Note: Descriptions are shown in the official language in which they were submitted.
~;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
115Z442
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
115244Z
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-
-- 7 --
~ 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
- 8 -
~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
244Z
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
-- 10 --
~ 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
- - 12 -
~ 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
- 13 -
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
- 14 -
~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
-- 15 --
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,
- 17 -
- ~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
- 18 -
~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-
-- 19 --
;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
- 21 -
~ 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
- 23 -
~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,
- 26 -
52442
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
- 27 -
~ 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
- 28 -
~ 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
- 29 -
~5244Z
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-- 31 --
~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.
- 32 -
~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
- 34 -
~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.
- 35 -
1152442
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
- 36 -
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
- 38 -
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
- 39 -
~ 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.
- 40 -
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
- 41 -
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
- 42 -
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
.. . ~ ... . . .
- 43 -
~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 - $2.95
In still another broad form of the in~ention
illustrated in Figures 16-18, many of the advantages
of the already described embodiments are combined with
the relative simplicity of conventional vapor compression
systems to overcome the apparent ~hortcomings of such
conventional systems. -.Vapor- compression systems are .
w~ll known for the treatment of impure liquids~ However,
.. . . .
the system configurations heretofore known have suffered
from serious disadvantages which have limited their usefulness.
For example, vapor compression systems are typically
- 45 -
~lS2442
designed to accept and treat a particular liquid, e g ,
salt water, entering the system within a narrow range
of initial conditions. As a result, the system is incapable
of being used for other liquids or for other initial
conditions, and, therefore, its usefulness is limited.
In addition, conventional vapor compression systems must
operate a~ low compression ratios, e.g. 1.2:1 to 1.5:1,
to minimize cost. This means that the temperature in
the evaporator must be close to 212F
because such low compression ratios prevent drawing any
substantial vacuum in the evaporator. Moreover, since
the specific volume of water vapcr decreases rapidly
as temperatures drop below 212F and in view of the low
compression ratios which must be used, if the vapor
pressure exiting the compressor is to be high so that
the vapor temperature may be high, the evaporator must
j operate at or near 212F. ThiS effective temperature
lim$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
- - 46 -
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,
- 47 -
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.
- 48 -
~ 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
- 50 -
~ 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
-- 51 --
~ 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
- 52
~ 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;
- 53 -
~ 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
- 54 -
..
, .,
~ 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
- 55 -
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
- 56 -
1152442
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
- 57 -
.
115Z442
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
2 1 2
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 =
- 59 -
.
115244Z
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
- 60 -
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
61 -
, ,
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
.
- 62 -
~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.
.. . , . , .;, . . .
. . ~
.... ... .
- 63 -
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~ ~ 52442
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~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
- 66 -
~ , .
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
- 69 -
: 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.
- 70 -
~ 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
- 71 -
~ 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.
- 72 -
~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
0
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
_~
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~ O o - ~ ~ o o o o o o U~
O
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,~ _I ~ _I_I ~ U~ o o
-- 73 --
~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-
- 74 -
~ 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
- 75 -
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
- 78 -
(` (
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
- 81 -
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
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1152442
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
- - 83 -
.- ..
~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.
- 84 -
- ;
~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
- - 85 -
-- . .
~ 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
- 86 -
~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
- 87 -
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
- - 88 -
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
89 -
~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
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244Z
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.
.
, ., ~ . . . .
