Note: Descriptions are shown in the official language in which they were submitted.
- 1100317
The present invention relates to mechanical energy
generating systems and, more particularly, to a method and
apparatus for converting thermal energy to mechanical energy.
As is universally appreciated, the world's supply
of conventional fuels, such as natural gas, oil, coal, and
the like, is being rapidly consumed and the continued availab-
ility of such fuels has, in recent years, been seriously que-
stioned. Although these fuels are utilized for many purposes
in our soclety, perhaps no uses are more important than as a
source of thermal energy convertible to mechanical energy for
furnishing motive power for vehicles and boats or to electric-
al energy for powering our households and industries. Of
course, numerous alternative energy sources are available and
under development, e.g., solar energy, nuclear energy, and the
like, and to the extent that these alternative sources are
utilized they will partially alleviate the present and prosp-
ective fuel crises. ~owever, nuclear energy, for example, is
expensive and its use creates monumental environmental problems
which remain unresolved. Solar energy as a source of power
is still in the developmental stage and, at least at present,
is not totally practical for use in all climates, particularly
in areas where cloud, fog or smog cover is frequent. Indeed,
pd ~,
D3~,7
there is not yet available a viable alternative to the ever
increasing consumption of conventional fuels to supply the
power necessary to operate today's society.
There have been however, numerous suggestions for
systems which more efficientlyutilize energy available from
conventional sources by minimizing thermodynamic and fluid
.~ dynamic losses. Unfortunately, even these systems fail to
significantly improve energy conversion efficiency and, gen-
erally, omit even to make maxinum use of available energy
~- 10 conver~ion opportunities. For example, U.S. patent 3,358,451
discloses a system for con~erting the energy of a liquid
stream to power by heating a liquid working fluid to form a
two-phase fluid, accelerating the fluid, separating the liquid
and vapour phases, converting the kinetic energy of the liquid
phase to work, condensing the vapor phase to liquid and reun-
iting the liquid streams prior to heating again. This system,
however, neglects to extract the maximum work potential from
the available energy of the working fluid and, therefore,
relinquishes, rather than uses, valuable thermal energy to a
heat transfer medium in a condenser.
It is therefore an object of the present invention
to provide a system for producing mechanical energy which
e~tracts more of the available work from a working fluid than
has heretofore been possible.
It is another object of the present invention to
provide a method for producing mechanical energy which need
not consume or utilize the world's conventional fuel supply.
,' pd/ -3-
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It is yet another object of the invention to
provide an environmentally safe method for producing
mechanical energy.
It is still another ob~ect of the invention to
provide a method for producing mechanical energy from thermal
energy where the thermal energy is derived, at least in part,
from available ambient energy sources, such as the atmosphere,
rivers, oceans, waste heat sources, etc.
It is another object of the invention to provide
a method and system which is particularly useful in refrig-
eration and air conditioning applications and which substan-
tially reduces the energy requirements for such applications.
Thus, the present invention is broadly defined as
a continuous method for converting thermal energy to mechan-
ical energy comprising the steps of: providing a working fluid
stream in a substantially saturated liquid state containing
a predetermined quantity of thermal and static energy therein;
converting at least a portion of the energy of the fluid
stream to kinetic energy by accelerating the stream whereupon
a high velocity, reduced temperature and static pressure stream
is obtained, said fluid vaporizing in part under the reduced
pressure to form a vapor fraction in the liquid stream; trans-
ferring a portion of the vapor fraction momentum and thermal
energy to the liquid fraction whereby the vapor fraction is
compressed and a portion of the vapor fraction condenses
transferring a portion of its heat of condensation to the
liquid fraction, the liquid fraction increasing in velocity,
static pressure and temperature; extracting energy from the
two phase stream and converting the energy to shaft work;
adding sufficient static energy to the stream to condense a
ma~or proportion by volume of the vapor fraction thereof and
- 4 -
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to permit the addition to the stream, without vaporization,
of thermal energy sufficient for at least one additional
cycle of the stream; passing the stream in heat absorbing
relation with a source of thermal energy whereby the stream
absorbs thermal energy from the source; and repeating steps
b through f.
A more specific aspect of the present invention
relates to the improvement in a method for achieving heating
or refrigeration including the steps of compressing a refrigerant
vapor stream, condensing the vapor, expanding theresulting fluid
stream to reduce its saturation temperature and pressure, and
passing the expanded fluid stream in heat exchange relationship
with a source of thermal energy to re-vaporize at least a portion
: thereof, the improvement comprising: expanding the fluid stream
by passing it through at least one area of constricted flow
while maintaining the flow in a substantially non-throttling
condition to increase the kinetic energy thereof, and converting
the kinetic energy of the resulting fluid stream to shaft work.
A still further aspect of the present invention relates
to the improvement in a closed cycle mechanical vapor heating
or refrigeration system includins compressor means for raising
the temperature and pressure of a refrigerant vapor stream,
condenser means for cooling the vapor stream to at least a
substantially saturated liquid condition, means for expanding
the liquid stream for reducing its saturation temperature
and pressure, and evaporator means for passing the expanded
stream in heat exchange relationship with a source of thermal
energy to re-vaporize at least a portion thereof, the improve-
ment comprising: the means for expanding the liquid stream
comprising energy conversion means including at least one
substantially non-throttling nozzle section to increase the
-,il
- 4a -
03~7
kinetic energy of the stream, and expansion engine means
for converting the stream kinetic energy to shaft work.
Other ob~ects and advantages will become apparent
from the following description and appended claims, taken
together with the accompanying drawings in which:
FIGURE 1 illustrates, in schematic form, one embodi-
ment of the method and system of the present invention which
utilizes a single working fluid stream.
FIGURE 2 illustrates an embodiment of an energy
conversion tube useful in the system of the present invention.
FIGURE 3 illustrates a preferred form of energy
conversion tube useful in the system of the present invention.
FIGURE 4 illustrates, in schematic representation,
the ~asic elements of a conventional mechanical vapor refri-
geration system.
FIGURE 5 illustrates, on temperature-entropy coordi-
nates, the thermodynamic performance of the system of FIGURE
4.
,~
- 4b -
~100317
FIGURE 6 illustrates, in schematic representation,
the FIGURE 4 system modified to include a nozzle and turbine
in accordance with one aspect of the present invention.
FIGURE 7 illustrates, on temperature-entropy co-
ordinates, the thermodynamic performance of the system of
FIGURE 6.
Referring to the drawings, and particularly to
Figure 1, there is shown a continuous closed cycle system for
converting the energy potential of an appropriately selected
pressurized working fluid into mechanical shaft energy with
system energy losses, including useful shaft work, made up
by drawing energy, in the form of heat, from an available
thermal energy source, such as radioisotopes, nuclear reactors,
combustion heat (particularly from burning of non-convention-
al fuel sources such as garbage), solar energy and from
ambient thermal sources where available in sufficient quan-
tity (e.g,, the atmosphere, rivers, oceans, waste heat
sources, etc). The system illustrated utilizes only one
working fluid stream. Although there are numerous working
fluids which may be used, as a general matter any liquid
i8 suitable which is useful in an expansion work cycle taking
into account the maximum and minimum temperatures and press-
ures of the selected cycle and the need for vaporization
and condensation therebetween. For a system operating at
or slightly above or below normal ambient temperatures
those working fluids are most advantageous which are low
boiling, and preferably those which boil substantially
below the freezing point of water. Typical of these kinds
of working fluids are carbon dioxide, liquid nitrogen and
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the flourocarbons. Examplary of useful fluorocarbons are
difluoromonochloromethane, pentafluoromonochloroethane, dif-
luorodichloromethane, and the mixtures and azeotropes thereof.
For higher temperature, higher pressure systems the fluids
may include water or other well known coolants, even includ-
ing the liquid metals, e,g., sodium, potassium, mercury, and
the like.
In the practice of the invention, the working fluid
st~eam is directed to a means for converting the non-kinetic
10 energy of a stream, e.g., its static pressure, thermal and/or -
potential energy, to velocity or kinetic energy, such means
hereinafter referred to as an energy conversion tube (ECT),
as will be more fully described. In the ECT the velocity
of the stream is caused to increase while at the same time
causing the static pressure and temperature of the stream
to substantially decrease. As the pressure decreases, some
of the thermal energy contained in the liquid is liberated
and a portion of the liquid is vaporized. The resulting
stream contains an increased proportion, by volume, of vapor.
In a preferred form of the invention, the energy
conversion tube includes at least one nozzle section. Desir-
ably, depending upon the system in which it is used, the ECT
comprises a plurality of longitudinally spaced apart nozzle
sections (see Figures 2 and 3) interconnected by a plurality
of recovery sections. Thus, the ECT may comprise a single
nozzle section alone, a nozzle and a recovery section, or
a plurality of nozzle sections separated by a plurality of
recovery sections. In a plural section ECT, the liquid work-
ing fluid, having high potential or static energy lhigh static
pressure), and being substantially saturated, as defined here-
inafter, enters the first nozzle section and is accelerated
p~ 6-
11(~0317
therein to convert it to a high velocity, relatively lower
static pressure stream or jet of fluid flowing axially through
the tube, The velocity of the fluid increases, as does the
kinetic energy, due to the decreasing cross-sectional flow
area as the fluid moves through the nozzle section. As the
fluid accelerates and the static pressure thereon decreases,
the saturated liquid begins to vaporize and, in so doing,
consumes some of its thermal energy. The result is an in-
creased volume, increased kinetic energy, decreased static
pressure, decreased temperature and increased vapor content
stream exiting the nozzle section. By "su~stantially satur-
ated" as used herein, it is meant that the liquid is either
saturated or so nearly saturated that under the flow condi-
tions experienced in the first nozzle section, the liquid
will vaporize at least in part. Most preferred is the cond-
ition wherein the liquid is in fact at saturation at the
entrance to the first nozzle section of the ECT. It is
particularly preferred that the liquid be at saturation at
each nozzle section of the ECT.
The high velocity fluid stream, consisting of a
relatively high velocity liquid fraction and a substantially
higher velocity vapor fraction, exits the nozzle section and
enters a pumping and recovery section wherein the momentum
of the vapor fraction is converted to additional velocity
and increased temperature and static pressure of the liguid
fraction. This is accomplished by transferring a portion
of the kinetic and thermal energy of the vapor fraction to
the liquid fraction whereby the liquid fraction is energized
or regenerated for another expansion cycle through the next
nozzle section. It is believed that in the pumping and re-
covery section, consistant with the conservation of momentum,
the relatively fast moving vapor impacts with the relatively
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slow moving liquid resulting in a momentum exchange between
liquid and vapor and a reduction in vapor fraction velocity.
The velocity reduction is accompanied by a static pressure
increase (compression process) without a net expense of work
causing at least a portion of the vapor to condense and to
transfer its latent heat of condensation to the liquid frac-
tion, The net effect on the motive stream working fluid is
to further increase the velocity and kinetic energy of the
liquid, to condense part of the vapor, to recover a portion
of the static pressure which was converted to kinetic energy
in the nozzle section, and to increase working fluid tempera-
ture (to a value higher than at the nozzle section e~it but
lower than at the nozzle section inlet) so that the increased
static pressure, increased temperature liquid is ready for
_ further acceleration and expansion in the next nozzle section.
The progressive temperature increase of the liquid through
the pumping and recovery section is believed due to the vapor
continuously applying its stagnation pressure to the liquid
during the pumping process. This repeated vaporization-cond-
ensation sequence, occuring within the energy conversion tube,directs the work of vaporization downstream toward the low
pressure area rather than more or less uniformly dissipating
it in all directions. The effect is to create a pumping
action upon the liquid stream with the result that a high
velocity, and thus a high kinetic energy, is imI~arted to the
liquid stream exiting the tube. ~t is believed that the
resulting velocity and kinetic energy is greater where a
plurality of nozzle sections, rather than a single nozzle
section, is employed, Moreover, the use of a plurality of
spaced nozzle sections permits a portion of both the latent
and kinetic energy content of the vapor following initial
nozzl~g to be transferred back to the liquid where its thermal
portion is susceptible of reuse and conversion to additional
.~ ~
pd~ ~ -8-
31~
kinetic energy. By contrast, for example where only a single
nozzle section is used, the possibility of reusing the latent
thermal energy of the vapor is reduced, necessitating the
rejection of an increased amount of thermal energy to the
heat transfer medium, e.g., in a condenser, rather than fur-
ther using it to do additional work. This recovery and reuse
of the latent heat is believed to be one important aspect
of the improved performance obtainable with the system of
the present invention.
A plural nozzle energy conversion tube configured
for subsonic flow conditions is shown in Yigure 2. lt will
be appreciated by those s~illed in the art that a correspond-
ing tube having nozzles suitably configured for supersonic
flow conditions can readily be provided by those skilled in
the art. Tube 100 includes a plurality of spaced apart nozzle
sections 102, 104, 106, 108 interconnected by a plurality
of generally cylindrical recovery sections 110, 112, 114.
In the embodiment illustrated in Figure 3, energy conversion
tube 200 comprises a plurality of longitudinally diverging
diffuser sections 210, 212, 214. It is generally preferable
to utilize diffuser sections as the recovery sections since
they act as a controller, helping to pre~ent the static pres-
sure front within the tube from being so low at any point
that radial, rather then downstream directed, expansion of
a liquid droplet occurs,
As another important aspect of the present inven-
tion, the energy conversion tube also desirably includes means
for disrupting metastable flow conditions therein. Any known
methods for disrupting metastable flow may suitably be used.
3~ For example, a secondary flow stream, such as a mercury stream,
may be added to the working fluid stream. Alternatively,
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mechanical means may be used, such as interposing diverters
or turning vanes in the working stream. Still other methods
for disrupting metastable flow involve use of non-mechanical
means, e.g., using sound or radio waves. As has already been
indicated, it is important that the vaporization of the work-
ing fluid stream in a single nozzle or the vaporization-cond-
ensation sequence of the stream in a plural nozzle-recovery
section ECT take place under controlled conditions such that
the vaporizing action of the working fluid on experiencing
reduced pressure occurs within the ECT where the thermal
energy given off can be utilized by the stream and converted
to kinetic energy. If the vaporization occurs outside the
ECT, as might be the case if actual vaporization trails the
attainment of a sufficiently reduced pressure for vaporiza-
tion, then a metastable flow condition exists and the thermal
energy released will not be efficiently converted to kinetic
! energy of the stream. Thus, continued disruption of meta-
stable flow conditions, even in a single nozzle section ECT,
is a very desirab~e aspect of the present invention. By
disrupting metastable flow while at the same time utilizing
spaced apart nozzle sections wherein the throat pressure of
each nozzle section is less than the corresponding throat
pressure in nozzle sections upstream thereof, the desired
pumping action in the tube and increased liquid flow velocity
can be most efficiently achieved. For subsonic flow condi-
tions in the E~T, this means that the flow area of each nozzle
section is smaller than the flow area of nozzle sections
upstream thereof. For supersonic flow conditions, the flow
area of each nozzle section is larger than the flow area of
nozzle sections upstream thereof.
Referring to Figure 1, it will be appreciated that
the fluid employed can be any of the working fluids described
1~-! pd/J-o --10--
llOQ317
hereinbefore. Since the system of Figure 1 is continuous
and closed, for descriptive purposes the inlet to ECT 50 has
arbitrarily been selected as the system starting point. The
fluid enters ECT 50 in liquid form, preferably saturated,
at temperatures and pressures corresponding to an enthalpy
content at least as high as the anticipated energy losses,
including useful shaft work, in the system. In passing
through ECT 50, which is a nozzling device such as the energy
conversion tubes hereinbefore described, the potential, therm-
al and static energy of the fluid stream is partially convert-
ed to kinetic energy and the velocity of the stream is consid-
erably increased while the static pressure and temperature
of the stream is considerably decreased. Since kinetic energy
is a function of both mass and velocity, where it is desirable
to decrease flow velocity without decreasing stream kinetic
energy the mass of the stream can be increased by adding a
secondary flow stream, e.g., mercury. The addition of the
secondary stream decreases the overall flow velocity while
the mass increase keeps the kinetic energy substantially
constant. The secondary stream also functions as an aid in
disrupting metastable flow in the ECT, Vaporization of a
portion of the liquid stream occurs within ECT 50 changing
the physical ~ature of the stream from substantially all
liquid to largely vapor, by volume.
The relatively high velocity stream exiting ECT
50 impinges directly upon the blade portion of turbine 52
to convert a portion of the kinetic stream energy to mechan-
ical shaft energy of the turbine. If desired, an optional
throttle valve (not shown) can be inserted downstream of ECT
50 and upstream of turbine 52 to control the amount of
kinetic energy converted in the turbine.
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The stream exiting the turbine is spent and, if
the system is to be continuous and closed and the spent stream
is to be ~ecycled, the energy content of the stream must be
raised. This is preferably achieved in one embodiment of the
invention by separating the liquid and vapor fractions of the
turbine exhaust stream, increasing the static energy content
of each fraction, and then recombining the fractions. One
means of accomplishing the separation is by gravity separation,
for example by vertically stacking the liquid and vapor frac-
tion removal lines 54 and 56, respectively, with liquid line54 below vapor line 56. In addition the diameter of vapor
line 56 is made considerably larger ~for example by a factor
of 10) than the diameter of liquid line 54 to encourage the
vapor-liq~ separation. Vapor compressor or vapor pump 58 in
line 56 and liquid pump 60 in line 54 operated by the shaft
energy produced by turbine 52 increase the static pressure and
energy content of the vapor and liquid fraction streams. At
the same time the vapor compression in pump 58 increases the
vapor temperature to a value considerably in excess of the
liquid fraction temperature. The liquid and vapor streams are
reunited downstream of pumps 58 and 6~. The cold liquid
ser~es as a heat sink for the vapor, for example by passing
the vapor onto a thin film of the liquid in conventional
fashion, causing substantially immediate ~apor condensation
so that the combined streams at the inlet to pump 62, which is
also operated by the shaft ener~y produced by expansion 52, is
substantially all liquid. Pump 62 is a liquid pump which in-
crease the static pressure of the liquid stream to a pressure
sufficient to handle, without vaporization, the thermal energy
increase which the stream experiences in absorber section 64.
At the same time, the pump increases the energy content of the
stream.
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Pump 58, 64 and 62 also perform the necessary func-
tion of rapidly removing the turbine exhaust stream to prevent
back pressure build-up at the turbine which could adversely
affect its efficient operation. It will, of course, be appre-
ciated that it is not necessary to employ a three pump arrang-
ement as shown in Figure 2. Instead, for example, a single
centrifugal pump can be used in lieu of pumps 58-60, or in-
deed, any pump configuration is suitable which will perform
the two basic functions of pumps 58, 60 and 62, i.e., to
remove the turbine exhause stream and increase the static
head and energy level thereof.
The liquid stream leaving pump 62 is directed
through an absorber section 64 wherein thermal energy is
added to the system to make up for energy converted to mech-
anical shaft energy in the turbine 52 and for energy losses
due to friction and other thermodynamic inefficiencies
elsewhere in the system. The absorber section 64 should be
of sufficient length or area to permit absorption of some
predetermined quantity of energy and, to this end, the absor-
ber section 64 includes control means (not shown) whereby thequantity of energy absorbed can be closely controlled. In
passing through absorber section 64, the stream is re-energ-
ized to the desired extent by thermal absorption. The temp-
erature of the stream leaving pump 62 should be considerably
below the temperature of the thermal source in order that the
stream may be re-energized to the desired extent by drawing
upon the thermal energy available from the source, The
amount of thermal energy added to the stream must be
sufficient to make up for the energy converted to mechanical
shaft energy in turbine 52 which is not added back into thc
stream in pumps 58, 60 and 62 and for energy losses due to
~riction and other thermodynamic inefficiencies elsewhere in
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the system. It will be appreciated however, that whatever the
thermal source employed, thermal energy must flow from it to
the system. For this reason, until the thermal source is
selected and its the~mal conditions are defined, an appropr-
iate working fluid and the pressure and temperature parameters
of any particular system cannot be finally selected. The
stream leaving absorber section 64 is substantially saturated
and has a sufficient energy content to start the cycle over
again at the inlet nozzle 50. Alternatively, if desired to
control the net power output, some thermal energy could be
added to the fluid prior to adding the static energy thereto.
The following example is intended to illustrate
one set of ope~ating parameters for a system in which the
thermal energy source is the ambient (postulated to be in
the range 85-100F) and in which freon 22, commercially avail-
able from E . I . duPont de Nemours & Company, Inc., is utilized
as the working fluid.
Example 11
Freon, difluoromonochloromethane, was employed as
the working fluid in the system illustrated in Figure 1.
The fluid energy content parameters in BTU per pound of work-
ing fluid at the indicated locations in the system as weil
as energy additions ~+) to and energy losses (-) from the
system are set forth below;
AT THE ECT INLET
AT THE ABSORBER SECTION OUTLET
Assuming the working fluid to be a liquid having
no substantially velocity, all of the fluid energy is potent-
ial, i.e. static and thermal (hereinafter "PE") and none is
kinetic
~hereinafter "KE"):
. .
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PE = 33.7 BTU/lb,
KE = 0
AT THE ECT OUTLET
AT THE TURBINE INLET
_
PE - 23.7 BTU/lb
KE = 10.0 BTU/lb
AT THE TURBINE OUTLET
AT THE INLETS TO PUMPS 58 and 60
PE = 27 .1 BTC/lb
Work done by fluid in turbine,
assuming a turbine inefficiency
loss of 33~ - 6.6 BTU/lb
IN THE ABSORBER SECTION
~ .
Thermal energy transferred from source = +4.2 BTU/lb.
As hereinabove indicated, once thermal source
temperature conditions are known and the desired turbine mech-
anical energy output is selected, the turbine exhaust stream
temperature can be determined and from,this value an approp-
- ria'te working fluid and minimum pressure and'temperature
operating parameters can be identified. In this connection,
while it is appreciated that there is considerable latitude,
from the standpoint of operability, in selecting temperature
and pressure operating parameters, the size and cost of the
system components are closely related to the operating temp-
eratures and pressures. Thus, as a practical matter, the
ultimate used of the system, e.g., for vehicle motive power
wherein size may be critical or as a home power source
wherein size may not be very important, may influence the
physical size of the system and thereby limit the choice of
operating temperature and pressure parameters.
Environmental heating and cooling is a major user
of conventional energy sources and represents an area in which
the present invention is particularly applicable. The most
efficient conventional method of accomplishing enviromental
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110C~317
heating and cooling is via the conventional compressor powered
air conditioner for cooling and its counterpart, the heat
pump, for heating. It is not uncommon ~or such a system to
deliver more than 2.5 units of heat energy for every unit
of shaft power energy it uses. This cycle is essentially
the reverse of the common heat/power cycle as power is its
input and heat is its output, It is similar to the common
heat/power cycle in that its working fluid is processed be-
tween a high and low temperature and heat energy content
points (Tl and T2). It differs from the common heat/power
cycle because in practice, the extraction of useful shaftwork is omitted during the expansion of the.working fluid
between Tl and T2. In accordance with the present invention,
the efficiency of the conventional refrigeration/heat pump
cycle can be improved and the energy input requirements for
operating the cycle reduced. In so doing the quantity of
energy source.material consumed will be appreciably reduced,
thus reducing the overall cost to the consumer of environ-
mental heating and cooling,
In order to appreciate the significance of the pre-
sent invention in refrigeration/heat pump type applications,
it will be useful to review briefly a conventional mechanical
vapor refrigeration system (a discussion of its heating coun-
terpart, the heat pump, is therefore omitted). The basic
elements of such a refrigeration system 300 are shown in
Figure 4 and include a compressor 302, a condenser 304, an
evaporator 306 and an expansion valve 308, A conventional
refrigerant ~apor at relatively low pressure is drawn into
the compressor 302 to raise its pressure and temperature to
a level which, taking into consideration the reasonable
availability of a cooling medium, will permit heat to be
rejected in the condenser 304. Generally, the compressor
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110~317
302 will superheat the vapor for this purpose. In the condens-
er 304, the superheated vapor is cooled to a saturated or sub-
cooled liquid condition by a water or air cooling medium, The
saturated or sub-cooled liquid passes to the expansion valve
308 wherein it is throttled down to evaporator pressure at
essentially constant enthalpy in order to reduce the satur-
ation temperature of the liquid. Simultaneously, there occurs
an unavoidable flashing of a fraction of the liquid with the
result that the fluid leaving the valve is not all liquid.
This flashing is, of course, an undesirable incident of the
expansion since the vapor produced by the flashing has
~lready absorbed heat and is essentially useless as arefriger-
ant in the evaporator. The liquid-vapor mixture may be
permitted to pass directly to the evaporator 306 or, in some
instances, the vapor is bled off to lessen the flow load to
the evaporator. In either case, the useful refrigerant flow
to the evaporator, (or space to be cooled) is a fluid at
relatively low pressure. In the evaporator 306 the liquid
absorbs heat rom the space to be cooled and vaporizes to
a saturated or superheated vapor which becomes the feed stream
to the compressor 302 in the next refrigeration cycle.
The thermodynamics of the basic vapor refrigera-
tion cycle, expressed in temperature-entropy coordinates,
is shown in Figure 5 for an idealized refrigeration cycle.
Although it is appreciated that the idealized refrigeration
system depicted in Figure 5 is not possible of attainment it
establishes a criterion of maxiumum performance against which
actual systems can be measured. The idealized system envisions
four thermodynamic processes corresponding to the processes
outlined in connection with Figure 4. Thus, a liquid refriger-
ant at a undergoes an isentropic change to a liquid-vapor
mixture by expending along path ab. Heat is isothermally added
Pd/~ 17-
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to the refri~erant during evaporation along path bc during
which useful refrigeration or cooling is obtained. The pOillt
c assumes a dry compression cycle, i.e., one in which the
compressor suction vapor is dry or slightly superheated. The
vapor is compressed isentropically along path cd to a high
enough pressure to permit heat rejection in the condenser along
path dea. Initially, vapor superheat along path de is reject-
ed after which the heat of vaporization is rejected along ea.
In this idealized system both heat absorption in the evapor-
ated bc and heat rejection in the condenser dea are presumedto ~e constant pressure processes.
Referring to Figure 5, it can be seen that the heat
added to the system, Ql' at Tl, which is a measure of the
idéalized refrigeration achieved, is represented by the area
bczy. Heat is rejected at T2 in an amount Q2 represented by
area deayz. Therefore, the net work which must be provided
by an external power source, represented by area deabc, can
be expressed as
Net Wark = WN = ~ (Ql Q2)
In a conventional refrigeration system, all of this work is
supplied as input power to drive the compressor.
The coefficient of performance ~COP) for a refrig-
eration system is known to be the ratio of the refrigeration
effect to the work required to produce it, or
COP = Qr
WN
For the idealized refrigeration system, Qr/WN can be written
as
COP = Ql
Q2 Ql
Of course, an ideal refrigeration system is not
attaina~le in practical application and the expansion (thrott-
ling)-valve neither performs external work nor actually ~unc-
. ~
pd/~ 18-
1100317
tions without the gain or loss of heat through the pipe or
valve walls. Therefore, dotted line af in Figure S more
relistically shows the performance of an actual throttling
valve. Inasmuch as the refrigeration effect performed by the
system is represented in the ideal case by area bczy and in
the throttling case by area fczx, the area bfxy represents
the loss of useful refrigeration in departing from the ideal.
Now, it is well known that the idealized throttling
process, visualized as a constant enthalpy process, i.e., a
flow process which takes place adiabatically without work
production, is disadvantageous in that it foregoes the oppor-
tunity to extract from the expanding refrigerant at least
some of the work that was supplied to it during compression.
For example, if an expander were employed instead of a
throttling valve, as been sug~ested (see, e.g., Macintire et al,
Refrigeration Engineering, 2d Ed.~, and the pressure drop
thereacross presumed to occur isentropically as shown in
Figure 5, a substantially fraction of the work done by the
expanding fluid could theoretically be used to furnish work
input to the compressor. Of course the price to be paid for
the ability to recover this work is the expense of the equip-
ment constituting the expander. Nevertheless, viewed strictly
from the energy conservation standpoint, substituting an
expanding engine for a throttling valve is a viable means for
improving the cycle efficiency as represented by an improved
coefficient of performance. This much appears to have been
appreciated by many over the years, yet there seems to be no
evidence that this theoretical possibility was ever developed
into an efficient and economical refrigeration system.
To be sure, the use of an expanding engine or
turbine has been tried in the past. See, for example, U.S.
patent No. 1,575,~19 - Carrier and ~.S. patent No. 2,519,010 -
pd~ " -19-
- llOV317
Zearfoss. However, in the system described in these patents,
as in all heretofore suggested systems, the disadvantages of
a throttling process are not eliminated, Instead, they merely
recognize that, as a practical matter, there is a working
fluid stream velocity increase which attends throttling and
that by interposing a turbine in the flow path of the stream,
some of the stream kinetic energy can be transferred into
mechanical energy which can at least theoretically be used
in the operation of the system. The result is that the use-
ful energy having potential for doing work exiting a thrott-
ling device and delivered to the turbine is minimal, in
reality 30% or less of the work potentially retrievable, and
certainly below the level at which the increased capital
costs attributable to the turbine could be amortized over a
reasonable equipment life and be offset by any energy savings.
Figure 6 illustrates the conventional vapor refrig-
eration cycle modified in accordance with the present
invention in which an accelerating nozzle 320 and a turbine
322 have been installed in lieu of a throttling valve. The
nozzle 320 increases the kinetic energy or velocity of the
fluid and impinges the fluid flow upon a properly designed
furbine 322 through which the fluid expands and cools as it
did in the throttling valve. This system differs from those
of the prior art in its intentional omission of throttling,
i.e., the ommission of an adiabatic, constant enthalpy
expansion in favor of an expansion which is substantially
isentropic in nature or at least closely compares with the
isentropic case. The expansicn of a fluid through a nozzle
is accomplished by continuously varying the flow area along
the nozzle length and permitting the pressure and velocity
of the stream to adjust. As a result, in the nozzle, the
stream will exhibit a maximum velocity and minimum enthalpy
20-
~10~317
at the nozzle exit, Choking can be avoided if local acoustic
velocities are not achieved. By contrast, choking is normal
at the exit of capillary tubes and flow restrictions result.
As a consequence, for equal flow inputs and identical initial
conditions, it has been found that the temperature of the two
phase flow stream prior to leaving a nozzle will be lower and
the velocity thereof higher than for the same stream exiting
a capillary tube. Where work is removed from the fluid stream,
as in a turbine, the system coefficient of performance is,
therefore, increased more by a nozzle than by a capillary
tube. In accordance with this embodiment, a far larger por-
tion of the expansion work can be converted to useful shaft
work which may advantageously be employed to reduce the
amount of external power needed to operate the compressor,
fans and pumps. Since the useful work developed by the
present system is so much greater than that delivered by
proposed prior art systems, the present system is believed
to be economical in the sense that energy savings exceed
amortized capital costs in a relatively short period of time.
If the expander or turbine could actually be op-
erated in an isentropic fashion, then dotted path ab in
Figure 7 would represent the expansion path of the refriger-
ant between the condenser and the evaporator. As a practic-
al matter, however, reduced turbine efficiency a~d other
factors make isentropic operation unrealistic and line ag
is a more practical indication of the path followed during
expansion through a turbine having an efficiency less than
100%. The refrigeration effect of this cycle is represent-
ed by the area ~ . Since the refrigeration effect, Qr
using a nozzle and turbine is increased compared to using
a throttling valve (area fczx) and inasmuch as the net work
input, WN, is decreased, the coefficient of performance is
pd/~ , ~ -21-
110~317
improved according to the expression:
COP = Qr
WN
In Figure 4, the area ~fxw represents the increased refrigera-
tion effect attributable to replacing a throttling valve with
a nozzle and turbine.
In a practical circumstance wherein a commercial
flurocarbon refrigerant is employed in a mechanical vapor
refrigeration cycle between an evaporator inlet liquid temp-
erature of 40F and a condenser liquid temperature of 160F
and assuming overall turbine and compressor efficiencies of
80%, the coefficient of performance using a throttling valve
can be calculated as about COP=2.26. By-comparison, where
a nozzle and turbine are used in lieu of the throttling valve,
the calculated COP is 3.04, which reflects an improvement
of about 34~.
~ he failure of a throttling valve to capture and
use fluid expansion wor~ is believed to waste, at assumed
equipment efficiencies of 80%, a quantity of work equivalent
to about 34% of the theoretical input work needed to operate
the cycle. Assumin~ more realistic combined nozzle-turbine
efficiencies averaging about 45~, which correspond to those
already experimentally attained, it is believed that about
20~ of the theoretical input work to the cycle can actually
be captured from the re~rigerant expansion by using a nozzle/
turbin~ configuration in lieu of a throttling valve. Applying
this 20~ improvement to the many billions of dollars expended
annually for energy sources used in environmental heating
and cooling amounts to a véry substantial national savings.
The savings to the ultimate energy consumer will li~ewise
be substantial. Although the initial equipment cost ~or
conventional mechanical vapor refrigeration/heat pump equip-
.,'
pd/~ 22~
~la~3l~
ment modified to include a nozzle-turbine configuration in
accordance with the present invention may be as much as 10%
greater than present costs for conventional equipment, this
additional cost would be repaid by the anticipated energy
savings in less than two years.
It can, therefore, be appreciated that best results
are achieved by employing a non-throttling nozzle configura-
tion or at least one which chokes under conditions which still
allow it to outperform, in terms of increased retrievable
work, a capillary tube or like pressure reducing means under
similar initial working fluid conditions, and by designing
the nozzle and system to approximate isentropic, rather than
constant enthalpy, operation. The nozzle configuration may
be converging, converging-diverging, or diverging, i.e., it
may generally encompass any of the configurations contemplated
by the description herein. The system operating parameters
generally maintain the pressure close to saturation and the
working fluid is typically cycled through pressure changes
which cross the fluid's saturation point. ~t times, the fluid
may exist, for short periods, in physical states which do
not correspond with the saturation conditions at the time,
e.g., a liquid failing to vaporize instantaneously as pressure
decreases below the saturation point. These, and other fac-
tors may cause metastable flow conditions and attendant er-
ratic and undesirable fluid stream behavior. For example,
expansion of a liquid may occur in a radial, rather than a
downstream, direction or the volume produced by a phase change
may reduce the effective flow area. Any such deviation in
behavior from design based upon the configuration of the ECT
reduces the potential for recovering work from the system and,
generally, impairs the usefulness of the system for its
intended function. Thus, it is particularly desirable to
Pd/d -23-
llG~317
avoid metastable flow conditio~s. To this end, the use of
means and methods, such as are hereinbefore described, for
disruptin~ metastable flow are recommended for usage in con-
nection with the ECT in this application as well.
Still another application for the present invention,
although not necessarily one which is commended by its energy
savings capability, is the installation of the ECT to control
flow in fluid container safety relief valves. Typically,
railroad tank cars transporting hazardous materials such as
propane, butadiene, ammonia, ethylene, vinyl chloride, and
the like utilize safety relief valves which include means
for sensing over-pressure conditions in the tank and means
for automatically opening to permit out-flow from the tank
to a lower pressure environment, generally to the ambient.
The valves are intended to open when the pressure in the
closed tank car reaches a predetermined value less than the
tank bursting pressure~ It has been found by a study into
the nature of tank car accidents that the relief valves pres-
ently in use are inadequate because thay are not able to deal
with the two phase flow ~onditions which are created when
there is violent boiling within the tank following a tank
car accident. The inability of tank car relief valves to
efficiently handle two phase flow at high flow rates coupled
with their generally inadequate sizing has created a problem
-~f considerable proportions. In accordance with the present
invention, a tank car safety relief valve including an ECT
coupling the tank car and the environment to which the tank
car contents is to ~e relieved is provided which will permit
efficient relieving of large volumes of two-phase fluid streams
under predictable and efficient conditions. Particularly
when the ECT includes means for disrupting metastable flow
therein, as has been previous~y descri ~ , the usefulness of the
Pd/d~ -24-
'317
ECT in safety relief applications is enhanced. In this
connection, it will be appxeciated that maximum energy removal
and high flow velocity and mass flow rate are among the prime
objections of such a valve. Therefore, any erratic flow
stream behavior which detracts from achieving these prime
objectives, such as radial rather than downstream directed
expansion, or effective flow area reductions, which might
result from the existence of metastable flow conditions, must
be avoided in this type application even more so than, for
example, in refrigeration cycle applications. It will there-
fore be appreciated that the configuration of the ECT is tied
closely to its intended application.
While the present invention has been described with
reference to particular embodiments thereof, it will be under-
stood that numberous modifications can be made by those skill-
ed in the art without actually departing from the scope of the
invention. Accordingly, all modifications and equivalents
may be resorted to which fall within the scope of the inven-
tion as claimed.
pd/~. -25-
