phase heat transfer systems are designed with a maximum pressure limit which is safely above the maximum pressure experienced during normal operation. However, over-pressure protection may be required when the system is de-activated by error, power supply failure, certain equipment failures, or because the CCU de-activates the system in response to an indication by a sensor in the substance being heated, that this substance has reached its maximum design temperature. For FR.C species systems, in which liquid refrigerant will not enter the absorber Inlet from the refrigerant circulation pump by gravitation al flow while the pump is not operating, the system is self-protecting in the event of a power supply failure and also for most equipment fail ures,which disable the condensate pump. In the case where the substance being heated is a fluid -if a failure occurs, which stops the circulation of this fluid through the condenser, overpressure protection is achieved by a mechanical interlock, or an electrical interlock controll ed by the CCU, which de-activates the system, in the event of such fail ure until manually reset. Alternatively, the system can be protected against failure of the circulation of the fluid to be heated by the tech niques described under the heading "Over-pressure and Over-Temperature Protection" starting on page 41.

With insolation present, but condensate pump 14 not operating, all the refrigerant in the absorber ^' will collect in the liquid piping, re ceiver 11, condenser 8, and the vapor piping by the process of gravity and vapor migration. While the refrigerant is evaporating from the absorber and condensing in other parts of the system, the temperature of the collector is held close to the condensation temperature until starvation commences and the exiting vapor becomes superheated. The temperature of the more elevated parts of the absorber will rise, but the vapor pressure remains close to that of the condensation temperature at all times, and will be below maximum operating pressure when there is no liquid refrigerant at all left in the absorber. Thus, the system and the individual components, should be sized so that the entire refrigerant charge can drain into the-Other parts of the fluid circuit below the bottom of the absorber, including the condenser, receiver, and/or piping so that none remains in the absorber in the event of malfunction.

This self-protection design prevents destructive pressure build-up in the absorber, even when the absorber temperature approaches stagnation temperature, and also limits temperatures in other parts of the system because there is negligible heat transfer from the superheated vapor in the absorber. The refrigerant is also protected from excessively high temperatures in the stagnation condition, except for the small amount present as superheated vapor in the absorber.

For NRC systems, and for FRC systems in which liquid refrigerant can drain by gravity into the absorber inlet when the refrigerant circulation pump ;is not operating, it is necessary to protect the absorber and other system components when the refrigerant exceeds the maximum design pressure under stagnation conditions.

A protection technique Is shown for an NRC species system in Figure 7 and is also obviously applicable to an FRC species system where liquid refrigerant can drain by gravity into the absorber inlet. This method is to position in the merged return line 2-39 an automatically controll ed valve 67 just ahead of the absorber inlet 2. This valve 67 could be but is not limited to, an automatic two-position valve that can be operated in a so-called "fail-safe" manner. Namely, the valve would be normally kept closed in the absence of a control "signal" by, for example, a spring or weight, and would be opened by the control signal. The means employed to determine whether the refrigerant has exceeded its maximum design operating pressure may either be a pressure sensor or a temperature sensor. In the latter case, the operating pressure would-be assumed to have been exceeded if the measured temperature exceeded the refrigerant saturated vapor temperature corresponding to the maximum operating pressure. Examples of pressure sensors are (a) pressure-actuated switches, in which case the valve would be actuated by an electrical signal and would probably be a solenoid valve (of the "normally-closed" kind); (b) bellows filled with a gas , in which case the valve would be actuated (closed) when the pressure exceeds the maximum operating pressure; (c) bourdon-type pressure gauge mechanisms, in which case the valve would be mechanically actuated by the movement of the bourdon tube as it straightens out, as the pressure inside it increases, and it comes into contact with the stem of the valve; and (d) crankcase-type pressure regulating valves that close on rise of pressure Examples of temperature sensors are (a) platinum resistance transducers, in which case the valve would be electrically actuated and usually be a solenoid valve; (b) bulbs containing a mixture of liquid and vapor of, for example, the same type of fluid as the refrigerant, in which case the valve would be actuated by the pressure of the fluid's vapor as, for example, in the case of a theπnostatic expansion valve; and (c) bimetallic strips, In which case the valve would usually be mechanically actuated. The sensing element of the foregoing sensors (not shown in Figure 6) would preferably be located at some point in the absorber 1 near the vicinity of the outlet and, in the case of temperature sensors would either be immersed in the refrigerant, or would be in thermal contact with the tubing containing the refrigerant or with the absorb ing surface in the vicinity of this tubing.

The presence of the automatically controlled valve 67 in this Rankine-type system of my invention prevents the refrigerant from exceeding its maximum design operating pressure in the event liquid refrigerant is temporarily trapped in the absorber passageways, because of equipment malfunction, or because the medium being heated reaches its maximum design temperature or pressure, and the system reduces the rate at which heat is transferred from the refrigerant to the medium being heated below the rate at which radiant heat is absorbed by the refrigerant. The presence of valve 67- prevents actuation of a pres sure relief valve, or equivalent device, that would necessarily be incorporated in any practical system. Thus, the valve 67 would automatically prevent the liquid refrigerant from entering the absorber passage ways through inlet 2 whenever the refrigerant saturated vapor pressure exceeds a pre-selected value below, preferably by about 15 to 25%, the setting of the pressure relief valve, or below the setting of other safety means to dump the refrigerant if its pressure exceeds the safety device setting.

In NRC species systems, the liquid refrigerant must also be prevented from entering the absorber through its outlet 3. To this end, sufficient volume must be provided by the refrigerant circuit outside the absorber passageways to ensure that the entire refrigerant charge can be held by the circuit and that the level of the liquid refrigerant in the circuit does not rise above the highest level of the tube from the absorber outlet to the separator outlet. This can usually be accompushed without providing a liquid refrigerant reservoir — especially in cases where the levels of the condenser and separator, liquid inlets are above the level of the absorber. In FRC species systems, the entire refrigerant charge can be stored in the condenser, the liquid receiver and associated tubing, by proper sizing in design of. the system, so that no refrigerant can remain in the absorber under stagnation conditions and the system pressure is the vapor pressure of the stored liquid.

In FRC species systems, recovery from a stagnation condition may in some installations require that the CCU controls the system turn-on sequence to prevent a pressure surge due to the heat stored in the absorber structure. The sequence is first to turn-on the pump which circulates the medium to be heated (assuming it is a fluid) through the condenser then later, after sufficient time for the condensing temperature to stabilize, to slowly increase the circulating pump speed from zero to its normal controlled speed by a control signal from the CCU.

In NRC species systems, recovery from stagnation should first turn on the medium circulation pump, then after a short delay activate valve

67.

Second, the separator inlet and vapor outlet must be high enough for the net static head in the auxiliary refrigerant circuit to be zero or positive at the maximum evaporation rate and the minimum evaporation operating temperature,

2. C o m b i n e d O v e r - P r e s s u r e a n d O v e r - T e m p e r a t u r e P r o t e c t i o n T e c h n i q u e s The foregoing over-pressure protection technique prevents the pressure of the refrigerant and the temperature of the substance to be heated from exceeding their design values. It does not, however, limit the temperature of the absorber or of the refrigerant in the absorber. The combined over-pressure and over-temperature technique described next provides absorber and refrigerant over-temperature protection in addition to the refrigerant over-pressure and the substance over-temperature protection discussed earlier. The combined protection technique^uses a sensor to indicate that the maximum design operating pressure or saturated vapor temperature, or dry-bulb temperature, has been exceeded by sensing absorber or refrigerant temperature, refrigerant saturated vapor temperature, or refrigerant pressure. This sensor activates a heat rejection circuit to reject unwanted heat to an external heat sink, which will usually be water or the outdoor ambient air. In the case where the heat sink is air, the means used to transfer heat from the refrigerant to the air Is any type of air cooled condenser; and, in the case where the heat sink is water, the means is any type of water cooled condenser. The water for the water sink can be, or come from, any natural or man-made supply, such as the sea, a lake, an artesian well, or water from the cold-water supply of a man-made water distri bution system. The air or water condenser hereinafter referred to as "heat-rejection condenser", may use either natural air and natural water circulation, respectively; or forced air and forced water circulation, respectively. The heat- rejection rate, of this rejection circuit, is designed so that the temperature and pressure in the two- phase heat-transfer design system remains within design limits under the highest stagnation temperature and the highest ambient temperature. The combined over-pressure and over-temperature technique uses an intermediate condenser which transfers heat from the main Rankine heat-transfer system to the heat-rejection circuit, which is a small ancillary Class A_NN or Class B_N NRC species system.

The main Rankine heat-transfer system may be a Class A_NN or a Class B_N system. A Class A_NN system is illustrated in Figure 11 without loss of generality, illustrates the case where both the main and ancillary Rankine heat-transfer systems are Class A_NN systems. The main inlet 69 of the intermediate condenser 68 is connected to the separator outlet 6 of the main system, and the main outlet 30 of this Intermediate condenser is connected to the refrigerant inlet 9 of the main condenser 8. The refrigerant of the ancillary system exits intermediate condenser 68 at 71, enters ancillary separator 72 at 73. The evaporated dry refrigerant exits 72 at 74 and enters ancillary condenser 75 at 76 from which it exits, mostly in condensed form, at 77 and is returned by gravity to ancillary inlet 78 of intermediate condenser 68. The non-evaporated portion of the refrigerant exits ancillary separator outlet at 79 and is also returned by gravity to 78 after flowing through mergence point 79.

Valve 80, which activates the ancillary heat-rejection circuit when it is open may be actuated by a pressure or a saturated temperature or a dry-bulb temperature rise in the principal refrigerant circuit, as described earlier under the heading "Over-Pressure Protection Technique", either through a bellows and mechanical actuator or by an electrical signal from the CCU, or directly by an over-pressure or over-temperature swith. If an electrically operated (solenoid) valve is used, it would be normally open, when it Is not energised, to provide fail-safe protection. 3. S ub - a t m o s p h e r i c P r o t e c t i o n

FRC species heat-transfer systems may be designed for use in environments where the winter ambient temperatures fall below the boiling point of the refrigerant at atmospheric pressure. Because the liquid refrigerant is free to drain out of the absorber inlet, in general, a self-sustaining Rankine cycle will transfer heat, by liquid evaporation and condensation, from the lower parts of the system, where the liquid collects by gravity, to the absorber as it cools. Sustained low ambient temperatures may, therefore, result In the absorber internal temperature falling below the atmospheric-pressure boiling point of the refrigerant, so that the vapor pressure in the system becomes subatmospheric. With certain types of pumps, rotating seals may admit air to the system at subatmospheric pressure. There are three methods of preventing subatmospheric pressure at the location of the pump, under minimum design temperature. The first method consists of making heat available from an auxiliary heat source, such as a tank of the substance being heated, so that the internal temperature of the absorber does not fall below the atmospheric boiling point. This method is useful when the absorber heat-loss coefficient is small and the equivalent outdoor temperature is not far below the refrigerant boiling temperature at atmospheric pressure (say no more than 10°C) . In this case, auxiliary heat may be supplied by thermosyphon flow, of the substance being heated, through the condenser which contains liquid refrigerant under quiescent conditions. The second method is to configure the system so that the condensate pump (and other components requiring subatmospheric protection) are placed at a low enough level below the absorber for the vapor pressure in the system, plus the- pressure due to the quiescent static head of liquid refrigerant (which will be at minimum temperature also), to exceed the atmospheric pressure at the mechanical seal of the condensate pump. This method could also protect a rotating seal in the overfeed pump of a Class A_FF system if the pump is placed at a low enough level below the absorber. The advantage of the second method over the first is that it does not rely on a continuing supply of heat to protect the system.

The third method is shown in Figure 12, for the case of Class A_FN systems, and it may abviously be applied in the same way to the other three classes of FRC species. A liquid control valve 64 is located in the liquid line 16-17 between condensate pump outlet 16 and absorber inlet 2. (This valve would be located in liquid line 16-24 in a Class A_FF system, and in liquid line 16-2 in a Class B_F of a Class C_NN. system.) This control valve is operated by, for example, a temperature transducer 65 in thermal contact with absorber 1. The valve operates to shut off the liquid flow when the temperature sensed by 65 is below a preset value, at which the refrigerant vapor pressure is safely above atmospheric pressure. While valve 64 is closed, liquid refrigerant from vapor condensing in the absorber (and its auxiliary refrigerant circuit in the case of Class A_FN and A_FF systems) will be prevented from returning to the lower parts of the system. Therefore, liquid refrigerant will eventually fill the absorber (and the auxiliary refrigerant circuit, if applicable) , and the liquid-vapor interface near the absorber can be positioned in a well-insulated section of the line 6-9 (or line 3-9 or 3-5 for Class B or Class C systems, respecti vely) from which the heat-loss rate is much smaller than that from the absorber. The temperature of the absorber (and its auxiliary refrigerant circuit, if applicable) may then fall further, so that it is filled with sub-cooled liquid, without appreciably affecting the temp erature and pressure at all refrigerant liquid-vapor interfaces in the system, which may be maintained above atmospheric pressure by a small heat input (such as leakage through Insulation) to the liquid remaining in the lower part of the system. 4. F r e e z e P r o t e c t i o n Some FRC species systems may utilize a refrigerant such as water, which will freeze in the absorber if exposed when deactivated, to the lowest ambient temperatures which the system may encounter.

There are two basic methods of freeze protection. The first method is the same as the first method described for sub-atmospheric pressure protection in the previous section. This method relies on an auxiliary source of heat, such as a hot-water reservoir, to supply the heat required, to maintain the absorber (and the absorber auxiliary refrigerant circuit, if applicable), above freezing temperatures by latent heat transfer caused by vapor evaporation from the lower parts of the system and vapor condensation in the absorber (and the absorber auxiliar circuits, if applicable). The basic system requirement, to achieve this operation is that liquid refrigerant fills at least part of the condenser and that the auxiliary heat be supplied to the condenser to evaporate refrigerant liquid by (a) a thermo-syphon action or (b) thermal conduction.

The heat lost by the auxiliary heat source, while preventing absorber circuit freezing, is replaced automatically during normal heat transfer system operation. The hazard of the method, is that the auxiliary heat source will become so depleted, by extended severe environments with no solar radiation available, that the heat transfer rate wil be insufficient to maintain the absorber circuit above the freezing point, when the vapor will condense directly to the solid phase and eventually block the natural circulation. After this happens, melting and freezing may generate sufficient pressure to rupture the system. Therefore, the first method would be of use where the heat loss rates from the absorber and its auxiliary exposed circuits are small enough for an auxiliary heat source to be economically practical for protecting the system fora.period In" excess of the worst (historical) environment conditions, with adequate-asafety factor.

The second method consists of protecting the absorber (and its auxiliary circuit, if any) by two valves which isolate the exposed portions of the system (above line L-L in Figure 13) from the lower, environmentally protected portions, thus preventing vapor condensation and the accumulation of liquid which may freeze in the exposed portions of the system while it is de-activated. Figure 13 shows this method applied, for example, to a Class A_FN system, but the method can obviously be applied to any other FRC species system containing a liquid receiver 11 or, alternatively, in the case of a Class C_FN system a separator 4, in which liquid refrigerant is stored.

The absorber 1 and the absorber auxiliary refrigerant circuit 2-3-5-7-17-2 is configured so that, when the system is de-activated, the liquid refrigerant in the absorber auxiliary refrigerant circuit will drain down through the liquid line 2-17-16 and back through pump 14 into liquid receiver 11, which is able to hold all the refrigerant charge. If the pump 14 is of a type which will not permit reverse flow through It when de-activated, then pump bypass circuit 96-95-97 is opened to allow the liquid from the absorber circuit to drain down from 17 to 11, past the pump. Normally open valve 95 is opened after system de-activation and closed during system activation, by control 98 from the CCU. Valve 81 In the liquid line between mergence point 17r.and 96 is located at a level which is at, or above, the level of one-way valve 94 (between condenser outlet 10 and receiver inlet 12) , and is activated by a temperature sensor 82, in thermal contact with absorber 1. Valve 81 closes when the temperature of sensor 82 falls to a preset value Tv , which is a selected amount above the temperature at which refrigerant liquid will freeze in equilibrium with refrigerant vapor (le. the triple point of the refrigerant) . One-way valve 94 closes whenever the pressu in condenser 8 is equal to or less than the pressure in receiver 11. After, the liquid refrigerant has drained out of the absorber and its auxiliary circuit, and the absorber has cooled sufficiently for valve 81 to close, the vapor pressure in the absorber circuit 17-3-10 will fall below the pressure in 11, because of the continued heat loss from the.,abs'orber.Therefore, valve 94 will close and the absorber circuit wil become isolated from refrigerant vapor and liquid in circuit 16-13-12, which is protected from freezing by a benign environmentr or additional heat sources using conventional techniques. In applications where the liquid refrigerant stored in the receiver is at a temperature well above the turn-off temperature for valve 81, which is usually true with small subcooling in condenser 8, a supplemental condenser 83, with inlet 84- connected to 12 and outlet 85 connected to liquid line 13-15 at 86, is used to reject heat from, the liquid refrigerant in 11 to the environment. This additional heat rejection is designed to cool receiver 11 faster than the absorber cools until valve 82 closes. The heat from 83 is .rejected by a two-phase heat-transfer circuit 87-88-92-93-89-87, using a non-freezing refrigerant, in which 83 acts as the evaporator and 91 as the heat rejection condenser. The heat rejection circuit is activated under control of the CCU, as indicated by symbol 90, which opens valve 89 on system deactivation, and closes 89 again when the signal from the temperature of sensor 99 indicates that the temperature of the refrigerant in the receiver has fallen below "Tv . The signal from 99 is sup- plied to the CCU as indicated by symbol 100. (In a class C_FN system, the one-way valve 94 would be located in the refrigerant liquid and vapor line 3-5 so that the condenser 8 and the evaporator 4 would be isolated from the absorber when valve 94 is closed.)

Claims

I CLAIM:

1. An improved heat-transfer system for absorbing heat from solar radiant energy, and for transferring the absorbed heat to a substance to be heated, of the type having a refrigerant and a principal refrigerant circuit within which the refrigerant is circulated, which includes

(a) means for absorbing heat from the radiant energy, and for evaporating at least a portion of the liquid phase of the refrigerant entering said absorbing means, said absorbing means having an inlet and an outlet, and (b) means for condensing essentially all the evaporated portion of the refrigerant exiting said absorbing means, said condensing means having a refrigerant inlet and a refrigerant outlet; wherein the outlet of said absorbing means is connected to the refrigerant inlet of said condensing means, and the refrigerant outlet of said condensing means is connected to the inlet of said absorbing means; and wherein the improvement comprises collective means for ensuring that at least four internal operating conditions are satisfied for all external conditions under which the heat-transfer system is designed to operate; said internal operating conditions being: first, the refrigerant entering said absorbing means exists exclusively in its liquid phase and has a mass flow rate large enough to maintain the amount of superheat of the evaporated portion of the:refrigerant exiting said absorbing means below a preselected, positive, upper limit; second, the refrigerant vapor entering said condensing means is maintained in a dry state; third, the amount of liquid refrigerant backing-up in said, condensing means is small enough to maintain the effective condensing surface area of said condensing means above a preselected, positive, lower limit; and fourth, the absolute value of the difference between the saturated vapor temperature of the refrigerant exiting said absorbing means and the saturated vapor temperature of the refrigerant entering said condensing means is maintained below a preselected upper limit.

2. A heat-transfer system in accordance with claim 1, wherein the refrigerant fluid passages of the segment of the principal refrigerant circuit, between the outlet of said absorbing means and the refrigerant inlet of said condensing means, are large enough for the friction induced pressure drop in said segment — at the highest evaporation rate and the lowest evaporation temperature, at which the system is designed to operate — to be small enough for the fourth internal operating condition to be satisfied.

3. A heat-transfer system in accordance with claim 1, wherein the first, second, and third, internal operating conditions are achieved without employing a valve to throttle the evaporated refrigerant transferred from the exit of said absorbing means to the refrigerant inlet of said condensing means.

4. A heat-transfer system in accordance with claim 2, wherein the rate at which refrigerant is circulated in the principal refrigerant circuit is controlled primarily by a condensate pump, said condensate pump having an inlet and an outlet.

5. A heat-transfer system in accordance with claim 2, wherein the rate at which refrigerant is circulated in the principal refrigerant circuit is controlled solely by the net static head, in the principal refrigerant circuit, resulting from the combined action of the heat absorbed from solar radiant energy and the local gravitational field.

6. A heat-transfer system in accordance with claim 4, wherein said collective means includes means for separating the evaporated portion and the non-evaporated portion of the refrigerant exiting said absorbing means; said means having an inlet connected to the outlet of said absorbing means, a vapor outlet connected to the refrigerant inlet of said condensing means, and a liquid outlet connected to the principal refrigerant circuit at a mergence point located between the outlet of the condensate pump and the inlet of said absorbing means; and wherein the rate at which liquid refrigerant is circulated in a natural-circulation, absorber, auxiliary, refrigerant-circuit is controlled solely by the net static head resulting from the combined action of the heat absorbed from solar radiant energy and the local gravitational field.

7. A heat-transfer system in accordance with claim 4, wherein said collective means insludes

(a) means for separating the evaporated portion and the non-evaporated portion of the refrigerant exiting said absorbing means; said means having an inlet connected to the exit of said absorbing means, a vapor outlet connected to the refrigerant inlet of said condensing means, and a liquid outlet; and

(b) an overfeed pump, having an inlet connected to the liquid outlet of said separating means and an outlet connected to the principal refrigerant circuit at a mergence point located between the outlet of said condensate pump and the inlet of said absorbing means; and wherein the rate at which liquid refrigerant is circulated in a forced-circulation, absorber, auxiliary, refrigerant-circuit is controlled primarily by varying the effective capacity of the overfeed pump.

8. A heat-transfer system in accordance with claim 4, wherein said collective means Includes

(a) means for varying the effective capacity of the condensate pump to maintain the amount of superheat of the evaporated refrigerant exiting said absorbing means within preselected lower and upper limits, thereby satisfying the first and second internal operating conditions; and

(b) a receiver, having an inlet connected to the refrigerant outlet of said condensing means and an outlet connected to the inlet of the condensate pump, which has a large enough internal volume to store the total net change in the volume of liquid refrigerant present in said absorbing means and in said condensing means, over the entire range of values of the evaporation rate and the temperature of evaporation for which the heat- transfer system is designed to operate.

9. A heat-transfer system in accordance with claim 4, wherein said collective means includes means for separating the evaporated portion and the non-evaporated portion of the refrigerant exiting said absorbing means; said means having an inlet connected to the exit of said absorbing means, a vapor outlet connected to the refrigerant inlet of said condensing means, and a liquid outlef connected to the principal refrigerant circuit at a mergence point located between the refrigerant outlet of said condensing means and the Inlet of the condensate pump.

10. A heat transfer system in accordance witt claim 6 wherein said collective means includes

(a) means for varying the effective capacity of the condensate pump to maintain — for all external conditions under which the heat transfer system is designed to operate — the volume of liquid refrigerant in the natural-circulation absorber auxiliary ref_ϊ. rigerant in the natural-circulation absorber auxiliary refrigerant circuit (1) above a lower limit high enough for the evaporated refrig erant exiting said absorbing means not to be superheated, thereby satisfying the first internal operating condition, and (2) below an upper limit low enough for the refrigerant vapor supplied to said condensing means to be dry, thereby satisfying the second internal operating condition; and

(b) a refrigerant charge with a mass which is

(1) large enough for the condensate pump to maintain — for all external operating conditions under which the heat-transfer system is designed to operate — the volume of liquid refriger ant in acήatural-circulation absorber auxiliary refrigerant circuit above said lower limit, and the available net positive suction head to the condensate pump above the net positive suction head required by the condensate pump, thereby satisfying a fifth internal operating condition for the condensate pump ; and

(2) small enough for the condensate pump to maintain — for all external operating conditions under which the heat-transfer, system is designed to operate — the volume of liquid refrigerant in the natural-circulation absorber auxiliary refrigerant circuit below said upper limit and the amount of liquid refrigerant backed-up in said condensing means small enough to maintain the effective condensing-surface area of said condensing means above a preselected, positive, lower limit, thereby satϊr isfying the third internal operating condition.

11. A heat-transfer system in accordance with claim 7, wherein said collective means includes

(a) means for varying the effective capacity of the condensate pump to maintain — for all external conditions under which the heat-transfer system is designed to operate — the volume of liquid refrig erant in the forced-circulation absorber auxiliary refrigerant circuit

(1) above a lower limit high enough for the evaporated refrigerant exiting the absorbing means not to be superheated, thereby satisfying the first internal operating condition, and for the available net positive suction head of the overfeed pump to be above the required net positive suction head, thereby satisfying the fifth internal operating condition for the overfeed pump, and (2) below an upper limit low enough for the refrigerant vapor supplied to said condensing means to be dry, thereby satisfying the second internal operating condition; and

(b) a refrigerant charge with a mass which is (1) large enough for the condensate pump to maintain — for all external operating conditions under which the heat-transfer system is designed to operate — the volume of liquid refrigerant in the forced-circulation auxiliary refrigerant circuit above said lower limit, and the available net positive suction head to the condensate pump to be above the net positive suction head required by the condensate pump thereby satisfying the fifth internal operating condition for the condensate pump; and (2) small enough for the condensate pump to maintain — for all external conditions under which the heat-transfer system is designed to operate — the volume of liquid refrigerant in the forced-circulation absorber auxiliary refrigerant circuit below said upper limit, and the amount of liquid refrigerant backed-up in said condensing means small enough to maintain the effective condensing-surface area of said condensing means above a selected, positive, lower limit, thereby satisfying the third internal operating condition and

(c) means for varying the effective capacity of the overfeed punp to maintain — for all external conditions under which the heat-tran sfer system is designed to operate — the quality of the refriger ant vapor at the outlet of said absorbing means within a preselect ted lower limit greater than zero and a preselected upper limit up to and including one.

12. A heat-transfer system in accordance with claim 8, wherein said collecting means include

(a) means for measuring the amount of superheat of the evaporated refr rigerant exiting said absorbing means;

(b) means for controlling the effective capacity of the condensate pump to maintain the volume of liquid refrigerant in said absorbing means (1) above a lower limit to prevent the refrigerant vapor exiting said absorber from exceeding said upper limit of superheat, and (2) below an upper limit to prevent the refrigerant vapor exiting said absorber from exceeding said lower limit of superheat; and (c) a refrigerant charge with a mass which is (1) large enough for the condensate pump to maintain — for all external operating conditions under which the heat-transfer system is designed tύ operate — the volume of liquid refrig erant in said absorbing means above said lower limit and the available net positive suction head at the pump inlet to be above that required by the condensate pump, thereby satisfying a fifth internal operating condition for the condensate pump; and (2) small enough for the condensate pump to maintain — for all external conditions under which the heat-transfer system is designed topoperate — the volume of liquid refrigerant in said absorbing means below said upper limit and the amount of liquid refrigerant backed-up in said condensing means small enough to maintain the effective condensing-surface area of said conden sing means above a preselected, positive, lower limit, there- by satisfying the third internal operating condition.

13. A heat-transfer system in accordance with claim 9, wherein said collective means includes (a) means for varying the effective capacity of the condensate pumpto to maintain — for all external conditions under which the heat- transfer system is designed to operate — the quality of the refrigerant vapor exiting said absorbing means within a preselected lower and upper limit , greater than zero, and a preselected upper limit up to and including one, therebμ satisfying, the first intera nal operating condition, and (b) a refrigerant charge with a mass flow which is

(12 large enough for the condensate pump to maintain — for all external operating conditions under which the heat-transfer system is designed to operate — the greatest volume of liquid refrigerant required in the absorbing means for the vapor quality to be at the said lower limit and for the available net positive suction head at the pump inlet to be above that required by the pump, thereby satisfying the fifth inter nal operating condition for the condensate pump, and

(2) small enough for the condensates-pump to maintain — for all external operating conditions under which the heat-transfer system is designed to operate — the least volume of liquid refrigerant required in the absorbing means for the vapor quality to be at the said upper limit and the amount of liquid refrigerant in said separating means and backed up in said condensing means to be small enough that the second and third internal operating conditions are satisfied.

14. A heat-transfer system in accordance with claim 9, wherein said collective means includes

(a) a condensate pump with constant effective capacity

(1) high enough to maintaintthe quality of the vapor exiting said absorbing means at or below one — for all externalloperating conditions under which the heat- transfer system is designed to operate — thereby satisfying the first operating condition, and

(2) low enough that the vapor exiting said separating means is in a dry state, satisfying the second operating condition, and

(b) a refrigerant charge with a mass which is

(1) large enough to maintain — for all external operating conditions under which the heat-transfer system is designed to operate — the greatest amount of liquid refrigerant in said absorbing means for the vapor quality to be one or less and for the available net positive suction head at the pump inlet to be above that required by the pump, thereby satisfying the fifth internal operating condition for the condensate pump, and (2) small enough to maintain — for all external operating conditions under which the heat-transfer system is designed to operate — the least amount of liquid refrigerant in said absorbing means for the vapor quality to be one or less and the amount of liquid refrigeranttin said separating means and backed up in said condensing means to be small enough that the second and third internal operating conditions are satisfied.

15. A heat-transfer system in accordance with claim 9, which includes means for subcooling liquid refrigerant, having a refrigerant inlet connected to the outlet of the condensate pump and a refrigerant outlet; connected to the inlet of said absorbing means; wherein said collective means also includes

(a) means for varying the effective capacity of the condensate pump to maintain — for all external conditions under which the heat-trans fer system is designed to operate — the amount of sensible heat released by the refrigerant in the subcooler at a selected value, and (b) a refrigerant charge with a mass which is (1) large enough to maintain — for all external operating conditions under which the heat-transfer system is designed to operate — the greatest amount of liquid refrigerant in said absorbing means for the vapor quality to be one or less and for the available net positive suction head at the pump inlet to be above that required by the condensate pump, thereby sat isfying a fifth internal operating condition for the condensat pump, and (2) small enough to maintain ferfor all external operating conditions, under which the system is designed to operate — the least amount of liquid refrigerant in said absorbing means for the vapor quality to be one or less and the amount of liquid refrigerant in said separating means and backed up in said condensing means to be small enough that the second and third internal operating conditions are satisfied.

16, A heat-transfer system in accordance with claim 10 or 11 or 12, wherein said collective means also includes

(a) a receiver, having an inlet connected to the refrigerant outlet of said-Icondensing means and an outlet connected to the inlet of the condensate pump, which has a large enough internal volume to store the total net change in the volume of liquid refrigerant present in said absorbing means and said condensing means, over the entire range of values of the evaporation rate and the temperature of evaporation for which the heat-transfer system is designed to operate, and (b) means for measuring the level of the surface of liquid refrigerant contained in said receiver, and (c) means for controlling the capacity of said condensate pump in response to said surface level over the range of values of evaporation rate and temperature of evaporation for which the system is designed to operate.

17. A heat-transfer system in accordance with claim 16, wherein said collective means also includes means for controlling said pumping means as a function of the said liquid level in such a manner that the flow rate of the condensed working fluid is zero when said liquid level falls below a minimum fixed preselected level and is greatest when said liquid level rises above a maximum fixed preselected level.

18. A heat-transfer system in accordance with claim 16, wherein said collective means also includes (a) means for controlling said pumping means as a function of the said liquid level in such a manner that the flow rate of the condensed working fluid is zero when said liquid level falls below a minimum variable preselected level and Is greatest when said liquid level rises above a maximum variable preselected level, and (b) means for controlling said variable preselected level to accomodate the variation in volume of the liquid refrigerant within said absorbing means and said condensing means, over the entire range of evaporation rates and evaporation temperatures for which the system is designed to operate.

19. A heat-transfer system in accordance with claim 12, wherein said collective means also includes

(a) means for stopping the condensate pump when the surface level of the liquid refrigerant in said receiver is below a preselected lower limit, and

(b) means for measuring the superheat of the refrigerant vapor exiting said absorbing means, and

(c) means for increasing and reducing the effective capacity of said condensate pump as a function of the difference of said superheat measurement above said lower limit of superheat and below said upper limit of superheat.

20. A heat-transfer system in accordance with claim 19, wherein said collective means also includes

(a) means for measuring the temperature of the refrigerant vapor exit ing said absorbing means, and

(b) means for measuring the pressure of said refrigerant vapor, and

(c) means for computing the superheat of said vapor deriving the temperature of saturated vapor at said measured pressure and subtracting from the measured temperature of said vapor.

21. A heat-transfer system in.--accordance with claim 19, wherein said collective means alsolincludes

(a) means for measuring the pressure of saturated refrigerant vapor at

(b) means for measuring the pressure of said vapor, and (c) means for deriving from the two said pressure measurements a superheat measurement of sufficient accuracy.

22. A heat-transfer system in accordance with claim 13, wherein said collective means includes (a) means for stopping said condensate pump if the level of the liquid refrigerant in said separating means falls below a preselected limit, and

(b) means for measuring the liquid refrigerant volumetric flow rates in

(1) the liquid line from the outlet of said condensing means, and

(2) the liquid line from the liquid outlet of said separating means, to the mergence point in the liquid line leading to the inlet of the pump, and

(c) means from deriving a volumetric flow rate control for the con densate pump from the measures of said flow rate measurement means so that the vapor quality at the outlet of said absorbing means is a preselected function of the evaporation rate.

23. A heat transfer system in accordance with claim 1 , wherein said collective means includes means for measuring the evaporation rate and means for controlling the volumetric flow rate of said condensate pump to be a given function of the condensation rate.

24. A heat transfer system in accordance with claim 15, wherein said collective means includes

(a) means for stopping said condensate pump if the level of the liquid refrigerant in said separating means falls below a preselected lower limit, and

(b) means for measuring the temperature of the refrigerant at the inle of the subcooler and

(c) means for measuring the temperature of the substance being heated at the outlet port of the subcooling means for this substance, and

(d) means for deriving a volumetric flow rate control for the condensate pump as a function of the difference of the measured temperatures of the refrigerant and the liquid being heated at the ports of the cooling subcooling means relative to a preselected value.

25. A heat transfer system according to claim 17, wherein said collective means also includes means for measuring the level of the surface of the liquid refrigerant in said receiver which comprises a differential pressure transducer with inlet one connected to the liquid line between said condensing means and the inlet of said receiver, so configured that, during system operation liquid refrigerant will fill the line to Inlet one under the action of gravity, and with inlet two connected to the liquid line from the outlet of said receiver to the inlet of said condensate=pump, close to the outlet of said receiver, so that the liquid surface level is determined by the pressure difference of the two liquid columns above said inlets one and two.

26. A heat-transfer system in accordance with claim 18, wherein said collective means also includes (a) means for measuring the level of the surface of the liquid refrr rigerant in said receiver which comprises a diffential pressure transducer with inlet one connected to the liquid line between said condensing means and the inlet of said receiver, so configured that, during system operation liquid refrigerant will fill the line to inlet one under the action of gravity, and with inlet two connected to the liquid line from the outlet of said receiver so that the liquid surface level is determined by the pressure of the two liquid columns above said inlets one and two. (b) means for measuring the saturated vapor temperature at the outlet of said absorbing means and deriving, from the vapor temperature measurement, a measure of the density of the liquid refrigerant in said receiver to compensate said liquid level measurement for the temperature of evaporation.

27. A heat-transfer system in accordance with claim 18, wherein the means for capacity control of the overfeed pump, in accordance with claim 11, comprise means for measuring the speed of the condensate pump and the speed of the overfeed pump and means for deriving a control for the speed of the overfeed pump as a function of the speed of the condensate pump.

28. A heat transfer system in accordance with claim 18, wherein the means for capacity control of the overfeed pump, in accordance with claim 11, comprise means for measuring the volumetric flow rate of the condensate pump and the flow rate ofΛthe said overfeed pump and means for deriving a control for the speed of the overfeed pump - to make the volumetric flow rate of the overfeed pump a given function of the volumetric flow rate of the condensate pump.

29. A heat-transfer system in accordance with claim 16, or 18, or 23, wherein the means used to determine the evaporation rate include (a) means for measuring the stagnation temperature the absorber would have if heat were not being removed from it by the refrigerant, (b) means for measuring the evaporation temperature of the refrigeran and (c) means for computing the evaporatio rate of the refrigerant as a function of the difference between the said stagnation temperatur and the said evaporation temperature.

30. A heat-transfer system in accordance with claim 29 which also includes

(a) means for measuring the outdoor air temperature in the vicinity of the absorber, and

(b) means for computing the evaporation rate also as a function of said air temperature.

31. A heat-transfer system in accordance with claim 5, wherein said collective means includes means for separating the evaporated portion and the non-evaporated portion of the refrigerant exiting said absorbing means; said means having an inlet connected to the outlet of said absorbing means; a vapor outlet connected to the refrigerant inlet of said condensing means; and a liquid outlet connected to the principal refrigerant circuit at a mergence point located between the refrigerant outlet of said condensing means and the inlet of said absorbing means.

32. An improved method of performing the function of absorbing he from a solar radiant energy, and transferring and releasing it to a substance to be heated of the type wherein

(a) a refrigerant is circulated in a.:ρrincipal refrigerant circuit,

(b) heat is absorbed from said source, at least in part, by evaporating the liquid phase of the refrigerant in a means for absorbing radia and energy, and

(c) the absorbed heat is released to the sunstance, at least in part, by condensing essentially all the evaporated portion of the refrigerant in a means for condensing a vapor; wherein the improvement comprises satisfying at least four internal operating conditions for all external conditions under which the heat-transfer system is designed to operate ; said four conditions being

First, maintaining the refrigerant entering said absorbing means in its liquid phase and keeping the mass flow rate of the refrigerant entering said absorbing means large enough to maintain the amount of superheat of the evaporated refrigerant exiting said absorbing means below any preselected positive upper limit, Second, maintaining the evaporated refrigerant entering said condensing means in a dry state;

Third, limiting the liqud refrigerant backing up in said condensing area of said condensing means above a preselected, positive lower limit, and Fourth, maintaining the absolute value of the difference between the satu rated vapor temperature of the refrigerant exiting said absorbing means and the saturated refrigerant temperature entering said condensing means below, a positive, preselected upper limit.