CA1152442A - Method and apparatus for high volume distillation of liquids - Google Patents

Abstract

ABSTRACT OF THE DISCLOSURE

A method and apparatus for high volume distillation of impure liquid comprises evaporating the impure liquid in an evaporator to form a vapor at a temperature above the freezing point and below the critical point, preferably the boiling point, of said liquid at atmospheric pressure and at a pressure not exceeding a pressure corresponding to the evaporation temperature under saturated conditions;
compressing the vapor; passing at least a portion of the compressed vapor through an expansion engine to do work upon and motivate the engine and to produce shaft energy, whereby the vapor expands and cools; adding work to the expansion engine to make up the difference between the work done in compressing the vapor and the work done on the engine by the vapor expanding therethrough; compressing the expanded vapor to form a second vapor; cooling the second vapor in heat transfer relation with the impure liquid whereby the vapor at least partially condenses, transferring sufficient heat to the impure liquid for evaporating the liquid and to form the aforementioned vapor; and collecting the condensed vapor.

Description

~;z44Z ' ME'l'IIOD ~ND APPARI~TUS FOR HIGH VOLUME
DISTILLATION O~ LIQUIDS

The present invention relates to a method and apparatus for economically and efficiently purifying and recovering high ~uality water from waste water and, more part;cularly, t'o a method and apparatus which permits ev~poration'''and vapor compression treatment of large volumes' of impure water.
The need for very large volumes of high quality water arises in many contexts. Many industries require large quantities of good quality water as input or raw material in order to operate. For example, the paper or textile industries utilize tremendous volumes of such water for their dyeing and bleaching operations. Many more industries discharge large quantities of waste or contami-nated aqueous solutions to the environment. However, with the continuing decline in quality of the water in our lakes, rivers and streams and the continuing promulgation by federal, state and local governments of statutes and or dinances regulating the quality of water dumped into water-ways, there is an increasing need for economical methods by which industrial waste streams can be cleaned prior to discharge. Still another area which requires the treatment of large volumes of water in an efficient and economical fashion is the production of potable water from the oceans by desalination. A related a'rea for treatment of large volumes of water is the treatment of sea wate'r into which . . .
oil has been spilled to recover the oil and to d~salinate or purify the water. Thus, the problem of waste water treat-ment in high volumes includes the treatment of impure water ~k as well as sea or brackish ~ater. It also includes the treatment of water containing inorganic or organic impur-ities or materials where it is desired to separate and recover the water and/or to separate and recover the ma-terials. In a broader sense the problem is not limited to water or aqueous solutions but extends to non-aqueous solutions as we'll where the components can be substantially separated ~ ~he method of distillation. Therefore, all possible feed solutions for liquid separation of the solvent from other constituents of the solution, whether the solvent is a~eous or not, are encompassed within the term "impure liquid" as used herein.
There have been endless suggestions for treating industrial waste and sea water, incl~ding multistage dis-tillation plants, thermo-mechanical distillation systems, and the like. However, any system heretofore suggested which has been capable of treating the millions of gallons per day necessary to effectively deal with industrial waste or to produce meaningful quantities of potable water have been hopelessly impractical or uneconomical in terms of their capital equipment or energy requirements. A good illustration of this is the system disclosed in U.S. Patent No. 3,423,293 to Holden, which is a thermo-mechanical system for distilling impure waste at one atmosphere. The Holden system includes, seguentially, a boiler for evaporation of the water, a compressor, heat exchange means for adding heat to the compressed'vapar','a-t'urbine motor for driving the compressor and a condenser unit for extracting the heat .. :.. . . -of vaporization from the vapor and for transferring this extracted heat to the impure feed liquid at one atmosphere.
Although Holden makes a seemingly appealing case for the economics of his system, when practical thermodynamic con-siderations are imposed it bec~mes apparent that in order to treat large volumes of ~ater i~ the Holden system, e.g., 1,000,000 gal/day or 125,000 gal/hr, would require about 1,250,000 ft of condenser heat transfer area. Using com-mercially available condensers, this means that a typical 20 inch wide condens~r would have to be 18,266 feet long.
If the condenser size were increased to 5 feet wide, a condenser length of 2,03i running feet would be required.
The capital costs involved in building a support structure for such a condenser unit are too impractical to consider.
Other Uni ~ States patents which teach or disclo~ water distillation systems and which may be of some lnterest in connection with the present invention are the following: 1,230,417;
1,594,957; 2,280,093; 2,51S/013; 2,537,259; 2,589,406;

2,637,684; 3,412,558; 3,423,293; 3,425,914; 3,351,537;

3,440,147; 3,444,049; 3,476,654; 3,477,918; 3,505,171;
3,597,328; 3,477,918; 3,505,171; 3,597,328; 3,607,553;
3,649,469; 3,856,631; 3,879,266.
It i5 therefore an object of this invention to provide an economical yet practical system for high volume purification of impure liquid sources.
It is another object of this invention to pro-vide a thermo-~,echanical distillation system capable of purifying large volumes of impure liquids and convert-ing them to potable, or at least dumpable, liquid without imposing unreasonable equipment or energy requirements.
It is still another object of the fnvention to provide a heat and work input system wherein maximum heat and work input efficiencies are practiced.
It is yet another object of this invention ~ ;Z442 to provide a syst~m capable of purifying millions of gallons per day of waste water while at the same time providing a thermal energy reserve which can be used as such or converted to mechanical or electrical energy.

Other objects and advantages will become apparent from the follow'ing description and appended claims.
B'r'iefly stated, in accordance with the aforesaid objects' one broad aspect'of the present invention comprises a method, and a system for practiclng the method, for purifying large volumes of impure liquid by evaporating the liquid in a boiler, preferably under reduced pressure, substantially adiabatically compressing the resulting vapor to a pressure substantially in excess of the vaporiza-tion pressure, directing at least a portion of the compressed vapor through and substantially adiabatically expanding the vapor in an expansion engine, adding sufficient make-up work to the expansion engine such that the added work plu8 the work done by the vapor passing therethrough at least equals the work done by the compressor on the vapor, compressing the expanded vapor to form a second vapor and passing the resulting second vapor through a condenser, such as the condenser side of the boiler, wherein the second vapor will, upon condensing, give up at least enough thermal energy to vapori2e the feed liquid. The work added to the turbine can be added by directly mixing the compresse'd-vapor, under substantially isobaric conditions, with a volume of hot gas', e.g., .. .. ~ .
combustion gas, or by directly driving the turbinel e.g.

with an externally powered engine, by a combination of direct mixing and direct driving, or by other means well ~ 5Z442 known in the art.
In another broad aspect of the invention there is provided a method, and a system for practicing the method, for purifying large or small volumes of impure liquid by evaporating the liquid in a boiler under a pressure not exceeding the saturated liquid vapor p~essure, sub~tantially adiabatically compressing the resulting vapor to a pressure substantially in excess of the vaporiza-tion pressure in a compressor capable of prod~cing a variable compression ratio, and passing the resulting vapor through a condenser, such as the condenser side of the boiler, wherein the vapor will, upon condensing give up thermal energy to vaporize the feed liquid.
In an optional form of the invention, the compressed vapor is directed through and substantially adiabatically expands in a turbine before passin~ to the condenser.
The compressor is preferably driven by linking it to the shaft of an auxiliary turbine which may itself be driven by passing a volume of hot gas, e.g., combustion gas, steam, etc., therethrough. In one embodiment, the auxiliary turbine blading is annularly disposed with respect to the compressed vapor flow path and is driven by combustion gases produced in the annular space. Al-ternatively, the compressor may derive at least a portion Of its power from motor means shaft linked directly there-to.

.... .. . . .. . .
According to this method, maximum utilization -- - is made of available thermal energies with the result that more efficient and economical high volume purifica-tion can be accomplished than with any other method heretofore known. Moreover, the system of the present ~ 1524~2 invention, because its operation is independent of the method of evaporation, e.g., vacuum or flash distillation are both suitable, is extremely flexible in terms of its utility, and physical location. In the most common usage, where the impure liquid is impure water, the system is able to furnish large quantities of useful thermal energy, in the form of steam, in addition to large quantities of purified water.
The invention will be better understood from the following description considered together with the accompanying drawings, wherein like numerals designate like components, in which:
FIGURE 1 illustrates schematically a single stage embodiment of the purification system of the present invention showing an exemplary means and an alternative means (in phantom) for adding work to the turbine.
FIGURE lA illustrates schematically another single stage embodiment of the present invention including an independent compressor and exemplary and alternative (in phantom) means for operating the independent compressor.
FIGURE lB illustrates a variation of the FIGU~E
lA embodiment.
F~G'JRE 2 illustrates schematically the single stage embodiment of FIGURE 1, with the vapor treatment section deleted, including means for diverting a portion of the effluent vapor for direct mixing with the raw . ,........................... .. ... ~, .
feed liquid.
FIGURE 3 illustrates schematically another single stage vaporization embodiment of the present inven-tion.

ll~Z44Z
FlGURE 4 ill~strates schematically a multi-stage embodiment of the present in~ention, particularly suited for vacuum distillation-vapor compression treat-ment of waste water.
FIGURE S illustrates schematically a multi-stage embodiment of the present invention, particularly suited for flash distillation-vapor compression treatment of waste water.
FIGVRE 6 illustrates schematically another multi-sta~e flash distillation embodiment of the present invention.
FIGURE 7 illustrates schematically a clutched compressor unit which can be operated by a turbine motor as an optional turbine-compressor unit useful in the many embodiments of the present invention.
FIGURE 8 illustrates schematically two turbine motors operating a single turbine compressor as an op-tional turbine-compressor unit useful in the many embodi-ments of the present invention.
FIGURE 9 illustrates schematically a single turbine motor operating two turbine compressors as an optional turbine-compressor unit useful in the many embodi-ments of the present invention.
FIGURE 10 illustrates schematically two turbines, one of which can be powered by dirty, hot qases, operating a turbine compressor as an optional turbine-compressor .. . . . . . .
unit useful in the many embodiments of the present inven-.. tiQn.
FIGURE 11 illustrates schematically concentric compressor-turbine combinations, one of which combinations can be powered by dirty, hot gases, as an optional turbine-~ 5244Zcompressor unit useful in the many embodiments of thepresent invention.
FIGURE 12 illustrates schematically a centrifugal compressor operated by two turbine motors in tandem as an,optional turbine-compressor unit useful in the many embodiments of~the present invention.
FIGURE,13 ill~strates schematically a centrifugal compressor and a turbine compressor operated by a sin~le turbine motor as an optional turbine-compressor unit useful in the many embodiments of the present invention.
FIGURE 14 illustrates schematically an optional free wheeling compressor unit with two turbine driven compressors in tandem, which unit is useful as the turbine-compressor unit in the many embodiments of the present invention.
FIGURE 15 illustrate,s schematically a single stage embodiment of the present invention wherein the tu~rbine by-pass is eliminated'.
FIGURE 16 illustrates schematically a single stage embodiment of the purification system of the present invention 'in which the vapor treatment section of the system includes an in-line turbine as well as a variable ratio compressor and an auxiliary turbine configured to be operated by combustion gases produced by in situ combustion of fuel and air.
FIGURE 17 illustrates schematically an alterna-.. .. . . .
tive vapor treatment section comprising ,a compressor -,but no turbine, which section may be employed in conjunc-tion with or in place of the system of FIGURE 16.
FIGURE 18 illustrates schematically still another vapor treatment section useful in the embodiment of FIGURE

~SZ44Z
16, wherein the vapor treatment section includes a com-pressor, an optional in-line turbine, and an auxiliary turbine configured to be operated using available hot gases.
The invention will be better ~nderstood and appreciated fro~m a consideration of a preferred embodi-ment thereof which, for purposes of a descriptive clarity, includes only a single effect evaporative unit. It is of course appreciated, as is well known in the art, that multi-effect evaporative systems have many efficiencies which recommend them in practical usage. The present invention, as will be seen from the description of ad-ditional embodiments, contemplates the use of multi-as well as single-effect evaporative units. In addition, the invention contemplates both vacuum and flash evapora-tion as well as any other known evaporative techniques for producing high volumes of vapor at Pl, Tl, as will more clearly appear hereinafter. It is, however, preferrèd to use vacuum evaporation or vacuum distillation in most instances due to the greater flexibility it affords in terms of plant location.
Referring now to Figure 1, a vacuum distilla-tion-vapor compression system is shown generally at 900.
The system consists in its essential aspects of a boiler unit 904 including a condenser section 906 therein, a variable compression ratio turbine compressor 912 operated ..... . . ......................... .
through shaft 924 by turbine motor 916, turbine bypass arms 920, a mixing chamber 925 downstream of the turbine motor 916, and means for supplying additional or make-up work to turbine 916, i.e. work not done on the turbine by the vapors passing therethrough. The work supplying 3neans may be hot clean gas supplying means 934 ~or sup-plying hot gases, e.g. combustion gases, to mixinq chamber 914 for direct combinàtion with the compressed vapors from compressor 912 to motivate turbine 916. Alternatively, in lieu of hot clean gases, or in addition thereto, the turbine 916 can be directly driven thro~gh its shaft 924 ~y mo~or means 917, such as an electric or diesel powered motor, acting through shaft 913 and clutch and gear box 915 (shown in phantom). It will be appreciated, therefore, that the language "adding make-up work to the turbine" or similar expressions used herein are in-tended to contemplate any addition of work to the system, whether directly or indirectly to the turbine, where the effect of that work is to motivate the tùrbine.
To understand the operation of the system 900, the path of raw feed, e.g., impure water, therethrough can be charted. Initially, a starter motor, such as motor 917, is energized to rotate shafts 913 and 924 through clutch and gear box 915. Compressor 912 and turbine 916, which are linked to shaft 924, also rotate when the mo~or 917 is operated. During start-up, the compressor 912 is allowed to rotate for a time sufficient for a vacuum to be drawn on the evaporative side of boiler 904. The extent of the vacuum is predetermined, as will be seen hereinafter, based upon the desired operating parameters of the system and the temperature of the influent impure water and is controlled and monitored by variable pressure valve 911 in duct 910 joining the ~ ~ .. ... . .
boiler 904 and compressor 912. Optional means 934 for supplying hot gases to mixing chamber 914, if present, may be operated to motivate turbine 916 to keep it running ~ 15;~44Z
during start-up and to heat the tu~es 906 in the condenser section.
In this embodiment, a source 934 for clean hot gases is shown for supplying work to turbine 916 through duct 936 and may comprise a gas turbine sysltem, described in co~nection with Figùre lA, or any other known way of providing high temperature, high pressure gases, e.g., burning garbage at high temperature to pro-duce high temperature, high pressure steam, may be used.
Alternatively, the clean gas source 934 and mixing chamber 914 can be entirely dispensed with and a motor, such as motor 917, u~ed to provide the additional work to turbine 916 through shafts 913 and 924. If desired, both direct mixing with hot gases and direct mechanical drive can be used together, or any other suitable method employed for adding necessary work to the turbine.
Referring to Figure 1, which is described using direct gas mixing as the means for adding make-up work to turbine 916, it can be seen that the impure liquid feed enters system 900 through feed duct 902 and i5 rapidly heated to the boillng temperature, which depends on the vacuum level in the boiler 904, by heat transferred from the vapor condensing in hot condenser tubes 906. Un-vaporized concentrated feed liquid, contain~ng a large proportion of impurities therein, is removed from the boiler 904 through line 905. The vapor produced by .... ,. , . , ~ . ;
boiling at Pl, Tl is drawn through moisture separator ...... - --.908 and into duct 910 leading to turbine compressor 912.
The pressure Pl is maintained in boiler 904 at a level not exceeding a pressure corresponding to Tl under saturated conditions by pressure regulating valve 911 disposed ~ 1~;2442 in duct 910. The vapor is s~bstantially adia~atically compressed at a ratio of from 1.2:1 to 250:1, preferably 5:1 to 100:1 and more preferably 5:1 to 50:1, in compressor 912 to P2, T2 and, upon leaving compressor 912, can proceed either through mixing chamber 914 and turbine motor 916 or can be diverted by by-pass control valves 918 into by-pass arms 920. Although two by-pass arms 920 are shown for descriptive convenience, there may, in fact, be only one by-pass arm or there may be multiple by-pass arms. Moreover, the vapor which flows into the by-pass arms may be at the same or at a higher pressure than the vapor which proceeds through turbine motor 916.
Inasmuch as turbine compressors are freq~ently multi-stage units, and since the extent of compression depends on the number of stages through which the vapor passes, it is a simple matter to direct the flow into the by-pass arms 920 from a different compression stage than the flow which proceeds through turbine 916.
In accordance with this embodiment, it i8 contem-plated that as little as a fraction of 1% or as much as a raction less than 100% of the vapor flow exiting compressor 912, e.g., 0.01-99.9% by weight, preferably .15-95%, may be diverted into by-pass arms 920. Although it is unlikely that in practical operation the amol~nt of vapor by-passing turbine 916 will be at either extreme, as will appear more clearly from the description which , . . . .... . .
follows, the system 900 is operative at the extremes -- - as well as at any point therebetween. The selection of the amount of flow to be diverted depends upon the economics sought from the process, the volume flow rate reguired and whether reduced operating expenditures take ~ lSZ44Z
precedence over capital equipment expenditures, or vice-versa.
Assuming that direct mixing with hot gases is the method chosen to add work to the system upstream of or at turbine 916, the vapor which proceeds through compressor 912 is substantially isobarically admixed in mixing chamber, 914 with hot, clean gases s~pplied ~rom source 934 through duct 936 and emitted from in-jectors 922, The mixing chamber 914 may be a mixing injector, mixing aspirator, jet mixer or any other con-figuration k'nown to be suitable for mixing vapors having different pressures in such a manner that a partial vacuum is created upstream of the actual mixing point.
The partial vacuum is useful for drawing the non-injected vapor into the mixing chamber and thereby enhancing the mixing. The mixture of vapor and ~ases operate turbine motor 916 which is linked by shaft 924 to compressor 912. The temperature of the added gas is sufficiently greater than the temperature of the vapor to heat the vapor, at substantially constant pressure (i.e., P3=P2), by at least about 2K to T3 before the vapor does work W2 on turbine 916. Because of the direct shaft link between turbine 916 and compressor 912, the work W2 done on the turbine equals the work Wl done by the compressor on the vapor in substantially adiabatically compressing it. The vapor substantially adiabatically expands through , ., , ,,,, . ;
turbine 916 with a resultant pressure and temperature drop to P4, T4-The vapor which is diverted through by-pass arms 920 is at a temperature and pressure which equals T2, P2 in the case where all vapor is equally compressed 1~;24~Z
in compressor 912. The by-pass vapor is recombined with the vapor passing through the turbine in injector or mixing section 925 wherein the bypass vapor is injected through injectors 926 into the stream of vapor exhausting the turbine. Mixing section 925 can have any suitable configuration for efficient mixlng of vapors. The effect of this vapor mixing is to compress and heat the vapor exiting t~rbine 916 to ambient pressure, since the system downstream of turbine 916 is open to the ambient, and to T5, whereupon the mixed vapor proceeds through vapor return duct 928 to condenser tubes 906 in boiler 904.
The heat transfer temperature differential between the returning vapor at T5 and the feed water at Tl must be high enough that large volumes of feed water can be accomodated in this system within the practical limits imposed by reasonable condenser size. The vapor condenses in tubes 906 giving up its heat of vaporization to the feed li~uid entering the system through feed duct 902, Purified condensate may be removed from the system for general usage through line 930. Excess steam may be diverted through line 932 to keep the system in thermal balance; to heat the raw feed or to be injected into boiler 904, as will appear from a discussion of Figure 2, or for other purposes.
. It will appreciated that bypassing the turbine with at least a portion of the vapor together with 'the .... .. . . .. . .
mixing action created by injectors 922 upstream of the . tuFbine and injectors 926 downstream of the turbine have the net effect of creating a vacuum at the turbine inlet which materially eases the task of maintaining turbine rotation at a level sufficient that compressor 912 is ~15244Z
able to perform a quantity of work Wl in compressing the vapor. Nevertheless, a quantity of work W2=Wl must still be done on turbine 916 by the vapor passing there- _ through. Since the quantity of vapor passing through the turbine is decreased to the extent of the bypass, not as much vapor is available to run the turbine and the en~rgy content of the bypass vapor m~st be compensated for, as, for example,.by the addition of thermal energy via the gases, which may be combustion gases, injected into mixing chamber 914 through injectors 922. The hot gases as well as the additional thermal energy may be furnished in any form, as long as the gases are clean, from any available source. Suitable sources may include hot combustion gas sources, high temperature, high pressure steam sources, and the like. It will be appreciated, however, as previously indicated, that hot gas mixing to raise the thermal energy of the vapor and thereby permit the vapor to do the quantity of work W2 on the turbine is not the only means of adding make-up work.
Instead, the hot gas source 934, duct 936, injectors 922 and mixing chamber 914 can all be eliminated and the quantity of make-up work needed to reach W2 which is not supplied by the vapor can be furnished by directly driving the turbine through mechanical means, such as motor 917.
Where, however, hot gases are added to the .,. ., , ., ",.~ . .
vapor to raise its thermal energy, it is preferred that ` ~ direct mixing of gases occur in the space between the first compressor 912 and turbine 916. Alternative vapor heating configurations, such as by heat exchange through a conventional heat exchanger as taught in U.S. 3,423,293 3~15;Z442 - Holden, is wasteful of thermal energy due to transfer inefficiencies and the resulting need for higher tempera-ture heat transfer mediums, and is therefore uneconomical.
Improved vapor and combustion gas mixing and more uniform temperature distribution along mixing chamber 922 can be achieved by use of multiple nozzle injectors (not show~ in chamber 922.
The system illustrated in Figure l and the embodiments to be described hereinafter are useful even when the impure liquid feed contains dissolved salts which can precipitate and form scale on the outside of the condenser tubes and on the boiler walls at relatively high evaporation temperatures. Because scale deposits interfere with efficient heat transfer between the con-densing vapor in the tubes and the feed liquid in the boiler, it is undesirable to operate the system at a boiler temperature at which scaling occurs. Therefore, when sea water containing calcium sulfate, magnesium hydroxide, calcium carbonate, and the like, is the liquid feed, since these salts are more soluble in cold sea water than in sea water above about 160F, at temperatures above 160F scale will rapidly form on the hot tubes and condenser surfaces and will, in a short time, render the system operative only at very low thermal efficiencies.
Therefore, if sea water is the liquid feed, boiler tempera-ture (Tl) should be kept below 160F and preferably below 150F. The system can still treat very large volumes of liquid feed in an efficient manner by maintaining a vacuum in the boiler at a level such that the boiling of the liquid feed is accomplished within the no-scaling temperature limitations.

ll~Z442 The lower limlt of Tl is dictated by practical considerations since the system is uns~ited for treating solid feed. Therefore, for water feeds, Tl should never be below the freezing point of water at ambient conditions, which at 1 atm. is 0C ~32F) corresponding to a Pl under substantially saturated conditions of .006 atm. Tl is suitably at 33Y or above. Tl is preferably almost as high as the boiling point of water at 1 atm., which is 212F, e.g., at about 211F and 0.99 atm. For non-aqueous systems, which at 1 atm. boil above or below the boiling point of water, the preferred temperature limits of this system remain just above the freezing point to just below the boiling point. This is so even for so-called high boiling organic substances, which boil above 212F.
At the reduced pressure in the evaporator, even these type liquids boil at significantly lower temperatures and can be practically employed.
Under preferred circumstances highest volumes in gallonage are obtained when vapor is evaporated under saturated conditions at a vapor pressure less than one atmosphere. As a general matter,the lower the evapora-tion temperature, with the system in thermal balance, the higher the throughput volume and the higher the costs. Thus a water system utilizing an evaporator temperature of 170 to 211F produces an appreciable flow at relatively low cost. However, each system must be ... .. ... ... . ..
operated at evaporator temperatures and pressures, com-pression ratio , and the like, to meet the particular flow rate and cost requirements of each user. Therefore, depending upon whether a user desires to reduce operating costs at the expense of capital costs, or vice versa, - ~15;Z442 one or more systems can be operated together to yieI~
the desired flow rate and cost. The examples and data provided hereinafter are useful in making a choice of system parameter starting points necessary to meet a potential users needs.
Figure lA illustrates another embodiment of the prcsejnt invention which differs from the Figure 1 embodiment in the use of an independent second compressor 940 downstream of mixing section 925 and in the details of a motive power system 50 for furnishing hot, clean gases to injectors 922 and for driving the independent compressor 940. In the system of Figure lA, the T5, r P5 vapor from mixing section 925 may be further compressed in a substantially adiabatic fashion to increase its pressure to P6 and its temperature to T6. Since in this embodiment these pressure and temperature conditions, P6 and T6, represent the initial vapor conditions in the condenser tubes 906 as well, the compression ratio in compressor 940 is selected to provide a final pressure at least equal to ambient and to create the desired tempera-ture differential for effective heat transfer in the condenser tubes 906 from the condensing vapor to the feed water entering duct 902. Thus, one important purpose for includin~ an independent compressor in this system - is to provide great flexibility in operation at a relatively nominal cost, particularly where a motive system such . . . , - . ~ - .
as system 50 is operating to produce hot combustion gases - - for injection into mixing chamber 914. This flexibility is important to compensate for thermal imbalances which may occur in the system. Furthermore, steam injector load requirements may also be a factor that wiil make ~L~L52442 use of the independent compressor desirable, especially at low values of by-pass and/or low P2 pressure values, if difficulty is encountered in achieving the flow rates shown in the Tables and Examples. The cost per lO00 gallons when an independent compressor is used is higher than the cost values set forth in the tables and examples.
~rhis h~1gher cost, Cost IC~ may be calculated by using the following relationship:
Cost IC = (2-O.OlBP) X ~Cost from Tables) For example, using the first entry from Table I where Tl=207F, BP=12.7 and cost from the table=$0.15/lO00 gal-, the CostIc is CostIc= (2-0.01 x 12.7) x $0.15 CostIc= $0.28/lO00 gal.
Motive system 50, which may be a gas turbine engine, includes, a combustion chamber 52 wherein hot ;
combustion gases are produced, a turbine motor 54 operated by the hot combustion gases, and the compressor 56 linked to turbine 54 through shaft 58, shafts 60 and 62 linking compressor 56 through clutch and gear box 64 to independent compressor 940, and duct 66 for carrying the hot combustion gases to mixing chamber 914 through duct 936. Combustion chamber 52 is supplied by a compressed air duct 68 and ;
a fuel duct 70 through air and fuel injectors 72. The fuel to air ratio is maintained for complete combustion of all fuel. Preferably, the burning fuel is supplied with an excess of air through duct 68, which may use as its source a small compressor or super charger ~not shown) operated from shafts 58 or 60, so that the fuel burns to completion producing only carbon dioxide and r steam as clean combustion products. The clean combus-;Z4~2 tion gases together with the air drawn through compressor 56 operate turbine 54 and the combustion gas and air exhausting from the turbine exits by duct 74, controlled by servo-operated valve 76 which monitors the temperature in the space downstream of mixing chamber 914, and duct 66, which supplies clean combustion gases to the mixing chamber 914 through gas injectors 922. When the tempera-ture downstream of the gas injectors 922 becomes too high, valve 76 opens to divert some of the combustion gas away from the mixing chamber 914 until the tempera-ture stabili'zes to the desired level. An optiGnal com-bustion gas cleaning unit 67, shown in phantom, may be interposed along duct 66 to clean the gases in the event that comb~stion is incomplete or impurities enter the system with the fuel or air. Suitable gas cleaning units are well known and include, for examples scrubbers, electrostatic precipitators, chemical precipitators, and the like.
The independent compressor 940 need not, of course, be operated by a motive power system 50 as shown.
Instead, the compressor could be operated directly by motor means 941 (shown in phantom), such as electrical, gasoline or diesel motors. In such a case, if direct mixing of hot gases is to be used to supply the make up work to turbine 916, injector feed gas duct 936 could be connected to an alternative supply source for clean, . ..
hot gas, such as a pre-existing combustion gas source if system 900 were physically located near an industrial clean waste gas source, a separate fuel and air combustion gas generating source such as the combustion chàmber, fuel and air supply ducts and injectors shown in this ~ - 20 -~ 52~4Z
Figure, or, a steam production means with thermal energy supplied by burning inexpensive fuel, such as garbage, or by other suitable means. Alternatively, the use of hot gases to provide additional energy or work to the turbine can be entirely dispensed with and motor means 917 (shown in phantom) or any other thermal, electrical or mechanical energy source used to furnish the make-up work to turbine 916.
Figure lB illustrates still another embodiment of the present invention wherein the system of Figure lA is modified by adding thereto a third mixing section 948, similar to mixing sections 914 and 925, wherein vapor flowing in bypass arms 920 may be injected down-stream of independent compressor 940 through injectors 942. Such an arrangement provides a large degree of operational flexibility and permits continuous operation even under adverse conditions. Whether vapor flowing in bypass arms 920 is admixed with vapor expanding through turbine 916 in mixing chamber 925 through injectors 926 or with higher pressure and temperature vapor downstream o~ independent compressor in mixing chamber 948 through injectors 942 is controlled by bypass flow control valves 944 and 946, respectively. As in the embodiments of Figures 1 and lA, the additional energy needed to drive turbine 916 may be furnished from clean gas source 934 as thermal eneryy, from motor means 917 as mechanical ,,, , , . .;..
energy, or from any other suitable source. In a similar manner, independent compressor 940 may be directly driven through motor means 941 or may be driven in any other suitable way.
With the foregoing general descriptionof the ~ Z442 operation of a few embodiments of a single stage vacuum distillation-vapor compression system serving to set forth the fundamentals of the present invention, before other embodiments and variations are described, it will be useful to consider the following more specific examples of the oper~ation of the instant system. Accordingly, the following illustrative examples are offered by way of further explanation and are not intended to expressly or impliedly limit the scope of the invention.

EXAMPLE I
This Example, employing the embodiment of Figure 1, utilizes impure water as the feed liquid and assumes an initial boiler temperature Tl of 122F or 582R from which the initial vapor pressure in the boiler Pl can be determined from standard charts to be 1.789 psia.
The enthalpy of thé saturated vapor under these conditions is given by standard tables to be hl=1114 BTU/lb. The chosen compression ratio for variable compression ratio compressor 912 is 15:1, i.e., P2/Pl=15/1.
From the ideal gas law applied to adiabatic compressions and expansions and assuming that the heat capacities at constant volume and pressure, Cv and Cp, are constant, it is known that:

T2/Tl = ~P2/Pl) where b =~ and ~= Cp/Cv-Adopting the physical constants for water disclosed in U.S. 3,243,293 - Holden, b=0.2445, and substituting P2=15P
. ... .
and Tl = 582R into equation (1):

T2 = 582 (15)0.2445 = 1128R(668F) ~1~;24q~Z
Inasmuch as P2 = 15Pl; P2 = 26 836 psia. From the saturated steam tables it can be seen that at T2 = 668F, the saturation pressure is 2498.1 psia. Since the actual pressure, P2, is only 26.836 psia it will be appreciated that the steam is unsaturated. The enthalpy of unsaturated steam at T2 = 668F, P2 = 26.836 psia can be determined by interpolation in standard water and steam tables to be h2 = 1368 BTU/lb.
The demand work, WD = Wl, or work done by compressor 912 on the vapor is defined by the relationship:
WD = Wl = h2 hl where hl is the enthalpy of the uncompressed vapor at ~1 = 122F, Pl = 1.789 psia. Subs~ituting the known values of h2 and hl yields WD = Wl = 254 BTU/lb.
Upon exiting compresser 912, a portion of the compressed vapor at P2 proceeds through byp~ss arms 920.
This percent bypass (BP) or fractional bypass ~.OlBP) does not expand through turbine 916. Rather, it expands in a substantially adiabatic manner through injector nozzle 926 rom P2 to PBp. However, since the system downstream of turbine 916 is effectively open-to ambient, PBp = 1 atm, and the resulting t~mperature, TBp of the vapor exitin~ injectors 926 is given by the adiabatic fQrmula for ideal gases as:

BP T2 (PBp/P2)0 2445 Since PBp = 1 atm., P2 = i5Pl and T2 = Tl(15)0 2445, TBp becomes:
TBp = Tl (l/Pl)0 2445 At the same time the fraction of the compressed vapor which does not bypass the turbine expands through ~l~Z~42 the turbine to T4, P4. It can reasonable be assumed, in view of the direct shaft link between compressor 912 and turbine 916, that the expansion in the turbine will not exceed the compression in the compressor and, there- . .
.fore, that the limiting value of P4 is Pl and of T4 is Tl. Taking~the system at its limit, the vapor exha~sting turbine 916 is at T4 = Tl, P4 = Pl. This vapor is compressed in a substantially adiabatic fashion in the vent~ris in mixing section 925 to TR, PR. Since PR = 1 atm., TR can be calculatèd as follows TR = T4 (1/P4)0 2445 Substituting T4 = Tl and P4 = Pl, TR ~ Tl (l/Pl) Thus, TR = TBp and, irrespective of the value of BP, the temperature, T5, of the mixed vapor downsteam of mixing section 925 is T5 = T~ = TBp. For Tl = 582R
and Pl = 1.789 psia, T5 = 514F.
The enthalpy of the combined vapor stream at T5, P5 is denoted h5 and may be used to determine the bypas8 percentage, BP, for any Pl, Tl and compression r~tlo. Realizing that the enthalpy released by bypass vapors expanding through injectors 926 equals the enthaipy gained by the turbine throughput vapors compressing in-the mixing section venturis, and specifying the enthalpy released as .OlBP ~h2-h5) and the enthalpy gained as (l-.OlBP) (h5-hl,), and equating the enthalpy released to the enthalpy gained:
---. . BP = 100 (hs-hl)/h2-hl Substituting the known values for hl and h2 and determining h5 = 1295 BTU/lb from standard tables, BP = 71.3% s ~1524~2 EXAMPLE II
- In systems such as the one exemplified in Example I, it has been determined that the temperature of the vapor in the condenser, T5, exceeds the saturation tempera-ture for P5 = l atm. of 212F. This means that the heat released by the vapor in condensing, Qc' is greater than the heat of vaporization, Qv' with the res~lt that some fraction o the vapor, Fu, is uncondensed. This ~raction depends upon the quantity of surplus heat, Qs' released beyond the heat of vaporization,.or Qs Qc Qv Since Qc is the amount of heat released by the vapor at T5 and 1 atm. condensing and cooling to Tl(liquid), Qc ~ hs~hl~liq).
and Qv is the heat given up by the vapor at Tl condensing to a liquid at Tl, Qv hl-hl(liq).
Substituting for Qc and Qv' Qs 5 the fraction uncondensed, Fu = Qs/Qv' becomes:
Qs/Qv Fu ~ hs-hl/hl-hl(liq) Using the known values for hl and h5 and finding hl (liq) in the steam tables to be hl(liq) = 90 BTU/lb, the values f Qs~ Qc~ Qv and Fu can be calculated to be:
Q~ ~ 1205 BTU/lb.
. Qv = 1024 BTU/lb.
Qs = 181 BTU/lb/
.: . . ., , , , ;. , Fu = 0.177 . If a diverter line 932 (shown including a valve in Figure l) is junctioned into vapor return line 928 to permit the quantity of vapor passing into the condenser tubes 906 to be contolled so that only the amount necessary ~ 1~2~42 to vaporize the raw feed at Tl reaches the condenser, the remainder can be diverted to other uses. As a result, instead of only condensate alone being produced in the system, both condensate and superheated steam becomes available from the system.
Both the condensate and steam have a number of uses' Lor example:
(a) the condensate can be used for drinking water or for industrial purposes that require pure water;
(b) the steam can be used for heating or for producing electrical power;
(c) the condensate can be heated by the steam to any temperature up to the boiling point by indirect heat exchange;
(d) the steam can be condensed at little cost, e.g., by using a finned radiator cooled by air blown over it where the blower is powered by the motive power system;
(e) the steam can be diverted to duct 950 shown in Figure 2 for direct injecting into the raw feed in boiler 904 to heat the raw feed.

EXAMPLE III
To demonstrate that the instant system can in fact purify large volumes of impure water using equipment, specifically a condenser, of reasonable si~e and availablili-ty, it is assumed herein that compressor 912 can maintain the boiler pressure Pl at 1.789 psia by removing vapor therefrom as rapidly as it is produced. In this case, the rate of flow of vapor is solely dependent on the ; rate that the heat of vaporization is transferred to j the feed liquid. The heat of vaporization of water boiling ~ ~ at 122F and 1.789 psia is Qv = 1024 BTU/lb and the tempera-', ture difference between the condensing vapor and the feed liquid~at P5 = 1 atm. is ~ TLM- ~ TLM is the log mean temperat~re difference during condensation which together with the initial temperature of the impure liquid, I Tl, and the desired final distillate effluent temperature, j TD, determines the re~uired conaenser size.
LM max ~ min/ ( ~ max/ ~ Tmin) , ~ Tmax T5 Tl~ ~ Tmin=TD-Tl, and TD is selected 3 to be equal to or less than the vapor condensation tempera-J ture and greater than Tl The surface area A in square feet of a condenser j required to condense R gallons/hr of condensate at 122Fr~' having a heat of vaporization Qv of 1024 BTU/lb through a temperature differential Of 392F in a stainless steel condenser having a coefficient of heat transfer "h" of 250 BTU/hr - F - ft2 can be determined from the following relationship:
A ~ RQv/h ~ TLM
Rewriting Equation 11 in terms of R:
R = Ah ~ TLM/QV
j It is known that a conventional condenser unit, such as is ma~ufactu~çd by the Pfaudler Comp~ny of Rochester, New York, which is 5 feet long and 5 ~eet ! wide has an effective surface area for heat transfer of 2988 ft. . Therefore, the length L of such a unit ¦ necessary to provide A ft.2 of surface area is denoted ~ 24~z by the formula:
A/2988 x5=L
A=2988L/5 Inserting the aforementioned val~es for h, and A and assuming L=40' yields:
R _ 5,976,000 ~TLM/QV
At~TLM -I99F and Qv = 1024 BTU/lb. a flow of R =145,201 gallons/hr can be acc~modated and condensed.
, EXAMPLE IV
The cost to produce the flow R determined in Example III depends upon the make-up work, Wmu, which has to be done on the turbine. The makeup work is that fraction of the demand work, WD, which is lost when vapor proceeds through the bypass arms 920 rather than through the turbine:
Wmu = .OlBP x WD
In the case illustrated in the foregoing Examples I-III, .OlBP = .713 and WD = 254 BTU/lb. Substituting, we find that Wmu ~ 181.1 BTU/lb.
This work, Wmu, is the work that must be added to the system by direct driving the compressor- turbine through motor means 917 or by addition of hot gases through injectors 922, or otherwise. The cost can be determined by assuming that the cost to produce energy is about $2.70/ l,OOO,OOO BT~ Therefore, the cost/l,OOO
gallons to operate the present system is the cost of the make-up work. Expressing this in terms of percent bypass, we find:
Cost/l,OOO gal = 2.16x~0 4(BP) (WD) This works out, when BP=71.3~ and WD=254 BTU/lb are ~ 1~;244Z
substituted, to be:

i~ Cost/l,000 gal = $3.90 This cost value is, of course, idealized and does not take into account system inefficiencies. Therefore, 1 actual costs will be somewhat higher.
. EXAMPLE V
~; The values calculated by the methods described in Examples I-IV have been determined for other initial i temperatures ~Tl) and other compression ratios in turbine compressor 912. Table I shQws these values for a represen-tative sampling of Tl values at compression ratios of 2:1, 5:1, 15:1, 25:1, 50:1, lO0:1 and 200:1, although it will be appreciated that the only limitation on com-¦ pres~ion ratio i~ the availability o equipment, .., . . . ,., f ~5244Z
-~ ~ ~ ~D O u~ u~ ~ ~ o u~
o _ O _ O~ u~ ~ o o a~
O o o u~ ~ _ o o ~ ~ ~ o o Oo ~D

_ ~ _ ~ _ _ ~ oo o ~ _ ~oo oo a~ O ~ ~ o ~ O ~ ~ ~ ~D
~, ~o o o ~ o o ;~ o ~ U~ o ~ ~ ,~
_ LL. 00 ~
_ t~l N
- L C~ _ ~ uo~ uo~ o o _ o o o o _ o o ^ :~ a~ oo O ~0 0 cr~ o o oo . ~
oL ~ ~ ~ ~ ~ o o I ',, C~ u~ ~ ~ O ~ D O
. ;~ u~ o ~ ~ r~ oo ~ ~
, _ ~ ~ ~ r~ ~ ~ ~ ~ ~ ~ ~ ~ , V~ I~ ~ oo _ ~ _ _ ~ _ ~ _ U~ o C\' ~ ~ ~ ~, ~ oo _ r~
-_ ~ ~ ~ ~
1- ' ~ ~ o o a' a~ ~ o ~ O ~` 1`
_ _ oo u~ o ~ u~ O ~ ~ O
~ o~ ~ ~ O O ~ ~ O ~
I~ ~ o 1~ ~ _ _ o ~ -- _ ~ oo 1 :~ ~ O ~ I~ I~ ~ o r~ r~ _ ~ o ~D O
1~ a o ~ -- o o o ~ -- -- o o ~ ~

~_ _ ~ _ ~ ~ -- :J~ J ~J
IL1 ae ~ O
._ ~ , O a~
_ ~ I~ oo ~ ~ U~ ~ o 1~ ~ N _ -- -- ~1 ~ _ a~ _ o~ ~ _ a~
o ~ ~ ~ ~ ~ ~ ~ ~ ~ v~
--_ ' ~ oo ~ O 1~ oo O 1~ ~D 1~ ~ X
- - - - - - ~ - - -~ ~ ~ ~ ~ ~o - ~ ~ oo ~o ~ o oo u~ o ~o V Oo o~ ~ ~ ~ ~
r - - ~ ~ ~ - - ~ ~ ~ - - :
- - - - - - - - - - - - - - -O~ O 00 0 ~ ~ ~ O O~ O
~--~ ~ ~ o ~ ~ -- ~ oo o - l - - - - - - - - - - - - -O~ O ~ O _ ~ ~ O ~D
;~ ~ o -- ~ ;~
-- -- -- -- ~ -- o o -r~
~ a~ ~D O u~ D ~ ;i o lL ~ ~ ;t -- ~ O O O ;~
~ --c~ t~ ~ v~ ~ ~ ~ o o ,~ ~ c~ o a~ oo ~ ~ ;~ O O
~ - ~ - o l~ oo o ~ ~o l~ ~ o oo ~ ~
Q, ~ _ ~, o -- ~ _ ~ O O

IL _ 1~ <~ O -- O ~ a oo cr~ v~ o o ._ O _ o ~ ~ a~ _ ~ ~ ~ cr~ o ~D 00 ~5i24a~2 O ' o o ~o U`~ ~ _ -- O O ~ oO 1 _~ V~
r- ~~ ~ ~ I~ ~D O ~ O 1~ ~ ~ 1~
_~oo o CO V~ ~ o oo ~ r~ O O~ '`
_ C~ _ ~u~ ~D O ~ U~
~_ _ (~ N
O l_ N ~ N NO 2 L ~ A ~1 ~ O O r~
~ t~ O ~
.' y _ ~
,, . '.
a I~ O ~ O _ ~
,~ _ O ~~ U~ 1~ 00 ~ 00 In O~ ~ ~ _ ~ O_ ~00 ~ ~_ ~ ~ o a~oo 1~ ~ ~ oO ~
_~ ~_ U"~ ~_ o a~ a~ o o~ ~o ~ a~ a~ o~ o o _ _ _ _ _ _ g r~
_ O O _ ~ ~O O
_ _ _ _ _. __ _ _ _ _ _ ~ _ O 00 1~ ~ O~
_ jO O ~ ~ _ ~ ~ ~ O ~ O
ct~ a~ ~D ,~~ O ~ o~ ~ O _ ~ ~ cr~ r~
O l_ ~ J o oo u~
_ cr ~ ~~ ~ O r~

_ ~ ~ N ~ o ~ o ~ o o r~l ~ O 00 ~ ~ ~ ~D ~0 __ _ _ _ _ __ . _ _ _ _ _ ~ -- o~ $ ~ ' S _ _ _ _ o o _-- _-- O O O
o o u~ ~ o I.L ~ ~ ~ o ~ ~ O~ g ~ _ _ _ _ _ __ _ _ _ _ _ .. _ . . o U~ ~ g 8 8 8 o g 8 o g _ _ _ _ _ ,-- ;~ ~ ~ o ~ o ~ ~ o ~ l~ - o o - 4~ ~o - o o o ~ -- o o~ ~ ~ oo-- ~ o ~
O ~_ O ~ I~ O~ D 00 0 ~ ~D 1 ~ -- -- -- -- ---- -- ~

~5Z4~2 Substantially similar results as those attainable with the vapor treatment sections of the embodiments illustrated in Figures 1, lA and lB can be achieved with-out need for bypassing the turbine. In such a system, illustrated in Fig~re 15, bypass arms 920, injectors 926, valves 918 and mixing section 925 can all be eliminated and the system operated substantially as described.
Referring now to Figure 15, a vacuum distillation-vapor compression system is shown generally at 10. The system consists in its essential aspects of a boiler unit 12 including a condenser section 14 therein, a variable compression ratio turbine compressor 16 operated through shaft 18 by turbine motor 20, means for supplying make-up work to the turbine motor 20, and an independent second compressor 24 downstream of the turbine motor 20. The means for supplying make-upwork may include motor means, such as motor 28, (shown in phantom) which can be powered by electricity, gasoline, diesel fuel, and the like, directly linked through shaft 29 to turbine shaft 18 for directly driving the turbine. Alternatively, or in addition, the means for supplying make-up work may include a mixing chamber 22 upstream of the turbine motor 20 and means 26 for supplying hot gases to mixing chamber 22. Other well known techniques for supplying energy aan also be used, but are generally less desirable.
To understand the operation of the system 10, the path , ., , .. . ....
of raw feed, e.g., impure water, therethrough can be - charted. Initially, a starter motor, such as motor 28, is energized to rotate shaft 18 through clutch and gear box 30. Compressor 16 and turbine 20, which are linked to shaft 18, also rotate when the motor 28 is operated.

~15244Z
During start-up the variable compression ratio compressor 16 is allowed to rotate for a time sufficient for a vacuum to be drawn on the evaporative side of boiler 12. The compression ratio and the extent of the vacuum is predeter-mined, as will be seen hereinafter, based upon the desired operating parameters of the system and the temperature o~ the influent impure water and is controlled and monitored by variable pressure valve 32 in duct ~2 joining the boiler 12 and first compressor 16. Means 26 for supplying hot gases to mixing chamber 22, when supplied hot gases are the means employed for employed supplying make-up work, are operated to motivate turbine 20 to keep it running during start-up and to heat the tubes 34 in con-denser section 14.
In this embodiment, motive system 50, as previously described herein, constitutes means for supplying the hot gases although it will be appreicated that any known way of providing high temperature, high pressure gases, e.g., burning garbage at high temperature to produce high temperature, high pressure steam, may be used.
At the same time, motive system 50 may be used to operate independent compressor 24. The independent compressor 24 need not, of course, be operated by a motive power system S0 as shown. Instead, the compressor could be operated directly by electrical, diesel or gasoline motor mean, such as motor means 25 (shown in phantom).

.: ........................... " , . ,,." , Assuming the system of Figure 15 to include a mixing chamber and a hot gas source as the means for supplying make-up work to turbine 20, typical operation of the system will be better understood from the following description.

:~524a~z Feed water enters system 1~ through duct 38 and is rapidly heated to the boiling temperature, which depends on the vacuum level in the boiler, by heat trans-ferred from the condensing vapor in hot condenser tubes 34. Concentrated feed water waste, containing a large p~oportion of the impurities therein, is removed via dischar~e line 33. The vapor produced at Pl and Tl (the pressure and temperature in the boiler) is drawn through moisture separator 40 into duct 42 joining the boiler 12 and the first compressor 16 and is substantially adiabatically compressed by compressor 16 to P2 with a resulting heating of the vapor to T2. The heated vapor mixes with the hot, clean combustion gases emitting from injectors 36 in mixing chamber 22, which may be a mixing injector, mixing aspirator, jet mixer or any other config-uration known to be suitable for mixing vapors having different pressures in such a manner that a partial vacuum is created upstream of the actual mixing point The partial vacuum is useful in drawing the non-injected vapor into the mixlng chamber and thereby for enhancing the mixing. The temperature of the combustion gas is higher than the temperature of the heated vapor at this point although there is a substantially smaller flow rate of combustion gases than of vapor. The direct m~xing results in a substantially isobaric increase of vapor temperature by at least about 2K to T3 while ,.. ... ....
pressure remains subtantially the same, i.e., P3 equals P2. The mixed vapor-combustion gas stream substantially adiabatically expands through turbine 20 to reduced pressure and temperature P4 and T4 and, in so doing, does work W2 on the turbine to operate it. Since the ~1~i;24a~z turbine 20 and compressor 16 are directly linked by shaft 18, the amount of work W2 done by the vapor on the turbine is equal to the amount of work Wl done on the vapor by the compressor, i e., Wl equals W2. Inasmuch as the combustion gas serves primarily to heat the vapor and since the combustion gas flow rate is only a small fraction of the vapor ~low rate (e.g., about 125,000 gal/hr of vapor to less than 1~000 gal/hr of combustion gas), the work W2 is largely done by the vapor in a steady state condition. The expanded and reduced temperature vapor exhausting from the turbine 20 then passes through indepen-dent compressor 24 and is substantially adiabatically compressed to increase its pressure to P5 and its tempera-ture to T5. These pressure and temperature conditions, P5 and T5, represent the initial vapor conditions in the condenser tubes 34 as well. Therefore, the compression ratio in compressor 24 is selected to provide a final pressure at least equal to ambient and to create the desired temperature differential for effective heat trans-fer in the condenser tubes 34 from the condensing vapor to the feed water entering duct 38. The heat transfer temperature differential must be high enough that large volumes of feed water can be accomodated in this system within the practical limits imposed by reasona~le con-denser size. It is for achieving reasonable condenser size that the independent compressor is so important . ,. . " , ... .
in this embodiment, particularly where, as here, the _. _ ~ -- compression ratio of the independent compressor can be adjusted to accomodate variations in feed water flow rate and feed water temperature. Following condensation, purified condensate is drawn off through duct 39.

In an alternative operative embodiment, make-upwork may be furnished by motor means, s~ch as motor 28, the independent compressor may be directly driven by motor means, such as motor 25, and means 26 and the ,associated mixing and gas supply apparatus partially or totally~eliminated.

EXAMPLE VI
This Example utilizes impure water as the feed liquid and assumes an initial boiler temperature Tl of 20C or 293K from which the initial vapor pressure in the boiler Pl can be determined from standard charts to be 0.02307 atm. The chosen compression ratio for compre5sor 16 is lS:l,i.e., P2/Pl=15/1.
From the ideal gas law applied to adiabatic compressions and expansions and, assuming that the heat capacities at constant volume and pressure, Cv and Cp, are constant, it is known that:

2/Tl = ~ P2/Pl ) ( 1 ) .
where b = ~ and ~ = Cp/Cv Adopting the physical constants for water disclosed in U.S. 3,243,293 - Holden, b=0.2445, and substituting P2 = 15Pl and Tl = 293K into equation (1):
T2 = 293 ~15)0.2445 = 568.1K ~295.1C) Inasmuch as the work Wl done by the compressor 16 on the vapor is equal to the work W2 done by th,e vapor and hot combustion gases on the turbine 20, the following formulae result:
Wl = W2 ~2) Wl = Cv ~T2~Tl); W2 Cv ( 4 3 1~52442 Cv (T2-T~ cv(T4-T3) T2-Tl = T3-T ~ (5) In order to minimize system costs, we allow the hot ^ combustion gases to heat the output of compressor 16 only slightly to raise its temperature from T2 to T2+2.
Th~s, substituting T3=T2+2 in eq~ation ~5):
~,, T2-Tl = T2+2 T4 (6) T4 = Tl+2 (7) Using the known values of Tl and T2, we find:
T3 = 568.1+2 = 570.1K (287.1C) T4 = 293~2 = 295K (22C) The present system can permit T3 = T2+2 because the system places no constraint on the value of P4.
Thus the vapor temperature in the system increases from Tl = 293.K in the boiler T2 = 568.1K following substantially adiabatic compression to T3 = 570.1K follow-ing direct mixing with the combustion gases and then decreases to T4 = 295K for the vapor exhausting in the turbine motor.
The vapor pressure in the system increases from Pl = .02307 atm. in the boiler to P2 Z 0.3460 atm.
following substantially adiabatic compression, remains constant at P2 = P3 = 0.3460 atm. during substantially isobaric heating in the direct mixing chamber and decrease~
to P4 following substantially adiabatic expansion in the turbine according to ,the ,following relationship:
'T4 = T3 (p4~p3~b (8) ~''~~ which can be written as:
P4 =P3 (T4/T3) ( ) but since P3 = P2 and b=0.2445:
- P4 = P2 (T4/T3)1/b P4 = .02338 atm.
- The temperature T5 of the vapor following ~ 15Z44Z
adiabatic compression in the independent compressor 24 can be calculated by using the appropriate adiabatic compression relationship, similar to Equations (1) and (9), once the vapor pressure P5 or compression'ratio has been selected:
T5 T4 (P5/P4) (10) , Ap~lying Equation 10 to instances where P5 = 0.6 atm., 0.8 atm., 1 atm. and 2.5 atm. yields the following result:
P4 (atm)T4(K) ,T5 (K) T5 (C) 0.6 0.02338295 652.2 379.2 0.8 0.02338' 295 699.8 426.8 1.0 0.02338295 739.0 466.0 2.5 0.02338295 924.6 651.0 EXAMPLE VII
To demonstrate that the instant system can in fact purify large volumes of impure water using equip-ment, specifically a condenser, of,reasonable size and availability, the instance in Example VI where P5 z 1 atm. has been selected for further illustration. It is assumed herein that compressor 16 can maintain the boiler pressure Pl at 0.02307 atmospheres by removing vapor therefrom as rapidly a~ ,it is produced, In this ,case, the rate of flow of the vapor is solely dependent on the rate that the heat of vaporization is transferred . .. ....... .
' ' to the feed liquid. The heat of vaporization of water boiling at 20C (68F) and .02307 atm. is 1053.8 BTU/lb.
according to published tables and the temperature difference between the condensing vapor and the feed liquid at 1~;2~Z
P5 = 1 atm. is ~ TLM as defined in Example III. Selecting TD to be 200F and substituting T5 = 870.5F and Tl = 68F, we find ~TLM = 371-9F-The surface area A in square feet of a condenserrequired to condense R gallons/hr of condensate at 20C
(68F) having a heat of vaporization Hc of 1053.8 BTU/lb through a log mean temperature differential of 371.9F
~ . , .
in a stainless steel condenser having a coefficient of heat transfer "h" of 250 BTU/hr - F _ft2 can be determined from the following relationship.
A = R~C/h ~TLM (11) Rewriting Equation 11 in terms of R:
R = Ah ~ TLM/HC (12) Inserting the aforementioned values for h, TLM and Hc yields:
R = 11.029A (13) It i8 known that a conventional condenser unit, such as is manufactured by the Pfaudler Company of Rochester, N.Y., which is 5 feet long and 5 feet wide has an effecti~e surface area for heat transfer of 2988 ft.2 Therefore the length L of such a unit necessary to provide A ft2 of surface area is denoted by the formula:
~A/2988) xS ~ L ~14) A=2988L/5 (15) . Assuming a practical condenser length of 25 feet in Equation (15) indicates that a flow of R = 164,766 .... ... ... ... .
gallons/hr can be accomodated and condensed.
In systems such as the one exemplified in Example VI the heat of vaporization, Qv' is always less than the heat released by the vapor in condensing, Qc' because condensation always takes place at a higher temperature ~ 152~42 and pressure than vaporization. This means that some fraction of the vapor, Fu, is uncondensed. This fraction depends upon the quantity of surplus heat, Qs' released beyond the heat of vaporization, or Qs Qc Qv (16) Since Qc is the amount of heat released by the vapor at T5 and l atm. condensing and cooling to TD (liquid), Qc h5-hD (liq) where h5 and hD(liq) are the enthalpies before and after condensing and cooling. Since Qv is the heat given up by the vapor at Tl condensing to a liquid at Tl, Qv hl-hl(liq) Substituting for Qc and Qv in the expression for Qs and substituting for Qs and Qv in the expression Fu = Qs/Qv' we find F = [h5-hD(liq)-hl+hl(liq)]/hl 1( q Inasmuch as the enthalpies can be determined from the steam tables as h5 = 1468 ~TU/lb, hD(liq) = 168 BTU/lb, hl ~ 1091.2 BTU/lb and h1(liq) = 36.1 BTU/lb, Fu ~ 0.2321 If a diverter line 35 (shown including a valve in Figure 1) i8 junctioned into vapor return line 37 to permit the ~uantity of vapor passing into the condenser tubes 34 to be controlled so that only the amount necessary to keep the system in balance (0.7679 lbs./lb feed) reaches the condenser, the remainder (0.2321 lbs./lb feed~ can .. ... . . . .. . . .
be diverted to other uses. As a result, instead ~of only condensate alone being produced in the system, both condensate and superheated steam becomes available from the system.

llSZ9~42 Both the condensate and steam have a number of uses, for example:
(a) the condensate can be used for drinking water or for industrial purposes that reguire pure water;
(b) the steam can be used for heating or for produci,ng electrical power;
(c) the condensate can be taken off at any temperature up to the boiling point at 1 atm.
but not by indirect heat'lower than about 3 or 4F above Tl with the result, even if all condensate is taken off at 212F, that 0.7795 lbs of 212~ water/lb of feed water v~porized and 0.2205 lbs of steam at 870.5F/lb of feed water vaporized can be produced;
- (d) the steam can be condensed at little cost, e.g., by using a finned radiator cooled by air blown over it where the blower is powered by the motive power system. , EXAMPLE VIII
The output of the system of Example VI can be determined on the same basis a~ in Example VII for the instance wherein P5 is selected to be 0.6 atm. instead, of 1 atm. and the vapor temperature exiting the independent compressor is 379.2C (714.6F). As in Example VI, the l'iquid feed is presumed to boii at Tl =,20C,~68~) at a,pressure of . 02307 atm. and to have a heat of vapori-zation, Hc, of 1053.8 BTU/lb. The log mean temperature difference ~ TLM = 324.2F for a TD = 200F.

From equation 12, substituting the known values l~S2~42 of h,~TLM and Hc yields:
R = 9.614 A (17) Inserting Equation (15) for A in Equation 17 we get:
a R = 5745.3L (18) Assuming a practical condenser length of 25 ~eet in Equation 18 results in R = 143,633 gallon/hr condensate.

EXAMPLE IX
Example VI was repeated using a feed liquid consisting of impure water and assuming an initial boiler temperature Tl of 50C or 323K from which the initial vapor pressure in the boiler Pl can be determined to be 0.1217 atm. The compression ratio of compressor 16 is selected to be 15:1, i.e., P2/Pl = 15/1.
Applying Equation ~1~, T2 is 626.3K (353.3C), P2 = l5Pl = 1.8255 atmospheres. Making the same assumption as in Example VI with respect to isobaric mixing in the mixing chamber, T3 = T2~2 and T4 = Tl+2. Thus, the vapor temperature in the system increaseS from Tl = 323K in the boiler to T2 = 626.3K following substantially adiabatic compression to T3 = 628.3K following substantially isobaric mixing and decreases to T4 = 325K for the vapor exhausting the turbine motor.
The vapor pressure in the system increases - from Pl = 0.1217 atm. in the boiler to P2 = P3 = 1.8255 atmospheres during substantially adiabatic compression _ . ...
,, . ~ ._ . ~ . . , and substantially isobaric heating and decreases to P4, which can be determined from Equation (9) to be 0.1323 atmospheres, following substantially adiabatic expansion l~Z44Z
in the turbine.
Applying Equation (1) to instances where P5 = 0.6 atm., 1 atm. and 2.5 atm. yields the following result:
P5(atm) P4(atm) T4(K) T5(K) T5(C) 0.6 .1232 325 478.6 205.6 0.8~ .I232 325 513.5 240.5 1.0 .1232 32.5 542.3 269.3 2.5 .1232 325 678.5 405.5 EXAMPLE X
The output of the system of Example IX can be determined on the same basis as in Examples VII and VIII with P5 selected for illustrative purposes as:
~ a) 1 atmosphere;
(b) 0.6 atmospheres.
- The liquid feed is presumed to boil at Tl =
50C at a pressure Pl = .1217 atm. and to have a heat of vaporization, Hc, of 1024.0 BTU/lb. The log mean temperature differential, ~TLM depends on the selected P5. ~or each P5 selected, the ~ TLM and value of R
calculated from Equations (12) and (15) for a stainless steel condenser and assuming TD = 200F are as follo~ts:

. T5 TLM R (based on A) R(based on L) (a) 269.3C 90.70C(195.3F) 5.96A 3561.7L
:; ........................... .. . ........ . .
(b) 205.6C 70.06C(158.1F) 4.82A 2883.3L

.. . ~ ... . . .

~15~2442 Assuming L = 40 feet in order to get results comparable to the Tl = 20C cases, the condensate flow rate is calculated as follows:
(a) R = 142,468 gallon/hr (bl R = 115,332 gallon/hr EXAMPLE XI
This Example, employing the system of Figure 15, utilizes impure water as the feed liquid and assumes an initial boiler temperature Tl of 140F from which the initial boiler vapor pressure Pl under assumed saturated conditions is 2.889 psia. The compression ratio for compressor 16 is 15:1. Therefore P2 = l5Pl = 43.335 psia.
From the ideal gas law applied to adiabatic compressions and expansions, it is known that T2/Tl = ~ P2/Pl ) Solving for T2 and substituting:
T2 = 703 F
In this Example, all make-up work added to turbine 20 is provided by direct driving the t~rbine using an externally powered motor. Nevertheless, because of the direct shaft link between the turbine and the compressor, Wl, the work done by the compressor on the vapor equals W2, the work done on the turbine by the vapor plus the direct drive work added to the turbine.
- Wl = Cp (T2 q!~) , Choosing an average value of Cp = b. 4667 and .,, . ~ . ,." . ., . _ .
substituting for T2 and Tl:
Wl = 263 BTU/lb.
Since direct drive is used, the temperature of the vapor entering the turbine, T3, eguals T2 and ~ 1~2442 assuminy Wl=W2, the temperature and pressure of the vapor exhausting the turbine T4, P4 equals Tl, Pl.
Assuming that the independent compressor increases the vapor pressure to an ambient pressure of 1 atm.:
T5 = T4 (P5/P4) T5 =ATl (l/Pl) T5,= 433F
Adopting the equation for flow rate, R, from Example III and sùbstituting for A:
R = 2988Lh ~TLM/Qv. 2 Substituting h=250 BTU/hr-F-ft , L=40', TD=205F, TLM=151.4F and QV=1053 ~TU/lb at 433F:
R = 107,415 gal/hr.
Calculating cost ùsing Cp=.4667 and assuming that the cost to produce energy is about S2.70/1,000,000 BTU, we find:
Cost = Cp tT5-Tl) ~$2.70/1,000,000 BTU) Converting units into gallons and substituting y~elds:
$/1000 gal - $2.95 In still another broad form of the in~ention illustrated in Figures 16-18, many of the advantages of the already described embodiments are combined with the relative simplicity of conventional vapor compression systems to overcome the apparent ~hortcomings of such conventional systems. -.Vapor- compression systems are .
w~ll known for the treatment of impure liquids~ However, .. . . .
the system configurations heretofore known have suffered from serious disadvantages which have limited their usefulness.
For example, vapor compression systems are typically ~lS2442 designed to accept and treat a particular liquid, e g , salt water, entering the system within a narrow range of initial conditions. As a result, the system is incapable of being used for other liquids or for other initial conditions, and, therefore, its usefulness is limited.
In addition, conventional vapor compression systems must operate a~ low compression ratios, e.g. 1.2:1 to 1.5:1, to minimize cost. This means that the temperature in the evaporator must be close to 212F
because such low compression ratios prevent drawing any substantial vacuum in the evaporator. Moreover, since the specific volume of water vapcr decreases rapidly as temperatures drop below 212F and in view of the low compression ratios which must be used, if the vapor pressure exiting the compressor is to be high so that the vapor temperature may be high, the evaporator must j operate at or near 212F. ThiS effective temperature lim$tation considerably reduces the usefulness of the conventional vapor compression system by limiting the type~ of liquids which may be treated, by restricting the l~quids treated to a low solids content, and by precl~d~ng the distillation separation of liquids, such as oil and water, which is most easily accomplished at low temperature.
The embodiment of Figure8 16-18 provides an economical yet extremely flexible vapor compression ... ... . . . .; ................... ..
system, which is capable of high volume purification of--impure liquid sources; provides a vapor compression system capable of accepting as input a diverse selection of impure liquids over a broad range of influent liquid temperature and pressure conditions; and provides a vapor 3~1524~2 compression system which can be rapidly adapted to treat a diversity of impure liquids and which can utilize as an energy source available clean or dirty gases, or most r fuels, e.g., natural gas, jet fuel, methane, coal, garbage, etc., to generate such gases. Briefly stated, this embodiment comp~rises a method, and a system for practicing the method, for purifying large or small volumes of impure liq~id by evaporating the liquid in a boiler under a pressure not exceeding the saturated liquid vapor pressure, substantially adiabatically compressing the resulting vapor to a pressure substantially in excess of the vaporiza-tion pressure in a compressor capable of producing a variable compression ratio, and passing the resulting vapor through a condenser, such as the condenser side of the boiler, wherein the vapor will, upon condensing give up thermal energy to vaporize the feed liquid.
In an optional form of the invention, the compressed vapor is directed through and substantially adiabatically expands in a turbine before passing to the condenser.
The compressor is preferably driven by linking it to the shaft of an auxiliary turbine which may itself be driven by passing a volume of hot gas, e.g., combustion gas, steam, etc., therethrough. In one embodiment, the auxiliary turbine blading is annularly disposed with respect to the compressed vapor flow path and is driven by combustion gases produced in the annular space.
. .
Alternatively, the compressor may derive at least a ~ ~-~~ - portion of its power from motor means shaft linked directly thereto. The system of the present invention, because its operation is independent of the method of evaporation, e.g., vacuum or flash distillation are both suitable, l~S2442 is extremely flexible in terms of its utility and physical location. In the most common usage, the impure liquid is impure water and the system is able to furnish large quantities of purified water and, under some conditions, useful thermal energy as well.
Referring now to Figure 16, a vacuum distilla-tion-vapor compression system is shown generally at 110.
:,. . .
The system consists in its essential aspects of a boiler unit 112 including a condenser section 114 therein, a variable compression ratio turbine compressor 116 operated through shaft 120 and linked by the shaft to turbine motor 118, and means 1700 for supplying energy to operate compressor 116, i.e., energy not furnished by turbine 118. The energy supplying means may be hot clean or dirty gases, e.g. combustion gases, passing through the blading of an auxiliary turbine. In lieu of hot gases, or in addition thereto, the compressor 116 can be directly driven through shaft extension 122 by motor means 126, such as an electric or diesel powered motor, acting through motor shaft 122a and clutch and gear box 128 (shown in phantom). It will be appreciated, therefore, that the language "adding energy to the compressor" or similar expressions used herein are intended to contemplate any addition of enerqy, whether directly or indirectly to the compressor, where the effect of that energy is to operate or power the compressor.

., To understand the operation of the system 110, the path of raw feed, e.g., impure water, therethrough can be charted. Initially, a starter motor, such as motor 126, is energized to rotate shafts 120, 122 and 124 through clutch and gear box 128 and motor shaft 122a.

~ lS2442 Compre~sor 116 and turbine 118, which are linked to shaft 120, also rotate when the motor 126 is operated. During start-up, the compressor 116 is allowed to rotate for a time sufficient for a vacuum to be drawn on the evapora-tive side of boiler 112. The extent of the vacuum is predetermined, as will be seen hereinafter, based upon the desired operating parameters of the system and the temperature of the influent impure water and is controlled and monitored by variable pressure valve 130 in d~ct 132 joining the boiler 112 and compressor 116.
Referring to Figure 16, which is described using fuel combustion for producing hot gases as the means for driving an auxiliary turbine for adding energy to operate compressor 116, it can be seen that the impure liquid feed enters system 110 through feed duct 113 and is rapidly heated to the boiling temperature, which depends on the vacuum level in the boiler 112, by heat transferred from the vapor condensing in hot condenser tubes 114.
Unvaporized concentrated feed liquid, containing a large proportion of impurities therein, is removed from the boiler 112 through line 115. The vapor produced by boiling at Pl, Tl is drawn through moisture separator 129 and into duct 132 leading to turbine compressor 116. The pressure Pl is maintained in boiler 112 at a level not exceeding a pressure corresponding to Tl under saturated conditions by pressure regulating valve 130 dispo~ed in duct 132. The vapor is substantially adiabatically compressed at a ratio of from 1.2:1 to 250:1; preferably ",, ~.. . ... . . _ 3:1-250:1, more preferably 5:1 to 100:1 and still more - preferably 5:1 to 50:1, in compressor 116 to P2, T2 and, after leaving compressor 116, proceeds through turbine motor 118. The vapor substantially adiabatically expands . _ 49 _ ~ 15;244Z
thro~gh turbine 118 with a resultant pressure and tempera-ture drop to P3, T3 and then proceeds through vapor return duct 134 to condenser tubes 114 in boiler 112. The heat transfer temperature differential between the returning vapor at TF, i.e., the temperature of the vapor entering condenser tubes 114, and the feed water at Tl must be high enou~h that large volumes of feed water can be accom-odated in this system within the practical limits imposed by reasonable condenser size. The vapor condenses in tubes 114 giving up its heat of vaporization to the feed liquid entering the system through feed duct 113. Purified condensate may be removed from the system for general usage through line 136 Excess steam, if any, may be diverted through line 138 to keep the system in thermal balance, to heat the raw feed or to be injected into boiler 112,as will appear from a discussion of Figure 2, or for other purposes. If desired, the vapor in return duct 134 may pass through an optional independent compressor 140 (shown in phantom) where it is compressed in a substantially adiabatic manner to a pressure greater than ambient and at least greater than the saturation pressure of the liquid at Tl. Use of an independent compressor assures a continuously high pressure vapor flow into the condenser tubes, irrespective of opera-tional variations which may occur upstream thereof and reduces surges and eliminates any back pressure from ... . ........... . .
the condenser. The independent compressor 14~ may be ~ - dr;ven by hot gases operating a linked turbine (not shown) or by motor means (not shown), such as electrical, gasoline or diesel engines.
In this embodiment, the energy to drive compressor ~ 152442 116, in addition to coming from coaxial tur~ine 118, is furnished by a completely concentric auxiliary compressor-turbine combination surrounding and directly linked to compressor 116. In this configuration, the outer compressor-turbine combination supplies rotary power to the inner system to improve the performance of the inner system.
Extending rpm the spindle of compressor 116 and from the spindle of turbine 118 are shaft-extension members 122 and 124, respectively. Connected to shaft 122 are supports 1704 which rotate auxiliary compressor 1706 through its hollow spindle 1708. Connected to shaft 124 are supports 1710 through which shaft 124 is rotated by the hollow spindle 1712 of auxiliary turbine 1714.
The blades 1707 of auxiliary compressor 1706 and blades 1713 of auxiliary turbine 1714 are arranged in an annular space 1716 surrounding the inner compressor-turbine unit 116,118. The annular space 1716 is separated from the clean vapor flow space 142 by a solid partition 1701 and sealing rings 1702. In a preferred form of the inventlon, auxiliary turbine 1714 is operated by in situ produced combustion gases. Annular space 1716 operates a~ a combustion chamber into which fuel is admitted through injectors 1718 and air is admitted through space 1720. In space 1716 the fuel is mixed with air and igniters 1703 initiate combustion of the fuel and air.
The resulting hot combustion gases are mixed with air ..... ... .. .....
drawn into space 1716 via space 1720 and control valve 24 by rotation of compressor blading 1707, which air is compressed by compressor 1706 in passing therethrough.
After passing auxiliary turbine 1714, the hot combustion gases and compressed air exhaust through space 1722 and ~ 15244Z
never come i.n contact with the clean vapor which moves throu~h space 142 and return duct 134. As the comb~stion gases and air drawn into space 1716 pass through turbine 1714, they do work on the turbine blades 1713 causing turbine 1714 to rotate and to transmit power through supports 1710 ~o shaft 124, which power is utilized by coaxial compressor 116 in doing work on the vapors flowing in space 142 and by auxiliary compressor 1706 in com- .
pressing air drawn by it into space 1716. In an alterna-tive form of this embodiment, combustion or other gases from an external source may be drawn into annular space 1716 via space 1720 and valve 1724, in which case space 1716 need not operate as a combustion chamber.

The dirty hot combustion gases or other gases in space 1716 exhausting turbine 1714 still possess substan-tial thermal energy and are directed, for disposal or use, either through space 1722 or into heat exchanger section duct 1723 via duct valve 1725 and then through heat exchanger 1727. When passed into the heat exchanger 1727, heat rom the exhausting gases is transferred to the clean vapor in return duct 134. Since exhaust com-bustion gases are at a temperature in excess of 500F
and a pressure of 25 psia or greater, they can substan-tially increase the vapor temperature, T3, to T4 before the vapor enters the condenser tubes 114. In this way ... ... . ,, .. ~ .
the temperature difference in the condenser, TF-Tl, which .~.~;..~~ -- .in-this case is T4-Tl, is increased, thereby permitting the system to accomodate a greater flow rate or to minimize condenser size. The hot gases exhausting through space . 1722 can also perform useful work such as operating a ~ 1524~2 low pressure turbine (not shown) for driving optional independent compressor 140, heating the influent raw feed in a heat exchanger (not shown) disposed in d~ct 113 and/or heating the raw feed in evaporator 112 by means of heat exchan~e coils (not shown) in the evaporator.
In the case where the hot gas flowing through the auxiliary turbine 1714 are clean gases, such as steam, the clean gases can be injected back into the vapor in return duct 134 at a point upstream of condenser 114 or directly into condenser 114.
Additional flexibility can be built into the system by using variable ratio compressors and variable .ength telescoping condenser section tubing. The latter can be achieved using telescoping condenser tubes which can be telescoped to the desired condenser area by mechanical or hydraulic means. The former can readily be achieved in a number of ways, for example:
1) at least some of the compressor rotor blades can be made to telescope into and out of the spindle by mechanical or hydraulic means;
2) the airflow passage through the compressor - can be varied by varying the distance between the stator walls and the spindle using mechani-cal or hydraulic means;
3) at least some of the stators can be made to telescope into the walls by mechanical or ..... , ., - ,;, . ;
hydraulic means;

4) at least some of the compressor stages may be made to be declutched from the power supply shaft so as to offer resistance to vapor flow therethrough;

~ lSZ442

5) the compressor may be geared and clutched to the power supply shaft so that compressor speed can be varied Numerous modifications can be made to the auxiliary compressor-turbine configuration illustrated in Figure 16 to alter it and/or improve it for particular usa~es. ~hus, s~pports 1704 and 1710 could be formed into alr foil shaped fans to assist in the movement of large masses of gas. Still another modification involves clutching and gearing the outer co~pressor-turbine combina-tion to the inner compressor-turbine combination in order that the rate of rotation of the latter could be varied with respect to the former. Another useful modification is the addition of further compressor-turbine combinations in concentric rélationship to the two shown in Figure 16, all with the purpose of increasing the motive power available for compression in compressor 116 and of utilizing available energy sources, such as dirty combustion gases, in as economical a manner as is possible. ~he fundamental advantage of the configuration of Figure 16 is that it enables utilizat~on of as many different combustion gas sources and/or combustible fuels as may be available at the system location for supplying economical power to compress the vapor~ flowing in space 142.
An alternative and somewhat simpler embod~-ment of the present invention is illustrated in Figure ... . . , .; . ;
17 which shows a vapor treatment section simi,lar to the .corresponding section of Figure 16 except that coaxial turbine 118 and compressor-turbine shaft 120 have been eliminated. This configuration is especially useful where compressor 116 has a low compression ratio and ..
, ., ~ 15Z44Z
where the evaporator temperature Tl is about 212F and the influent raw feed temperature is relatively low.
In this type of system, it is desirable to operate the condenser 114 at a pressure somewhat above ambient in order to increase the rate of condensation therein.
When,comparing the operational and cost character-istics of~the sy~tems of Figures 16 and 17, it is noteworthy ~see Table II) that the cost for the Figure 17 embodiment increases as compression ratio increases, all else being equal, because increased energy is'required in the auxil-iary system to operate at the higher compression ratios.
However, higher flow rates are attainable in the compressor only form of the invention because the temperature differen-tial in the condenser is normally higher. On the other hand, in the Figure 16 embodiment, increasing the compression ratio does not increase operational costs because the coaxial turbine is able to extract more work from,the higher pressuré, higher temperature vapor exiting the compressor. In fact, since turbines are notoriously more efficient at higher pressures, increasing the com-pression ratio also increases the efficiency of the energy exchange in the turbine. However, the fixed costs of capitalization do increase,as the compression ratio -increases although even at high compression ratios the present system is anticipated to cost less than heretofore known systems taking into account system flexibility .: ~ . .. , . . . ;. .
and the like. A comparison of the relative effect of - using or omitting coaxial turbine 18 is detailed in Examples XII-XV.
A unique aspect of the Figure 17 embodiment resides in the optional ability to divert a portion of 1~5244Z
the P2, T2 vapor exiting compressor 116 to flow directly throu~h the auxiliary turbine blading to supplement and mix with the flow of combustion gases or other gases therein which normally drive the auxiliary turbine.
The effect of this diversion is to increase the shaft energy available to drive compressor 116 and thereby to increase the vacu~m drawn in evaporator 112 or increase ~he compression ratio or decrease the input of energy from an external source. Of course, diverting a portion of the compressed vapor will resul~ in lower flow rate of distilled, purified liquid. Howëver, the flow rate reduction may be an acceptable alternative for reducing the cost of operation per thousand gallons in cases where only relatively small flow rates are needed and where external energy sources to drive the auxiliary turbine are costly. To achieve the desired diversion of com-pressed vapor flow, a fraction of the flow, controlled by bypass valve 146 ~shown in phantom), is directed into conduit 144 (shown in phantom) connecting flow space 142 downstream of compressor 116 with annular flow space 1716. The diverted flow in conduit 144 passes through solid partition 1701 and is preferably injected into flow space 1716 using nozzles or injectors 148 (shown in phantom).
The systems illustrated in Figures 16 and 17, as with the embodiments described hereinbefore, are use-ful even when the impure liquid feed contains dissolvéd salts which can precipitate and form scale on the outside of the condenser tubes and on the boiler walls at relatively high evaporation temperatures. Therefore, if sea water is the liquid feed, boiler temperature (Tl) should be kept below 160F and preferabIy below 150F, by maintaining a vacuum in the boiler at a level such that the boiling of the liquid feed is accomplished within the no-scaling temperature limitations. It is very important to be able to evaporate at low boiler temperatures, particularly below 160F, a~range in which conventional vapor compression systems c3nnot operate.
The lower limit of Tl is dictated by practical considerations since the system is unsuited for treating solid feed. Therefore, Tl sh~uld ~ever be below the freezing point at ambient conditions of the liquid being treated, which for water feeds at 1 atm. is 0C (32F) corresponding to a Pl under substantially saturated conditions of .006 atm.. Tl for water feeds is most suitably at 33F or above. Tl is preferably almost as high as the boiling point of the liquid under ambient conditions, which for water at 1 atm. is 212F, e.g., at about 211F and 0.99 atm. For non-aqueous systems, which at 1 atm. boil above or below the boiling point of water, the preferred temperature limits of this system remain from just above the freezing point to just below the boiling point. This is so even for so-called high boiling organic substances, which boil above 212F.
At the reduced pressure in the evaporator, even these type liquids boil at significantly lower temperatures and can be practically employed. In a particular form .:........ , .................... ~ ,. , .;.~ , ,, of this embodiment which illustrates the advantages of - this embodiment over conventional vapor compression systems and the advantages of evaporating at low pre-ssures in the boiler, Tl is in the range from just above the freezing point, which for water feeds is 33F, to .

at least 10F below the boiling point, which for water feeds is conveniently about 200~F, and more desirably 33-160F. At these low temperatures, the compression ratio should be in the range 3:1 to 250~1 and desirably 5:1 to 250:1.
With~the foregoing general description of the operation~of a few embodiments of a single stage vacuum distillation-vapor compression system serving to set forth the fundamentals of the present invention, it will be useful to consider the following more specific examples of the operation of the instant system. Accordingly, the following illustrative examples are offered by way of further explanation and are not intended to expressly or impliedly limit the scope of the invention.

EXAMPLE XII
.
This Example, employing the embodiments of Figures 16 and 17, utilizes impure water as the feed liquid and assumes an initial boiler temperature Ti of 198F or 658R from which the initial vapor pressure in the boiler, Pl, can be determined from standard charts to be 11.058 psia. The enthalpy of the saturatéd vapor under these conditions is given by standard tables to be h1=1145 BTU/lb. The chosen compression ratio (CR) for variable compression ratio compressor 16 is 15:1, .... - ... . .. ... . .
i.e., P2/P1 15/1.
From the ideal gas law applied to adiabatic compressions and expansions and assuming that the heat capacities at constant volume and pressure, Cv and Cp, are constant, it is known that:

1~5Z442 where b =(~ and~ = Cp/Cv.
Adopting the physical constants for water disclosed in U.S. 3,243,293 - Holden, b=0.2445, and substituting P2=15Pl and Tl = 658R in to above equation:
T2 = 658 (15)0-2445 = 1~76oR(8l6oF) aInasmuch as P2 = 15Pl; P2 = 165.87 psia. From the steam tables it can be seen that at T2 = 816F, P2 . = 165.87 psia, the enthalpy of the compressed vapor can be determined to be h2 = 1435 BTU/ib.
The demand work, WD = Wl, or work done by compres-sor 16 on the vapor, is defined by the relationship:

WD = Wl = h2-hl where hl is the enthalpy of the uncompressed vapor at Tl = 198F, Pl = 11.058 psia. Substituting the known values of h2 and hl yields ' WD = Wl = 290 BTU/lb.
The final temperature, Tp, of the vapor reaching the condenser tubes, assuming no indepéndent compressor and valve 1725 closed, is TF = T2 = 816F in the Figure 17 embodiment where there is no turbine 118 present.
The final temperature, TF, where there is a coaxial turbine 118 present ~Figure 16), can be deter- ' mined from the followin,g expression for a substantially adiabatic expansion through the turbine:
TF = Tl ~PF/pl) ,",...

-,.,Assuming PF = 14.696 psia, and substituting known valves for Tl and Pl, TF = 245F
In the compressor only configuration, hF =

.

h2 = 1435 BTU/lb. In the eompressor-turbine configuration, hF at 245F and 14.696 psia can be determined from the steam tables to be 1166 BTU/lb.
The ener~y amount which must be added to the system, either through the auxiliary turbine or by direct driving the compressor, to power compressor 116 may be defined a~ m~ke-up work and designated as WMu. For the compressor only configuration, WMv - WD = 290 BTU/lb.
For the compressor-turbine configuration:
MU F hl Substituting the ~nown yalves for hF and hl:
WMu = 21 BTU/lb EXAMPLE XIII

To demonstrate that the instant system can in fact purify large volumes of impure water using equipment, specifically a condenser, of reasonable size and availabil-ity, it is assumed herein that compressor 116 can maintain the boiler pressure Pl at 11.058 psia by removing vapor therefrom as rapidly as it is produced. In this case, the rate of flow of vapor is solely dependent on the rate that the heat of vaporization is transferred to the feed li~uid. The heat of vaporization of water boiling at 198F and 11.058 psia is Qv = 979 BTU/lb and the effective temperature difference between the condensing .:............................ ., . , - .;, . , vapor and the feed liquid at PF = 14.696 psia is ~ TLM. ~
.TLM is the log mean temperature difference during condensation which, together with the initial temperature of the impure liquid, Tl, and the desired final distillate effluent 1iS244Z
temperature, TD, determines th.e required condenser size.

LM ~ max ~Tmin/ln~ ~5Tmax/ ~ T
max TF Tl~ ~ Tmin=TD-Tl, and TD is selected to be equal to or less than the vapor condensation temperature and. greater than Tl. For this Example, TD=205F. Calculating QTLM for Tl -~ 198F and TF = 816F for the compressor only embodim nt and TF = 245F for.the compressor-t~rbine embodiment yields a TLM = 290F for the compressor only embodiment and a TLM = 21F for the compress~r-turbine embodiment.
The surface area A in square feet of a condenser required to condense R gallons/hr of condensate at 198F
having a heat of vaporization, Qv' f 979 BTU/lb through an effective temperature differential equal to a TLM
in a stainless steel condenser having a coefficient of heat transfer "h of 250 BTU/hr - F - ft2 can be determined from the following relationship:
A = RQV/h ~TLM
Rewriting in terms of R:
R 5 Ah ~TLM/QV
It is known that a conventional condenser unit, such a~ is manufactured by the Pfaudler Company of Rochester, New York, which is S feet long and 5 feet wide has an effective surface area for heat transfer of 2988 ft.2, Therefore, the len~th L of such a unit necessary to provide A ft. of surface area is denot~ed by the formula: , A/2988 x5=L
... _.
~ A=2988L/S
Inserting the aforementioned values for h, and A, assuming L-40' and converting units to gal/hr yields:
R.= 747,000 a TLM/QV

, , 11~;2442At a TLM = 290F and 21F and Qv = 979 BTU/lb. The follow-ing flows can be accomodated and condensed:
Compressor only R=104,051 gal/hr Compressor-turbine R=16~028 gal/hr .
, ....
EXAMPLE XIV

The cost to produce the flows R determined in Example XIII depends upon the make-up work, WMU, which has to be done.
The work, WMu, is the work that must be added to the system by direct driving the compressor through motor means 126 or by addition of hot gases through auxiliary turbine 1714, or both, or otherwise. The cost can be determined by assuming that the cost to produce energy is about $2.70/ 1,000,000 BTU, Therefore, the cost/l,000 gallons to operate the present system is the cost of the make-up work. Expressing this in terms of make-up work, we find:
Cost/l,000 gal = 2.15xlO (WMu) This works ou~, for each of the Figures 16 and 17 embodi-ments, to be:
Compressor only Cost/l,000 gal = s6-26 .: ........................... , . , , - .; . . , Compressor-turbine Cost/l,000 gal = $0.45 T~is cost value is, of course, idealized and does not take into account system inefficiencies. Therefore, actual costs will be somewhat higher. Furthermore, all .

~15244Z
thermodynamic calculations assume an isentropic reversible process which is an approximation of a real process.

- EXAMPLE XV

The values calculated by the methods describèd in Ex~mples XIIXiV have been determined for other compression ratios.in compressor 116 assuming the same Tl = 198F
to show the effect of compression ratio on cost. Table II shows these values for water for compressor only (0) and compressor-turbine (T) embodiments and for a representative sampling of compression ratios of 2:1, 5:1, 15:1, 25:1 and 100:1, although it will be appreciated that the only limitation on compression ratio is the availability of equipment. Table II also shows a sampling of calculated data for temperatures (Tl) above and below 198F. For purposes of constructing the table, distillate effluent temperature, TD, is taken as 205F for each example in which Tl is 198F or less and as 210F for Tl above 198F.

.. . , . , .;, . . .

. . ~
.... ... .

S ' ~ 0 ~ ~ ~ ~ ( ~ ~ ~ ~ ~ ~ u~ ~ ~ ~
~ ~ o u~ o o o o o o o ~
~ ` ` ` ` ` ` ` ` ` ` ~ ?A,~
O ~D ~ ~D ~ ~ ~D U~C~

J U~
o u~ ~ ~ ~r ~ ~ ~ ~ o ~r o o O .
.. ~ C~ o ~ o ~ o ~ o a~ o ~

,~
h i~l ~1 ~ ~1 O E~ ~ ~ o ~ o .

o N
. ~ ~ ~ n D
E~ D O ~ ~ U~
m ~ O ~ a~
~ ~ ~ ~ ~ ~ ~P ~ U~' O ~ ~ ~ In ~ o~
-- ~
D O O
C~ u~ In~r ~r a~ o~ I~ r~ _ ~7 E~
O O ~ ~ u~
H O O 0~ 1. U~
H ~ ~t'~lt'~lt'~l~ ~u7 ~ ~D U:~
~: _~ ~1_1 O ~ ~ U~ U~ C~ ~ O' r~

U~ . . . . . . .

a: ~ ~ ~ u~ o o o~
. ~ I` ~ o~ r~ 1~ 1` ~1~ 1~ r-a~
E~
m u~
~ er ~ ~r ~ ~ ~~ ~ .
s . . _ . ._ . . .

_~ Cl~ ~ 0 a~ 0 a~
.,~ ~ o O O O O O OO O O
~ 7 ., ~ a~ 0 0 ~ ~ a~
O ~
._ .n ~ p~ o o ~ ~ ~ ~ o cs~
~ ~ 52442 .

~J ~ n ~ ~ a) ~ o ~ a:~
o u~ ~ I_ O~ o ~ a) c~
o o .
O ~ O ~ ~ I` ~ r~ O ~

~I 0 a~ a~ ~ o ~ u~ "
_, ~ _, ~ U~ ~ , ~ ~ ~ 0 ~ o t- ' ct~ ~ ~ co . ~D ~ ~ Y~ ~D

E~
u~
~ ~ a~
,C ~ ~ ~ ~ ~1 ,~ ,~ ,~

~ 1~u~ ~ o ~ u~
O r~
_1 _1 0 0 a) a 1~ 1 0 ~ ~ ~ ~ ~ ~
~3 ~ D ~1 --I

E~
l ~ ~ ~U~
H
H ~::
m c~ o ou~ u~
o~ D~D ~ ~ ~ ~ ~ N
I
N N
~ ~ ~ r~ ~ ~ ~ ~ I
P~ u~
O O
t~

U o o ~ U~ o oU~ U~
O O ~ ~ O O
a~
Q :' r`1--t~lt~1~ ~`1~ 1 ~ ~ ~ ~ o o o o ~ a~

..., al . ... , .. ,.;. ~, , Da~ C~ o t~ ~ ~ o ~ ~ ~ o o ~ ~ ~
s --l~ ~l ~ ~ ~ ~ ~ ~
. .. ta a~
0 0 a~ ~ ~ ~,q u~ 0 a~
,~ ~1 o o r~ ~ o ot~
. . . . . . . L~
I ~ O
-~ ~
al . _ ~ 0 0 ~ ~ ~
~ E~ c~ ~ ~ ~ ~ o o s o-- ~ ~
. . o ~Z442 It will be appreciated that the a~xiliary com-pressor-turbine system liOO is in reality no different than a conventional gas turbine, the exhaust or combustion gases of which are at a comparatively high temperature.
For this reason the embodiments of Figures 16 and 17, involving direct combustion in annular space 1716, are not desirably employed in a system wherein the impure li~uid is or contains highly volatile inflammables.
If the Fîgure 16 or 17 system must be used with volatile combustibles, sufficient insulation must be provided to thermally isolate the auxiliary system flow space 1716 from the combustible-containing flow space 142.
In addition, in lieu of direct combustion, where possible the auxiliary turbine 1714 should be driven by lower temperature combustion gases or steam supplied from external sources. The system of Figure 18, which is ?
described more fully hereinafter, is particularly well suited for this type application.
Figure 18 illustrates an embodiment of the present system which permits the use of virtually any available hot gases, whether they be clean or dirty, combustion gases or steam, to provide motive power for driving the auxiliary turbine 1714 and, in turn, through the shaft link, or driving the vapor compressor 116 as well. In this embodiment, the gases passing through the auxiliary system do not actually mix with the clean .: . . . . . - .; . - . , ;
vapor in the primary system, and, therefore, the purity of the condensate produced by the system is not compromised, e.g., by use of dirty combustion gases for additional motive power. In Figure 18 there is shown a configuration which includes either the compressor-turbine combination ~ , .

llS2442 116, 118 taught in Fi~ure 16 or only the compressor 116 taught in Figure 17 as the components of the primary or internal system. Thus the turbine 118 is optional and is shown in phantom. The system of Figure 18 in-cludes compressor 116 linked through shaft 120 to optional turbine 118 and shaft portions 122 and 124 projecting axially fr~m,the,spindles of compressor 116 and optional turbine 118, respectively. The system also incl~des a clean or dirty gas operated auxiliary turbine 1714 which consists essentially of a hollow spindle 1712 and blades 1713 attached to the outside-surface of the hollow spindle. The spindle 1712 is drivingly linked to shaft portion 124 through supports 1710. Shaft portion 124 is operatively linked with the spindle of optional turbine 118 which spindle is joined through shaft 120 to the spindle of compressor 116. Where optional turbine 118 ls not used, shafts 120 and 124 merge into a single shaft which is herein designated 124. In operation, the system is energized ,by starting motor 126 acting through shaft extension 122a and clutch 128. Auxiliary gas turbine 1714 i8 disposed with its blades 1713 arranged in flow space 1716, which flow space is annularly arranged with respect to primary or clean vapor flow space 142 and which is separated therefrom by a solid paritition 1701 and.sealing rings 1702. In this manner, the hot gases, which may be dirty combustion gases, are directed through . .. . ... . ......... .
'space 1716 to act on turbine blades 1713J which, through spindle 1712 and supports 1710, rotate shaft 124. The expanded auxiliary gases exhaust from the turbine 1714 into space 1722 in such a manner that they never combine or mix with the clean vapor in the primary system unless ~ 15~44Z
it is specifically desired to cause them to combine.
In Figure 18, particularly in the form thereof wherein turbine 118 is omitted, it may optionally be desirable to divert a portion of the P2, T2 vapors exiting compressor 116 to annular flow space 1716 to provide a portion of the motive power used to operate auxiliary turbine 1714, T~us, a fraction of the compressed vapor flow, controlled by bypass valve 146 (shown in phantom), is directed into conduit 144 (shown in phantom) connecting flow space 142 with annular flow space 1716. The diverted flow in conduit 144 passes through solid partition 1701 and is preferably injected into flow space 1716 using nozzles or injectors 148 (shown in phantom).
It will be appreciated that the vapor treatment embodiments hereinbefore described in Figures 16-18, which permit varying the initial parameters in the evaporator and compression means, allow the rapid and economic treatment of practically any impure liquid, The flexibility of the system, which contemplates evaporation in multi or single stage evaporators, whether by vacuum distillation or flash distillation, offers the greatest potential for dealing with present ecological needs while at the same time achieving rapid purification. Thu~ it is practical to build an installation wherein a number of evaporators, arranged in parallel, feed into a vapor treatment section to allow various influents to be brought into holding.

.: ........................... .. . . ..... . .
tanks associated with the evaporators, and any evaporator . brought on line at any desired time. It is also contemplated that one evaporator could be fed through flexible influent conduit that could be sectionally assembled to be as long as is necessary, for example several miles, to permit . - 68 -~ 52442 the drawing of infl~ent from offshore points at sea.
This will allow a land-based system to effectively and rapidly deal with chemical or oil spills in offshore regions. Conventional vapor compression systems, typically employing low compression ratios and necessarily operating near the boiling point of the liquid under ambient condi-tions, are n~ither capable nor flexible enough to deal , . . . .
with the many diverse influents and influent conditions for which high volume, rapid purification may be desirable.
Fiqure 2 illustrates a modification to the present invention which is equally applicable to all embodiments of the present invention, indeed to all vacuum and flash distillation systems. In accordance with this modification, a fraction of the compressed vapor returning to the condenser tubes 906 through duct 928 is diverted and directly injected into the boiler 904 where it mixes with the impure feed water therein, giving up its latent heat of vaporization and raising the temperature of the feed water in the boiler to Tl. This is particularly useful and important where the raw feed entering duct 902 is relatlvely cold, e.g., water at about 33-70F.
If the temperature in boiler 904 is maintained at such a low temperature, it is necessary for P1 to also be low for boiling to occur at Tl. However, it is very expensive to draw and maintain a high vacuum in the boiler and, rather than do so, it may be desirable to raise .. .. . , - .;, . - . , the raw feed temperature to a value at which the system .may be more economically operated. The expense of raising the raw feed temperature to Tl by diverting a fraction of the returning vapor and direct mixing it with the feed water is readily measured since whatever flow is : L152442 diverted does not exit the system as purified liq~id throu~h line 930. On the other hand; direct mixing in the boiler is a far more efficient menas of heating the raw feed than, for example, by diverting the returning vapor thro~gh an external heat exchanger in which it can heat raw fçed or by passing all the returning vapor through condense~ tubes 906, as in the other embodiments of this invention.
In Figure 2, tne details of the vapor treatment section of the system are not shown since this modifica-tion is equally applicable to all embodiments described herein. Hot vapor directed to the condenser tubes 906 through return duct 928 is at a temperature, Tf, and has an enthalpy, hf. A portion of this vapor is diverted through duct 950 and its associated valve 952 into ducts 954,956,958 and 960 and their respçctive valves 955, 957,959 and 961 for injection back into boiler 904.
Although four injection ducts are shown, it will be appreciated that any number of such ducts may, in practlce, be used. The remaining or undiverted vapor continues through duct 928 into condenser tubes 906 and exits the system as purified effluent through line 930. The fraction of the vapor which must be diverted to heat the raw feed can be calculated by assuming that the temperature of the impure raw feed liquid in feed duct 902 is To and its enthalpy is ho~ The enthalpy change required, per ... . . ...
pound of raw feed, to heat from To to Tl is (hl-ho)~ -; ~~ ~~--- In-order to produce this change, a fraction, FD, of returning vapor, e.g., steam, at hf must be diverted through duct g50 and admixed with the feed liquid, condensing in the process and having a final temperature of Tl.

~ 1~i;2442 For one pound of returning vapor, , the enthalpy change is hf-hl and the fractional change is FD (hf-hl). Since the enthalpy change in the condensing vapor must equal the enthalpy change of the raw feed, it can be determined that:
FD = hl~hO/hf ho From this relatiQnship the fraction of compressed vapor :,. . .
diverted from duct 92B into duct 950 can be determined for various raw feed temperatures and desired boiler temperatures. By similar well known techniques the flow rate of effluent, RD, which continues on through the condenser tubes and exits line 930 can be readily calculated.
An optional aspect of the system shown in Fi~ure 2 involves the use of return line 970 and associated valve 972 (shown in phantom) to divert a small portion of the flow exiting the initial compressor 912 back to raw feed duct 902 wherein it is injected through injector 974 (shown in phantom). In this way, the vapor injected through injector 974 will create a pumping effect in duct 902 to aid the feed of liquid therethroogh while, at the same time, heating the incoming feed liquid.
Line 970 is optional, although useful, because its contri~u-tion to the heating of the raw feed is small compared to the vapors injected directly into boiler 904 through ducts 954,956,958 and 960 and because the vacuum drawn by compressor 912 is generally adequate to draw the raw . . ... . . . ;.... . .
feed into the boiler.

- EXAMPLE XVI
An impure liquid feed having an initial tempera-ture of 198F was fed into the system of Example I using ~ 152442 a compression ratio of lS:l, Pl at Tl = 198F is 0.7524 atm. P2, T2 and T4 can be calculated and Qv' h5, h and h2 determined as in Example I. From these values it is found, using the methods of Examples II through IV, that:
R = 16,028 gal/hr $;Cost/I,OOO gal = ~0.45 In order to keep the cost constant, if the raw feed water is at To = 70F, it can be heated to Tl=198F by diverting a fraction of the vapor at Tf, which is T5 in Example I and mixing the diverted fraction with the raw feed water. This fraction, FD, is determinable from the relationship:

FD hl98(~ h70(liq)/hf-h70 to be, FD = 0.1135.
It can be calculated that Fu, the fraction of vapor uncondensed, under these conditions is only 0.021. Therefore, there is no surplus vapor available and the amount of vapor diverted will decrease the vapor flow, R, produced by the system by the factor (l-FD) to RD.
RD = R (1-.1135) RD ~ 14,209 gal/hr.
It can thus be seen that only a relatively small flow reduction must be suffered to provide the flexibility of handling raw feed at 70F for the same .: ........................... ., . . . . .. , ;
cost as raw feed at 198F. From this type o analysis, ..a table can be constructed as set forth in Table III.

~lSZ44Z

o O . U~ ~D In In U~ ~ ' O U~
_I ~ o ~ u~ ~ cn 0 ~ .
V~ o o _, _, o ~ o o o o ~ a~ ~ o o ~ co o ~r ~
o U~ ~ ~ ~ ~ ~ ~ o o a ~ ~ O.
~ er ~ ~1 L~
~: a. ~ ~ r ~ er o ~ o o~
0 0 ~ O O 1` ~ O
~, H 1~ h ~o ~ u7 ~ ~D ~o ~1 ~ u~
H . t~ P d' ~~11~ ~ ~_1 a~ . ,i a~ ~ ~ co O _~ O O ~1 ~ ~1 O O O O O O O O O O
~1 'I 1-- 0 O~ cn 0 o 0 0 r-- ~
O ~ O ~ D0 al o o _~
o E~ ~ ~ o o o o o o ~ _~

~ O o - ~ ~ o o o o o o U~

O

o a I ~ U~ O O O
,~ _I ~ _I_I ~ U~ o o ~1~;2442 It will be appreciated that the foregoing embodi-ments may be employ~d in conjunction with various type single and multi-effect evaporator arran~ements and various compressor-turbine config~rations. Some exemplary arrangements and configurations are illustrated and described in conjunc-tion with Figures 3-14.
Referring now to Figure 3, the impure liquid feed enters the shell side 102 of the heat exchanger-condenser unit 104 where it i5 heated by passage of partially condensed vapor through the condenser side 106 of the unit. The heated feed in liquid form exits the shell side 102 via feed line 108 and enters vacuum distillation boiler 110 which is maintained at a selected vacuum Pl controlled by pressure regulating valve 910 in line 911. The feed liquid is rapidly heated to boiling temperature Tl by vapor passing through and condensing in the condenser section coil 112 in boiler 110. The vapors pass out of boiler 110 through moisture separator 114 via vapor line 116 and then via line 910 and valve 911 and pas~ into the vapor treatment section of the system (not shown). It will be appreciated that any of the vapor treatment section configurations shown in Figures l,lA, lB or 15-18 may be employed in this embodi-ment of the invention. In the vapor treatment seotion, the vapors are substantially adiabatically compressed by compressor 912 to P2, T2 and then further treated ... ., , .. . ".~ , in the manner described in connection with Figures l,lA, -. lB.and 15-18, dependinq upon the vapor treatment section configuration employed, before entering condenser section coils 112 via return line 928. It should be understood that hot combustion gases may be used to provide make-~ 15~442 up work in the turbine and a motive system, such as system 50, may be used to drive the independent compressor when such a compressor is ~sed. Alternatively, both the make-up work and the energy for driving the independent compressor can come from motor means, such as motors 917 and 941, or from other suitable energy sources.
In the cond;enser section 112 the vapors condense at least partially, transferring their latent heat to the feed liquid entering the boiler 110 through feed line 108.
The almost completely condensed vapor exits condenser section coils 112 via line 120 and passes into the con-denser side 106 of unit 104. Controlling flow through the condenser side 106, which is preferably a jet con-denser having pressure and expansion chambers therein, i5 servo or spring controlled pressure valve 122 which serves to maintain the pressurç on condenser side 106 of unit 104 and to assure that all vapors condense there-~n. Excess steam may be diverted by line 121 so that the system remains in balance and too high a temperature does not develop in the feed water. Line 124 carries condensed vapor into storage container 126 from which pùre condensate may be drawn for general usage through line 128. Non-condensible gases exit via vent 130.
Concentrated waste liquid is removed from boiler 11 via line 111.
When hot gases are directed to the vapor treatment section through injectors 922, a portion of the vapor 5~~~~--- in-return line 928 may be diverted via line 135 to duct 936 and then through injectors 922 to furnish an increased vapor flow to the turbine 916. ~f line 135 is utilized, the turbine 916 should preferably have waterways to take 1~5~44Z
into account the possibility that in expanding the increased vapor thro~gh the turbine 916, a portion of the condensible vapor will in fact condense. The effect of diverting vapor flow through line 135 to turbine 916 is to increase the efficiency of the turbine by extracting as much work as possible from the vapor passing therethro~gh.
A multi-stage embodiment of the present invention, , ;.., embodying a vacuum distillation-vacuum compression system is illustrated in Figure 4. As In Figure 3, any vapor treatment section configuration shown in Figures l,lA, lB,and 15-18 or otherwise described herein, may be used.
A vacuum distillation-vapor acuum compression system, as is well known in the art, has the advantage that, due to the multiple distillation stages, it can be con-structed using equipment which is significantly smaller than would be required with a single stage system. In addition, a multi-stage system is substantially more flexible in usage than is a single stage system and, by appropriate location of the valves,-one or more of the stages can be shut down during slack times, thereby producing a smaller quantity of distillate and permitting the cleaning and/or repair of stages which are not then in use. Multi-stage units are conventionally employed in flash distillation plants which usually require large bodies of cooling water, such as sea water, for efficient operation. The employment, as shown in Figure 4, of ~. ~ ... . .. . ;. . , a multi-stage system in a vacuum distillation embodiment has the advantage that it requires no large bodies of cooling water and can, accordingly, be located many miles from large bodies of water. Operating conditions for the multi-stage embodiment are substantially the same .

- 76 - , Z4~Z
as for the single stage embodiment with acceptable tempera-tures for water in the boiler (Tl) as low as about just above 32F, e.g., about 33F, corresponding to a pressure (Pl) of about .006 atmospheres and as high a temperature as is consistent with avoiding scaling in the boiler, where appropriate, ~hile at the same time maintaining an efective temperature difference between the vapor in the condenser return line and the condensing tempera-ture (Tl) in the boiler such that the system can effectively treat large volumes of impure feed liquid. Although the precise tempera~ure and pressure will vary from stage to stage by small amounts, as a general matter, the pressure and temperature is maintained substantially the same in all evaporation stages.
In the system designated by the numeral 200 depicted in Figure 4, the impure liquid feed enters the shell side 202b of the heat exchanger-condenser unit 202 where it i~ heated by passage of partially condensed vapor through the condenser side 202a. The heated feed exits heat exchanger-condenser unit 202 via line 203 and enters the shell side 204b of another heat exchanger-condenser unit 204 where it i5 further heated by passage through the condenser side 204a of additional partially condensed vapor. In a similar manner, the feed liquid is successively heated by passage through the shell sides of heat exchanger-condenser units 206,208 and 210. In each of these units heat is transferred to the feed liquid from partially condensed vapor passing through the condenser side 202a, 204a, 206a, 208a and 210a of the units and through lines 203, 205, 207 and 20g interconnecting the shell sides of the successive heat exchanger-condenser ~ - 77 -~ 15Z4~2 units. Finally, the heated feed liquid exits the shell side 210b of heat exchanger-condenser 210 thro~gh feed line 212 and enters multi-stage vacuum distillation boiler charnber 214 wherein it is heated to boiling in each of the stages 216, 218, 220 and 222 of the multi-stage chamber.
In chamber 214 the feed flows over and under a pl~rality of baffles 2~4a,i224b, 224c, 224d, 224e and 224f through all of the evaporation spaces until unevaporated concentrated li~uid feed containing the great bulk of impurities in the feed exits the multi-stage chamber 214 via line 226.
The pressure within the evaporation space in m~lti-stage evaporation chamber 214 is maintained substantially at Pl and Tl by pressure regulating valve 911, which may be a spring or servo-controlled valve. The vapor produced in chamber 214 at Pl, Tl exits the stages 216, 218, 220 and 222 through moisture separators 215 and vapor exit lines 230, 232, 234 and 236 respectively. The vapor r@combines in vapor discharge line 910 which directs the vapor through pressure regulating valve 911 and into the vapor treatment section wherein it is substantially adiabatically compressed in compressor 912, and then further treated in the manner described in connection with Figures 1, lA, lB and 15-18, depending upon the vapor treatment section configuration employed, before entering multi-stage evaporation chamber condenser section coils 242 via return line 928. It should be understood ... . .. ..... .
that hot combustion gases may be used to provide make-yp work in the vapor treatment section turbine and/or a motive system, such as system 50, may be used for driving the independent compressor, when such a compressor is used~ Alternatively, both the make-up work and the energy (` ( 1~5i2442 for driving the independent compressor can come from motor means, such as motors 917 and 941, or from other suitable energy sources. In the condenser section, the vapor is at least partially condensed, transferring its latent heat to the heated feed liquid entering the chamber 214 via feed line 212. Excess steam may be diverted through line 241 to keep the system in thermal balance.
;-, . . .
The almost completely condensed vapor is tapped from condenser coil 242 in each of the stages 216, 218, 220 and 222 via condensate return lines 244, 246, 248 and 250 and led to the condenser sides 202a, 204a, 206a, 208a, 210a, of heat exchanger-condenser units 202, 204, 206, 208 and 210 wherein the vapors completely condense giving up their remaining heat to the feed liquid passing through the shell sides of these units. Flow is cGntrolled through the condenser sides of the heat exchanger-units, which are preferably jet condenser units having pressure and expansion chambers therein, by servo or spring controlled pressure valves 252 in each of the condenser units, which valves serve to maintain the pressure on the condenser slde and to assure that all vapors are condensed therein.
The cooled condensate exits the condenser side of units 202, 204, 206, 208 and 210 via line 254 and its respective branches and is directed to storage tank 256 from which pure condensate may be drawn for general usage through line 258. Non-condensible gases exi~ via vent 260.
Inasmuch as the iiquid feed flows serially through the various stages 216, 218, 220 and 222 of the evaporation chamber 214, the feed liquid becomes more and more concentrated as it flows from feed line 212 toward concentrated liquid discharge line 226, thus --~Z44Z
increasing the possibility of scaling in evaporation spaces 220 and 222 as compared with spaces 216 and 218.
Proper control of the pressure and temperature in the multi-stage chamber 214 via valve 911 however, can avoid scaling. Another means of avoiding this increased likelihood of scaling is by modifying chamber 214 in s~ch a manner that the ~affles extend the entire height of the chamber 214, thereby defining enclosed evaporative spaces and by adding feed lines directly from the shell sides of heat exchanger-condenser units 202, 204, 206, 208, and 210 to each evaporative space so that fresh raw feed passes directly into each evaporative space independent of each other evaporative space.
As has been hereinbefore indicated, the instant invention is equally applicable to flash distillation as the evaporative mode for forming the vapor in the system. The embodiments of Figures 5 and 6 are generally directed to flash distillation-vapor compression m~lti-stage systems. As is well known, in conventional multi-stage flash distillation systems the flash chambers are interconnected with baffles and weirs to permit the flow of distilland from the first to the last flash chamber and each chamber is operated at a successively lower temperature and pressure than the preceeding chamber.
As a conseq~ence, each of the lower temperature and pressure stages are significantly less efficient than the first .. , . .,, . .. . ;., , , ;
flash distillation stage, which i5 one disadvantage of flash distillation systems. For example, U.S. Patent No. 2,759,882 discloses a seven stage combined flash distillation and vapor compression evaporator wherein it is disclosed that of the 8.2 lbs of distilled water , - 80 - t ' ' ' ' ' - ' t 11~2~42 prod~ced by the seven stages, the first stage produces 4.2 lbs. and the remaining six stages together only pro-duce an additional four pounds, with the average efficiency of the last six stages about l/6th the efficiency of the first stage. This disadvantage of multi-stage flash distillation systeTns is overcome in accordance with the present invention by maintaining the temperature and pressure at the same level in each of the flash chamber stages so that a high volume flow of distillate can be achieved. It is noteworthy that a characteristic of flash distillation systems is that the boilers do not contain heating means and, therefore, scaling of the heating means is not generally a problem. Of course, localized scaling is possible due to localized hot spots.
However, this can generally be eliminated by maintaining the flow of feed liquid therein reasonably rapid so that heat is absorbed and dissipated as fast as it is formed with the result that hot spots are substantially eliminated.
Turning now to the embodiment of the invention shown in Figure 5, the numeral 300 designates generally a flash distillation system into which impure liquid feéd is fed and purified condensate is removed in an economical fashion. The raw liquid feed enters the shell side 304 of the heat exchanger-condenser unit 302, which is preferably a heat exchanger-jet condenser unit, in which the liquid feed is heated by the passage of partially ~ , !
condensed vapor through the condenser side 306 of the unit. The heated feed in liquid form exits the shell side 304 through feed line 307 and is passed to t~e tube side 308a of heat exchanger 308 where it is heated by hot vapor conaensing in the shell side 308b. Th heated 11524~2 feed passes through line 309 directly into flash chamber 314 where it flashes under the reduced press~re Pl into the evaporative space above the liquid and flows as a vapor thro~gh moisture separator 348 and line 320, combined vapor line 910 and valve 911 to turbine compressor 912.
Valve 911 is a^pressure control valve which regulates the press~re at Pl within each of the flash chambers 314, 316 and 318. The f,eed liquid which does not flash in chamber 314 exits the chamber thro~gh line 311 and enters the tube side 310a of heat exchanger 310 wherein it is heated by the flow of condensing vapor in the shell side 310b, which condensing vapor entered the shell side of heat exchanger 310 through line 332 from heat exchanger 308. The heated feed exits heat exchanger 310 through line 313, flashes in flash chamber 316 under reduced pressure ~Pl) and flows as a vapor through moisture separator 348, line 322 and combined vapor line 910 to the vapor treatment section.In a similar manner, the u~evaporated heated feed passes from flash chamber 316 -through line 315 into the tube side 312a of heat exchanger 312 wherein it is further heated by vapor from heat exchanger 310 through line 334 condensin~ in the shell side 312b.
The feed continues through line 317 into flash chamber 318 where it is flashed at pressure Pl into vapor, passed through moisture separator 348 and led by vapor line ~24 into combined vapor line ~10 and then to the vapor treatment section. Any unflashed liquid feed exits the-sy~stem as concentrated waste through line 319. The combined evaporated vapors in line 910 passin~ valve 911 at pressure and temperature Pl, Tl are directed into the vapor treatment section of the system. It will be appreciated that any of the vapor treatment secti~n config-urations shown in Figures 1, lA , lB or 15-18 may be employed in this embodiment of the invention. In the vapor treatment section, the vapors are subtantially adiabatically compressed by compressor gl2 to P2, T2 and then further treated in the manner described in connection w,ith ~igures 1, lA, lB and 15-18, depending upon the vapor section configuration employed. It should be ~nderstood that hot combustion gases may be used to provide make-up work in the turbine and a motive power system, such as system 50, may be used for driving the independent compressor when such a compressor is used.
Alternatively, both the make-up work and the energy for driving the independent compressor can come from motor means, such as motors 917 and 941, or from other suitable energy sources. The compressed vapor returns to the shell sides of heat exchangers 308, 310 and 312 via return line 928 and lines 332 and 334, and is directed from the last heat exchanger shell 312b through line 336 into the condenser side 306 of heat exchanger-condenser unit 302. Controlling flow through the condenser side 306 is spring or servo operated pressure valve 346 which serves to maintain the pressure on condenser side 306 and to assure that all vapors are condensed therein.
Line 338 carries condensed vapor into storage container 340 from which pure condensate may be drawn for ~eneral .... , . - .; - , ;
usage through line 342. Non-condensible,gases exit via ,,vent 344. Excess steam may be diverted from return line 928 through line 331 to keep the system in thermal balance.

A preferred form of flash distillation-vapor compression system is illustrated in Figure 6~ In the .- ..

~15Z44Z

system of Figure 6, designated generally as 400, the raw liquid feed separately enters the shells 402b, 404b, 406b, of heat exchanger condenser units 402, 404 and 406, which are preferably heat exchanger-jet condenser ~nits. In the heat exchanger-condenser ~nits, the raw feed is heatedAby the flow of partially condensed vapor through tbe condenser side 402a, 404a and 406a of the units. The partially heated feed passes out of the units 402, 404 and 406 through feed lines 408, 410 and 412, respectively, into the tube si~es 414a, 416a and 418a of heat exchanger units 414, 416 and 418. In these heat exchanger units, the feed is further heated by the condensing vapor entering the shell sides 414b, 416b and 418b of the heat exchanger units through vapor return lines 444, 446 and 448. The heated feed from each of the heat exchangers enters its respective flash chamber 426, 428 and 430 through feed lines 420, 422 and 424, respectively. The heated feed flashes under the reduced pressure Pl at a temperature Tl in each of the flash chambers. Any unflashed concentrated waste is removed from the flash chambers through lines 427, 429 and 431, respectively.
The flashing vapor passes moisture separators 425 and is collected in vapor lines 432, 434 and 436 and combined vapor line 910 and is passed through pressure control valve 911 into the vapor treatment section.Valve 911 regulates the pressure in each of the flash chambers .
426, 428 and 430 to Pl. The vapors passing valve 911 pass into the vapor treatment section of the system.
It will be appreciated that any of the vapor treatment section configurations shown in Figures 1, lA, lB or 15-18 may be employed in this embodiment of the invention.

- ;

~5244Z
In the vapor treatment section, the vapors are substantially adiabatically compressed and then further treated in the manner described in connection with Figures 1, lA, lB and 15-18, depending upon the vapor treatment section configuration employed.
It should be understood that hot combustion gases may be, used to provide make-up work in the turbine and a motive power system, such as system 50, may be used for driving the independent compressor when such a compressor is used. Alternatively, both the ma~e-up work and the energy for driving the inaependent compressor can come from motor means, such as motors 917 and 941, or from other suitable energy sources. ~he compressed vapors return to heat exchangers 414, 416 and 418 through combined vapor return line 928 and then through individual vapor return lines 444, 446 and 448 to the shell sides 414b, 416b, 418b of the heat exchangers where the hot rqturning vapor~ at least partially condense, transferring their latent heat to the feed liquid on the tube sides of the respective heat exchangers. The almost completely condensed vapor exits the heat exchangers through lines 450, 452 and 454 and flows into the condenser side 402a, 404a and 406a of units 402, 404 and 406 wherein further condensation takes place and the heat thereby given up is .transferred to the entering raw liquid feed. Controlling flow through the condenser sides 402a, 404a and 406a are spring or servo-operated pressure valves 470 whicb serve to maintain the pressure on the condenser sides , . ~. ., . ~ . . . .
of units 402, 404 and 406 and to assure that all vapors are condensed therein. The condensate is carried through condensate return lines 456, 458 and 460 and combined -- . .

~ 15Z442 condensate return line 462 into storage container 464 from which pure condensate may be drawn for general usage through line 466. Non-condensible gases exit via vent 468. Excess steam may be diverted from return line 928 thro~gh line 443 to keep the system in t}-ermal balance.
I'he ~arallel-parallel embodiment of flash dis-tillation-va~or ~ompression system shown in Figure 6 is probably the most efficient type because concentration of waste can be individually adjusted from each flash chamber by adjusting the feed flow into each chamber.
In addition, different types of raw feed having a common carrier solvent, e.g., water, can be introduced into each chamber and valuable by-products can be separated from the common solvent of the feeds in each chamber and separately recovered.
The invention has thus far been described in its simplest forms and has, in each embodiment, utilized but a single turbine compressor operated by a single turbine motor. However, the configuration of the turbine compressor 912/turbine motor 916 need not be as simplistic as shown in Figures 1, lA, lB or 15-18. Rather, consider-able flexibility can be introduced into the system if the compressor, the turbine, the compressor-turbine combination or the compressormixing chamber-turbine combina-tion is configured to meet the requirements and demands of the particular system. For illustrations of particular arrangements which are useful and are generally operable --- in the systems shown in Figures 1, lA, lB, 2-6 and 15-18 attention is invited to Figures 7-14 and the descrip-tion thereof which follows in which the numerical designations of Figures 1, lA and lB have been used for ~lSZ442 convenience and in which it has been assumed that make-up work is supplied, at least in part, by direct mixing of hot gases. It will, of course, be appreicat~d that Figures 7-14 are equally applicable in conjunction with the other embodiments and/or where no hot-gas make-up work is utilized.
~ Referring first to Figure 7, there is illustrated schematically a clutched compressor unit designated by the numeral 500, which unit may be used in lieu of turbine compressor 912. The clutched compressor unit 500 is operated by a turbine 916 (partially shown) and includes a first compressor 502 having a compressor spindle 504 and a second compressor 506 having a compressor spindle 508 which is substantially larger than is spindle 504.
Spindles 504 and 508 are linked through shaft 510 and clutch 512. Clutch 512 can be a variable clutch which causes the smaller spindle to rotate at a different velocity than the larger spindle, i.e., clutch 512 may be a variable gear box generally similar to an automobile transmission, which permits the compression ratio to be varied at will.
Such a system is valuable as an aid in adjusting system operating variables depending upon the density of the vapor and the need to increase or decrease the flow rate through the system.
Figure 8 illustrates two turbine motors operating a single turbine compressor through a clutch and gear .: ........................... , ., , .. . ;. ~ . , box. Compressor 530 has its spindle 532-linked through ~ shaft S34 to clutch and gear box or transmission gear box 536. Shafts 538 and 540 link gear box 536 with turbine spindles 542 and 544 of turbines 546 and 548.

In operation, starting motor 550 acting through shaft llS2442 extensio~ 552 and clutch 554 s.tarts spindle 532 of compressor 530 rotating. Power is transmitted through shaft 534 to gear box 536 and, through shafts 538 and 540, spindles 542 and 544 of turbines 546 and 548 are also caused to rotate. Hot, clean cornbustion gases are mixed with the vapor flowing ~hrough space 556 as the gases are emitted into spac~e;556 through injectors 558. The combined vapor flow and combustion gases transmit rotary power to turbines 546 and 548 and through transmission gear box 536 to compressor 530. A particular advantage of this configuration is that it is more flexible than two separate compressor-turbine combinations and, at the same time, more economical.
Figure 9 illustrates a single turbine motor having a spindle 602 linked through shaft 604 to gear box 606 which gear box is directly linked through shafts 608 and 610 to the spindles 61.2 and 614 of compressors 616 and 618. In operation, starting motor 620 operating through shaft extension 622 and clutch 624 starts spindle 612 of compressor 616 turning and, in turn, causes compressor 614 and turbine 600 to also rotate. Hot, clean combustion gases are mixed with the vapor flowing through space 626 as the gases emit from injectors 628. The combined vapor flow and hot combustion gas flow motivates turbine.
600 which, through gear box 606, can operate either or both of the compressors 616 and 618. This configuration has advantages similar to those of the configuration .. ~. .. . .. .;.~ . - "
illustrated in Figure 8.
Figures 10 and 11 illustrate embodiments of the compressor-turbine combination which permit the use of hot, dirty combustion gases in addition to hot, clean combustion gases to provide additional motive power for llsz44z driving the turbine and, in turn, through the linked shaft, for driving the vapor compressor as well. In these embodiments, the hot, dirty combustion gases do not actually mix with the vapor in the system, and, there-fore, the purity of the condensate prod~ced by the system is not compromised by use of dirty combustion gases for additional motive power. Referring first to Figure 10, there lS shown a configuration which includes the con-ventional compressor-turbine combination and a mixing chamber for mixing hot, clean combustion gases with the vapor flowing through the turbine and the compressor.
In addition, the unit illustrated in Figure 10 includes a hot, dirty combustion gas driven turbine which increases the shaft power available for driving the compressor.
The unit of Figure 10 includes compressor 91Z linked through shaft 924 to turbine 916 and vapor-combustion gas mixing chamber 914 defining the space between the turbine and the compressor. Injectors 922 emit hot, clean combustion gases for mixing the vapor with the result that the combined flow of the vapor and the combustion gases operate turbine 916, which, through shaft 924, drives compressor 912. The system also includes a dirty combustion gas operated turbine 640 which consists essentially of a hollow spindle 642 and blades 644 attached to the out~ide surface of the hollow spindle. The spindle 642 is drivingly linked to shaft 646 through supports 648.
Shaft 646 is operatively linkëd with the spindle 91g of turbine 916 which spindle is joined through shaft ,. _ . ~ . . . .
924 to the spindle of compressor 912. In operatiOn, the system is energized by starting motor 650 acting through shaft extension 652 and clutch 654. Dirty combustion ~l~Z442 gas turbine 6sO is disposed with its blades arranged in flow space 656 which is annularly arranged with respect to vapor and clean combustion gas flow space 914 and which is separated therefrom by a solid paritition', and sealing ring 905. In this manner, hot, dirty comb~stion gases are directed through space 656 to act on turbine blades 644 which, through spindle 642 and supports 648, rotate shaft 646. The expanded dirty combustion gases exhaust from the turbiné 640 into space 658 in such a manner that they never combine or mix with the vapor or the clean combustion gases.
Figure 11 illustrates a completely concentric unit wherein one compressor-mixing'chamber-turbine combina-tion surrounds and is directly linked to another compressor-mixing chamber-turbine combination, In this configuration, the outer compressor-mixing chamber-turbine combination supplies rotary power to the inner system to improve the performance of the inner system. The inner system, which is the compressor-mixing-chamber. turbine combination disclosed in Figures 1, lA,and lB, includes compressor 912 linked through shaft 924 to turbine motor 916 and mixing chamber 914 between the compressor and the turbine in which clean combustion gases emitting from injectors 922 admix with the vapor flowing through chamber 914 to operate turbine 916. Extending from the spindle of compressor 912 and from spindle 919 of turbine 916 are .... .. , . . - .; -- , , shaft members 700 and 702 respectively. .Connected to ,,. s,h,,aft 702 are supports 704 which rotate compressor 706 through its hollow spindle 708. Connected to shaft 700 are suppo.rts 710 through which shaft 70Q is rotated by the hollow spindle 712 of turbine 714. The blades 707 - - , _ 90 --llSZ442 of compressor 706 and 713 of turbine 714 are arranged in an annular space surrounding the compressor-turbine unit 912,916. The annular space is separated from the vapor clean combustion gas flow space by a solid partition, and sealing ring 905. Turbine 714 is operated by combustion gases, which m~y be dirty gases, emitted into space 716 through ln~ectors 718. In space 716 the combustion gases may be mixed with air drawn therein from space 720 upstream of compressor 706 which air is drawn into the system and compressed by compressor 706. The air admixed with the hot combustion gases exhausts through space 722 and never comes in contact with the vapor and clean combustion gases which move through space 914. As the dirty combustion gases and air drawn in through space 716 pass through turbine 714, they do work on the turbine blades 713 causing turbine 714 to rotate and to transmit power through supports 710 to shaft 700, which power is utilized by coaxial compressor 912 in doing work on the vapors which are drawn into space 914. In an alternative form of this embodiment, space 716 may operate as a combustion chamber and injectors 718 used to inject fuel into the space for combustion with the air drawn in from space 702.
Numerous modifications can be made to the con-figuration illustrated in Figure 11 to alter it and/or improve it for particular usages. Thus, supports 704 and 710 could be formed into air foil shaped fans to ... , .. . .,, . ;. , assist in the movement of large masses of vapor. Still . another modification involves clutching and gearing the outer compressor-turbine combination to the inner com-pressor-turbine combination in order that the rate of rotation of the latter could be varied with respect to (--1~52442 the former. Another useful modification is the addition of further compressor-turbine combinations in concentric relationship to the two shown in Figure 11, all with the purpose of increasing the motive power available for compression in compressor 912 and of utilizing avail-able energy sources, such as dirty combustion gases, in as economical a manner as is possible. The fundamental advantage of the configuration of Figure 11 is that it enables utilization of as many different combustion gas sources as may be available at the system location for supplying economical power to compress the vapors flowing into space 914.
Figures 12 and 13 show still other configurations for the compressor-mixing chamber-turbine unit of Figures 1, lA and lB. Specifically, these Figures 12 and 13 illustrate the use of centrifugal compressors instead of or in addition to turbine compressors. Centrifugal compressors have the advantage that the~ readily pass condensed li~uid via the large waterways at the tips of the compressors impellers. Referring first to Figure 12, there is shown an inlet nozzle which leads from the evaporative unit directly to the impeller of a centrifugal compressor. Nozzle 750, which is optionally a venturi nozzle but may be merely an inlet duct, directs the hot vapor to impeller 752 of a centrifugal compressor which includes back plates 754 to prevent the flow of v~por straight-through and to assist impeller 752 in directing and concentrating the flow of vapor toward the sides 756 of the chambeL off the tips of the impeller. The compressed vapor passing centrifugal impeller 752 flows past back plates 754 and into space 758 where it mixes . . ' - .
~ 92 -~15244Z
with hot, clean combustion gases issuing from injectors 760 which are shown in Figure 12 to be optional multi-nozzle injectors. The flow of combustion gases through injectors 760 is controlled by flow valves 762 disposed in the arms 764 leading to the injectors. The vapor passing the ce~trifugal compressor admixes with the combus-tion gases and together the vapor and gases motivate turbines 766 and 768 disposed in tandem. As spindles 765 and 767 of turbines 766 and 768 are caused to rotate, they in turn rotate shafts 770 and 772 linked through clutch and transmission box 774 to shaft 776. Rotation of shaft 776 operates impeller 752 of the centrifugal compressor. As in the other configurations disclosed herein, the system can be started rotating initially utilizing a starter motor through a clutched system shaft-linked to one of the spindles 765, 767 of the tandem turbines. Optional butterfly valve 778 is shown disposed in the neck of entrance nozzle 750 to control the flow direction of the vapors entering from the boiler. The butterfly valve 778 is perferably arranged in s~ch a manner that arms 778a and 778b can be brought together to fully open nozzle 750 and, in that position, to offer little or no resistance to vapor flow therethrough.
Figure 13 illustrates turbine compressor 912 shaft linked through shaft 924 to turbine motor 916 and clean combus-tion gas injectors 922 disposed in mixing chamber 914 to emit clean combustion gases for combination with the ; ~ - vapor flowing through compressor 912 to conjointly operate turbine 916. Starting motor 786 and clutch 788 are provided for initial start-up of the system. In this embodiment, however, a centrifugal impeller 78~ is operated by shaft ': ` , . :, llS2442 924 in conjunction with back plates 782. As described in cc>nnection with Figure 12, the impeller together with the back plates directs and concentrates the flow of vapor toward the ends of the impeller into spaces designa-ted generally as 784 whereupon the vapors are additionally compressed prior to admixing in space 914 with the clean com~ustion gases,emitting from injectors 922.
Yet another useful configuration for the com-pressor-mixing chamber-turbine unit is illustrated generally at 800 in Figure 14. The unit shown consists of two compressor-turbine combinations in tandem together with a free-wheeling compressor upstream of the tandem com-binations. Specifically, free-wheeling compressor 802 is disposed in the path of vapor entering the unit and permitted to rotate at its own rate which is dependent only on the flow rate of vapor therethrough. Starter motor 828 and clutch 830 are shown operating on shaft 804 to which spindle 801 of the free-wheeling compressor is also connected. Hot clean combustion gases enter the system through feed linés 806 and are emitted into mixing chamber 808 of each tandem unit through injectors 810 therein. The hot, clean combustion gases admix with the vapor flowing through chambers 808 and the vapor and gases together operate on turbines 812 and 814.
Turbines 812, 814 are linked respectively, through shafts 816, 818 to compressors ~20, 822, which compressors are .:,........................... .. . ,.: ,, - "
operated by rotation of turbines 812 ~nd 814. As com-, pressor 820 and 822 are rotated, vapor is drawn into the unit past free-wheeling compressor 802 causing the compressor to rotate while supported by supports 824 and bearings 826. The configuration of Figure 14 has ~SZ442 the obvious advantage of affording a larger through-put while utilizing less power due to the presence of the free-wheeling compressor 802. Depending upon the motive , power necessary for compression in the system, either or,both of turbines 812 and 814 can be used.
The present invention in all its embodiments, has thus far,~been described in terms of its operation under the preferred conditions wherein the temperature in the boiler, Tl, is below the boiling point of the liquid under ambient conditions ana the pressure in the boiler, Pl, is below ambient pressure. It is anticipated that the vast m'ajority of users will wish to operate under these conditions and, in most circumstances, it is most economical to operate under these conditions.
However, there are circumstances where it will be desirable to operate at or above the boiling point of the liquid and at or above ambient pressure. For example, if the raw feed liquid is available from its source at or above its boiling point it may be more economical to operate , the system above ambient pressure. In some cases high evaporation temperatures will be beneficial where use of flash distillation apparatus is contemplated. It may also be desirable to employ high temperatures where the influent feed is sea water and a brine pre-heater together with chemical additions to the feed is employed to raise the feed temperature and prevent scaling. --~............................. ... . . ....... . .
However, absent some special circumstance, th,e p,resent ,~..,. ~~--- ,,invention is preferably operated between the freezing and boiling point of the raw liquid feed at ambient condi-tions and at a pressure below ambient pressure.
In those situations where the present invention ~ 152442 is to be practiced at or above the boiling temperature of the liquid determined at ambient pressure and at or above ambient pressure, the temperature in the boiler, Tl, should be less than about the critical temperature, i.e., the temperature above which the vapor cannot be condensed regardless of the pressure applied thereto, which or water 1s about 705.47F. For obvious reasons, as a practical matter, it is unlikely that one would choose to operate at such a high temperature in view of the very substantial equip~ent and energy costs which would be incurred. However, the system will operate as described herein at any temperature from boiling up to the critical temperature, determined under ambient conditions, provided only that the system parameters are controlled to assure a temperature differential in the condenser between the vapor in the condenser return line and the raw feed liquid. There should be no dif-ficulty in adjusting the system parameters to assure this temperature differential, although it should be understood that the system may have to operate at some-thing less than optimum cost conditions. The boiler temperature will, in most cases be less than about 350F
and the corresponding pressure, Pl, will therefore be a pressure not exceeding a pressure correspo~ding to the.evaporation temperature under saturated conditions.
Using the same calculational techniques employed .,. . ., . . ,;, . , in the Examples herein, and selecting P5 above 1 atm.
to insure efficient condensation, it can be seen that by appropriate selection of the system parameters, a system can be devised to produce whatever flow-rates may be required by the user, it being understood that the greater the flow rate the greater the cost of purifi-l~S244Zcation per thousand gallons, all else being equal. Thus, Table IV shows some approximate values of flow,- R, and cost per thousand gallons for a by-pass configuration, as shown in Figures 1, lA and lB, wherein Tl is selected to be 212F and 300F at a compres,ion ratio (CR) of lS:1 and where Tl is 250F at a compression ratio of 1.12:1. In ~hese instances P5 is arbitrarily selected to be twicë Pl.

Table IV -(F) (F) ~F) ~F) atm atm ~gal/hr) 1S D ~ LM CR 1 5 R $/1000 gal ~12336 232 57 15 1.00 2.0 19.02 43,802 1.~2 300 440 320 61.7 15 4.56 9.12 19.3 50,622 1.40 250 270 256 11.6 1.12 2.03 2.27 99.4 9,108 0.20 Where the independent compressor configuration of Figure lS i5 used, for a first compressor ratio of 15:1 Table V shows some approximate values of flow, R, and cost per thousand gallons for a 40' condenser wherein Tl is selected to be 300F for a pressure downstream of the independent compressor, P5, selected to be 1 atm.
greater than Pl. Table V also shows the cases, for Tl = 300F and 500F, where the independent compressor ratio is increased so that P5 is four-times Pl, ~

. ~. -- - .

' -~15Z442 ,, ., o~
o I~ U~ O~ ~ o o U~ o ". C~
)J
S O~
_, U~ o U~
.

_1 U~ CD cn ~ o E~ ~ ~r ~
_~ ~ o~
E~

n ~ U~ .
JJ ~ J~
~ a~

n _l ~
~r ~ r~l o o .... E~
~ U~ :~
.., o E~ ~ o ~

~ ~ :~ o o o E~ ~ o o ., r~

.. . .

.
.

while the present invention has been described with reference to particular embodiments thereof, it will be understood that n~merous modifications can be made by those skilled in the art without actually de-parting from the scope of the invention. Accordingly, all modifications and e~uivalents may be resorted to which fall within the scope of the invention as claimed.

.
, ., ~ . . . .

Claims (75)

THE EMBODIMENTS OF THE INVENTION IN WHICH AN EXCLUSIVE
PROPERTY OR PRIVILEGE IS CLAIMED ARE DEFINED AS FOLLOWS:

1. A method for high volume distillation of impure liquids comprising the steps of:
(a) evaporating said impure liquid in an evaporator to form a vapor at a temperature above the freezing point and below the critical temperature of said liquid and at a pressure not exceeding a pressure corresponding to said evaporation temperature under saturated conditions;
(b) compressing said vapor such that the ratio of vapor pressure of said vapor following compression to evaporated liquid vapor pressure is in the range 1.2:1 to 250:1;
(c) supplying sufficient energy for said compression step, a portion of said energy being supplied by passing a hot gas through a means for expanding said gas whereby a portion of the energy is produced for said compression step, said hot gas comprising at least in part a gas other than said vapor produced in subparagraph (a) hereof and a portion of said energy being supplied by passing at least a part of said compressed vapor through a means for expanding said vapor;
(d) cooling said vapor in heat transfer relation with said impure liquid whereby said vapor at least partially condenses, transferring sufficient heat to said impure liquid for evaporating at least a part of said liquid and to form a vapor therefrom having said temperature and pressure characteristics set forth in subparagraph (a) hereof; and (e) collecting said condensed vapor.

2. A method, as claimed in Claim 1, including the step of passing at least a portion of said hot gas f low downstream of said hot gas expanding means in heat exchange relationship with said vapor to transfer heat from said gas to said vapor.

3. A method, as claimed in Claim 1, wherein said expanded vapor is compressed prior to cooling in heat transfer relationship with said impure liquid to form a second vapor at a predetermined pressure corresponding to a predetermined temperature differential between said com-pressed second vapor and said impure liquid.

4. A method, as claimed in Claim 3, wherein said compression of expanded vapor is achieved in a substantially adiabatic fashion.

5. A method, as claimed in Claim 1, wherein said vapor expands in passing through said hot gas expanding means in a substantially adiabatic manner.

6. A method, as claimed in Claim 1, wherein said impure liquid is water and said evaporation temperature is in the range 33-200°F.

7. A method, as claimed in Claim 1, wherein said impure liquid is evaporated to form a vapor at a temperature in the range 33°F to 160°F.

8. A method, as claimed in Claim 1, wherein at least a portion of said energy is added by driving said compression means with an external mechanical energy source.

9. A method, as claimed in Claim 1, wherein said impure liquid is evaporated to form a vapor at a temperature in the range 33°F to 211°F.

10. A method, as claimed in Claim 9, wherein said vapor has a pressure below atmospheric and corresponds to the saturated vapor pressure of the liquid at the vapor temper-ature.

11. A method, as claimed in Claim 1, wherein said condensation of vapor occurs in said evaporator and said released heat is transferred to said impure liquid in said evaporator to evaporate said liquid.

12. A method, as claimed in Claim 1, wherein the ratio is in the range 5:1 to 100:1.

13. A method, as claimed in Claim 12, wherein the ratio is in the range 5:1 to 50:1.

14. A method, as claimed in Claim 1, wherein said vapor is compressed in a substantially adiabatic manner.

15. A method, as claimed in Claim 1, including the step of diverting a fraction of said vapor prior to cooling and injecting said diverted vapor fraction directly into said impure liquid at a point upstream of said evaporator, whereby said vapor condenses and said impure liquid is heated.

16. A method, as claimed in Claim 1, wherein said energy from expanding said hot gas is obtained by directing a flow of hot gas thorugh a space separate from the space in which said vapor flows, said hot gas flow in said hot gas flow space passing through means for expanding said gas.

17. A method, as claimed in Claim 16, including the steps of drawing air through said hot gas flow space for mixing with said hot gas flow therein, passing said air through means for compressing said air prior to mixing with said hot gas flow, and drivingly linking said hot gas expanding means in said hot gas flow space with said air compressing means in said hot gas flow space, whereby at least a part of the energy produced by expanding said hot gas flow is used to operate said air compressing means.

18. A method, as claimed in Claim 16, wherein said hot gas flow space is annularly disposed with respect to the space in which said vapor flows.

19. A method, as claimed in Claim 17, wherein said hot gas flow space is annularly disposed with respect to the space in which said vapor flows.

?

20. A method, as claimed in Claim 17, further including the steps of admitting fuel into said hot gas flow space upstream of said hot gas expanding means and igniting said fuel, whereby said hot gas flow is produced in said space.

21. A method, as claimed in Claim 20, wherein said hot gas flow space is annularly disposed with respect to the space in which said vapor flows.

22. A method, as claimed in Claim 1, including the step of passing at least a portion of said hot gas flow downstream of said hot gas expanding means in heat exchange relationship with said vapor to transfer heat from said gas to said vapor.

23. A method, as claimed in Claim 20, wherein said impure liquid is evaporated to form a vapor at a temperature in the range 33°F to 211°F, and wherein said condensation of vapor occurs in said evaporator and said released heat is transferred to said impure liquid in said evaporator to evaporate said liquid.

24. A method, as claimed in Claim 1, wherein said vapor temperature is below the bailing point of said liquid at ambient pressure.

25. A method for distillation of impure liquids comprising the steps of:
(a) evaporating said impure liquid in an evaporator to form a vapor at an evaporation temperature above the freezing point and below the critical temperature of said liquid and at a pressure not exceeding a pressure corresponding to said evaporation temperature under saturated conditions;
(b) compressing said vapor such that the ratio of vapor pressure of said vapor following compression to evaporated liquid pressure is in the range 1.2:1 to 250:1;

?

(c) supplying sufficient energy for said compression step, a portion of said energy being added by passing a hot gas through a space separate from the space in which said vapor flows, said hot gas flow passing through means for expanding said gas, said hot gas comprising at least in part a gas other than said vapor produced in subparagraph (a) hereof;
(d) drawing air through said hot gas flow space for mixing with said hot gas flow therein, passing said air through means for compressing said air prior to mixing with said hot gas flow, and drivingly linking said hot gas expanding means in said hot gas flow space, with said air compressing means in said hot gas flow space, whereby at least a part of the energy produced by expanding said hot gas flow is used to operate said air compressing means;
(e) cooling said vapor in heat transfer relation with said impure liquid whereby said vapor at least partially condenses, transferring sufficient heat to said impure liquid for evaporating at least a part of said liquid and to form a vapor therefrom having said temperature and pressure characteristics set forth in subparagraph (a) hereof; and (f) collecting said condensed vapor.

26. A method, as claimed in Claim 25, wherein said impure liquid is water and said evaporation temperature is in the range 33-200°F.

27. A method, as claimed in Claim 25, wherein said impure liquid is evaporated to form a vapor at a temperature in the range 33°F to 160°F.

28. A method, as claimed in Claim 25, wherein at least a portion of said energy is added by an external mechanical energy source.

29. A method, as claimed in Claim 25, wherein said impure liquid is evaporated to form a vapor at a temperature in the range 33°F to 211°F.

30. A method, as claimed in Claim 29, wherein said vapor has a pressure below atmospheric and corresponds to the saturated vapor pressure of the liquid at the vapor temperature.

31. A method, as claimed in Claim 25, wherein said condensation of vapor occurs in said evaporator and said released heat is transferred to said impure liquid in said evaporator to evaporate said liquid.

32. A method, as claimed in Claim 25, wherein the ratio is in the range 5:1 to 100:1.

33. A method, as claimed in Claim 32, wherein the ratio is in the range 5:1 to 50:1.

34. A method, as claimed in Claim 25, wherein said vapor is compressed in a substantially adiabatic manner.

35. A method, as claimed in Claim 25, including the step of diverting a fraction of said vapor prior to cooling and injecting said diverted vapor fraction directly into said impure liquid at a point upstream of said evaporator, whereby said vapor condenses and said impure liquid is heated.

36. A method, as claimed in Claim 25, including the step of diverting a fraction of said compressed vapor and directing said diverted fraction through said hot gas flow space.

37. A method, as claimed in Claim 25, wherein said vapor temperature is below the boiling point of said liquid at ambient pressure.

38. A method, as claimed in Claim 25, further including the steps of admitting fuel into said hot gas flow space upstream of said hot gas expanding means and igniting said fuel, whereby said hot gas flow is produced in said space.

39. A method, as claimed in Claim 38, wherein said impure liquid is evaporated to form a vapor at a temperature in the range 33°F to 211°F, and wherein said condensation of vapor occurs in said evaporator and said released heat is transferred to said impure liquid in said evaporator to evaporate said liquid.

40. A method, as claimed in Claims 25 or 36, wherein said hot gas flow is annularly disposed with respect to the space in which said vapor flows.

41. A method, as claimed in Claims 38 or 39, wherein said hot gas flow space is annularly disposed with respect to the space in which said vapor flows.

42. A method, as claimed in Claims 26, 27 or 28, wherein said hot gas flow space is annularly disposed with respect to the space in which said vapor flows.

43. A method, as claimed in Claims 29, 30 or 31, wherein said hot gas flow space is annularly disposed with respect to the space in which said vapor flows.

44. A method, as claimed in Claims 32, 33 or 34, wherein said hot gas flow space is annularly disposed with respect to the space in which said vapor flows.

45. A method, as claimed in Claims 35 or 37, wherein said hot gas flow space is annularly disposed with respect to the space in which said vapor flows.

46. A method, as claimed in Claim 38, wherein said hot gas flow space is annularly disposed with respect to the space in which said vapor flows.

47. A method, as claimed in Claim 36, wherein said hot gas flow space is annularly disposed with respect to the space in which said vapor flows.

48. A system for high volume distillation of impure liquids comprising:
(a) evaporator means, including means for supplying impure liquid feed thereto, for evaporating said impure liquid at ambient pressure and at a pressure not exceeding a pressure corresponding to said evaporation temperature under saturated conditions;
(b) first compressor means receiving said vapor from said evaporator means for increasing said vapor pressure and temperature;
(c) condenser means in heat transfer relationship with said impure liquid feed for receiving said vapor and for at least partially condensing said vapor whereby the heat released by said vapor is transferred to said feed liquid to supply at least a part of the heat energy necessary for evaporating said feed liquid;
(d) duct means communicating with said condenser means for carrying said vapor from said first compressor means to said condenser means;
(e) means for recovering condensate from said condenser means;
(f) means for removing unevaporated liquid feed from said evaporator means; and (g) auxiliary turbine means drivingly connected to said compressor means, said auxiliary turbine means including an auxiliary flow conduit for hot gas flow therethrough, said conduit being annularly disposed with respect to and separated from the space in which said vapor flows, and turbine blading in said conduit, said turbine blading drivingly linked to said compressor means whereby hot gas flow through said conduit does work on said turbine blading which work is transmitted to said compressor means.

49. A system, as claimed in Claim 48, wherein said first compressor means is drivingly connected to said auxiliary turbine means through a shaft, and said auxiliary turbine means includes a spindle supporting said blading and drivingly connected to said shaft.

50. A system, as claimed in Claim 49, wherein said spindle is hollow and said vapor flows therethrough.

51. A system, as claimed in Claim 48, wherein said gas flowing in said auxiliary flow conduit is dirty gas.

52. A system, as claimed in Claim 48, further including auxiliary compressor means in said auxiliary flow conduit upstream of and drivingly connected to said auxiliary turbine means, said auxiliary compressor means drawing air through said conduit whereby said air flow together with said gas flow in said conduit motivates said auxiliary turbine means.

53. A system, as claimed in Claim 48, further including auxiliary compressor means disposed in said auxiliary flow conduit, said auxiliary compressor means upstream of and drivingly connected to said auxiliary turbine means through said shaft and including a spindle supporting compressor blading in said conduit.

54. A system, as claimed in Claim 53, wherein said spindle is hollow ant said vapor flows therethrough.

55. A system, as claimed in Claim 53, further including means for admitting fuel to said auxiliary flow conduit upstream of said auxiliary turbine blading and means for igniting said fuel, whereby said hot gas flow is produced in salt conduit.

56. A system, as claimed in Claim 48, wherein said first compressor means comprises compressor means having a variable compression ratio.

57. A system, as claimed in Claim 56, wherein said compression ratio is in the range 1.2:1 to 250:1.

58. A system, as claimed in Claim 57, wherein said compression ratio is in the range 5:1 to 100:1.

59. A system, as claimed in Claim 58, wherein said compression ratio is in the range 5:1 to 50:1.

60. A system, as claimed in Claim 48, further including second compressor means receiving said compressed vapor upstream of said condenser means for forming a second vapor having increased vapor pressure and temperature and means for driving said second compressor means.

61. A system, as claimed in Claim 60, wherein said means for driving said second compressor means is a mechanical energy source drivingly linked to said second compressor means.

62. A system, as claimed in Claim 60, further including duct means upstream of said condenser means, downstream of said second compressor means and communicating with said evaporator means for diverting a fraction of said compressed second vapor directly to said evaporator means for admixture with said impure liquid feed therein.

63. A system, as claimed in Claim 48, further including duct means upstream of said condenser means and communicating with said evaporator means for diverting a fraction of said compressed vapor directly to said evaporator means for admixture with said impure liquid feed therein.

64. A system, as claimed in Claim 48, further including duct means downstream of said first compressor means for diverting a fraction of said compressed vapor directly to said means for supplying impure liquid feed to said evaporator means.

65. A system, as claimed in Claim 64, wherein said means for supplying impure liquid feed includes a feed duct and said duct means downstream of said first compressor means includes a vapor injector means communicating with said feed duct to inject compressed vapor therein.

66. A system, as claimed in Claim 48, wherein said condenser means is disposed within said evaporator means in heat transfer relationship with said liquid feed in said evaporator means.

?

67. A system, as claimed in Claim 48, further including means for regulating the vapor pressure in said evaporator means.

68. A system, as claimed in Claim 48, further including heat exchange means in heat transfer relationship with said vapor in said duct means for receiving at least a portion of said hot gas flowing in said auxiliary flow conduit downstream of said auxiliary turbine blading, said hot gas passing in heat transfer relationship with said vapor to transfer heat from said gas to said vapor.

69. A system, as claimed in Claim 48, further including bypass conduit means downstream of said first compressor means and communicating with said auxiliary flow conduit upstream of said auxiliary turbine blading for diverting a fraction of said compressed vapor directly to said conduit for motivating said auxiliary turbine.

70. A system, as claimed in Claim 69, wherein said bypass conduit means includes a vapor injector means communicating with said auxiliary flow conduit for injecting compressed vapor therein.

71. A system, as claimed in Claim 55, wherein said first compressor means comprises compressor means having compression ratio variable in the range 1.2:1 to 250:1, wherein said condenser means is disposed within said evaporator means in heat transfer relationship with said liquid feed therein; and further including means for regulating the vapor pressure in said evaporator means.

72. A system, as claimed in Claim 48, further including expansion engine means motivated by vapor from said first compressor means, said engine being drivingly connected to said first compressor means whereby the work done by said vapor in expanding in said expansion engine means is transmitted to said first compressor means, said vapor exiting said engine means being carried by said duct means

73. A system, as claimed in Claim 72, wherein said auxiliary flow conduit is annularly disposed with respect to and separated from the space in which said vapor flows, said first compressor means is drivingly connected to said expansion engine means and said auxiliary turbine means through a shaft, said auxiliary turbine means includes a spindle supporting said blading and drivingly connected to said shaft.

74. A system, as claimed in Claim 72, wherein said expansion engine means is coaxial with said first compressor means.

75. A system, as claimed in Claim 72, wherein said expansion engine means is coaxial with said first compressor means, said first compressor means comprises compressor means having a compression ratio variable in the range 1.2:1 to 250:1; said condenser means is disposed within said evaporator means in heat transfer relationship with said liquid feed therein; and further including means for regulating the vapor pressure in said evaporator means.