Note: Descriptions are shown in the official language in which they were submitted.
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TECHNICAL FIELD OF THE INVENTIO_
The field of art to which the invention pertains is auto-
motive including engines, air conditioning, intercooling, turbo
charging, aftercooling, fuel cooling and charge cooling.
BACKGROUND OF THE INVENTION
Waste heat from automotive (heat) engines is an essential
consequence of the thermodynamicsof converting the chemical energy
of the fuel into mechanical work. In fact, in a practical engine,
the greater portion of the fuel heating value must be removed from
the engine proper as a result of which the essential engine exhaust
and cooling requirements arise. While this heat is practically un-
available directly for producing further work within the engine it-
self, it can be made available indirectly by various means for re-
latively high level uses both within the engine, as by compounding
or turbocharging, and external to the engine, as by heating, re-
frigeration (air conditioning) or other auxiliary power applications.
In the first use, the waste heat provides the power to alter engine
input conditions or to add recovered power to the output, resulting
in increased power and/or fuel economy. The latter case represents
a saving in engine power by substituting a waste-heat-powered by-
product function for one which would otherwise take some additional
prime engine shaft power and fuel to produce it.
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The prese~t invention not only combines improved engine
cooling with a system for waste-heat-driven heat pumping but it
also permits the two types of waste heat performance augmentation
to be combined. In particular, high-rate, isothermal jacket cooling
is obtained by nucleate boiling heat transfer to generate a
motive fluid for a jet-type thermo-comprPssor operating in true
Rankine power cycle fashion. The power produced therefrom con-
stitutes the actuation of a reverse Rankine cycle (vapor compression)
heat pump which can be used to heat or cool various media. A par-
ticularly important application is the cooling and heating of the
engine intake charge. Ano~her is the cooling and heating of vehicle
passenger or cargo environments. Some examples will illustrate the
utility of the invention.
Private automobiles represent one of the largest single
classes of energy consumers in the United States. A significant
portion of fuel consumption of automobiles is due to the growing
use of air conditioners. Additionally, air conditioning equipment
constitutes a significant fraction of the initial cost of the pri-
vate automobile. Such equipment typically is provided with a re-
ciprocating piston compressor which is mechanically coupled to the
engine and is generally complete in itself as an add-on component.
The additional vehicle weight and engine power involved requires
more fuel consumption and increases fuel consumption by as much as
15~.
On the other hand, waste heat from automotive engines is
available virtually free of cost and is responsive to relatively
high energy level uses. Within minutes, even at idle, tail pipe
temperatures exceed 500F. Fully 40% of the heating value of
total fuel consumption of a spark ignition engine exits at the tail
pipe~. Another 35 to 40% leaves by the cooling system. At cruising
conditions, exhaust gas temperatures exceeding 1200F are produced.
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Engine material temperatures must be maintained at between 150 anc
350~F by the cooling system to ensure engine integrity.
Accordingly, it would be desirable to provide some means
of high rate utilization of the engine waste heat to take economic
advantage thereof. The present invention provides a means for
effectively and economically controlling engine temperatures while
utilizing the otherwise wasted heat generated by an automotive ensfne
to produce refrigeration and heating for passenger comfort control
without requiring additional shaft power from the engine. Application
of the present invention to passenger automobiles can obtain milease
improvements of as much as 2 to 4 mpg when air conditioning is in
use.
Turbocharging internal combustion engines produces spec-
tacular gains in specific power. However, the results of charge
compression can be greatly offset by the attendant charge tempera-
ture rise from compression. High charge temperatures not only
diminish the power recovery of turbocharging but also intensify
engine cooling requirements inasmuch as the higher engine inlet
temperature will cause a greater portion of the recovered exhaust
energy to flow as heat into the engine parts rather than contribute
to shaft power. A further consequence is an increase in engine
octane requirement if a spark ignition type engine is involved.
Charge cooling, i.e. intercooling, aftercooling or eva-
porative cooling, is usually employed in high performance turbo-
charging applications. Evaporative cooling, e.g. the use of excess
fuel or water injection as in racing and aircraft applications, can
result in excessive fuel consumption and/or power loss. Inter-
cooling, sometimes called aftercooling, may require extensive heat
transfer surface and volume in the intake system. This usually
involves bulky and expensive heat exchangers and the volume of
intake charge hold-up may be increased to the point where engine
throttle response is adversely affected. A lower temperature
medium and a heat rejection system are also required. If ambient.
l~Z65~3
air is the lower temperature medium to which the heat is to be re-
jected, then significant increases in vehicle drag, frontal area
and/or power may be required to achieve the required heat flow for
a significant degree of charge cooling.
Without intercooling, the torque gain from charge com-
pression has diminishing returns such that at 60 in. Hg. boost,
only about ha7f the potential is realizable. The greater the
pressure boost, the more effective is the intercooling.
In a marine engine application these difficulties are
easily avoided because of the ready availability of ambient w~ter
for heat rejection. Such a medium has the desired properties of
low temperature (usually lower than ambient air), high heat capa-
city, high heat transport rate and facile pumping characteristics.
However, in other transport applications where only an
ambient air heat sink is available, substantial difficulties arise
in accomplishing a significant amount of charge cooling by the
customary method. Engine jacket cooling water has been used as
an intermediate heat transfer medium in intercoolers but the
effectiveness is limited due to the engine cooling load which
must be shared in the heat rejection system (radlator) and also
because of the sensible temperature changes which accompany heat
transfer. A separate sensible coolant loop whether of the "direct"
or "indirect" type is similarly limited. Heat pumping by vapor
compression refrigeration overcomes the limitations of both the
ambient heat sink and the sensible temperature gradients. How-
ever, a significant amount of power is required to drive the heat
pump which, if derived from the engine shaft, would greatly offset
any thermal advantage.
1~23
ESCRIPTIO~ OF THE PRIOR ART
A review of the prior art reveals several attempts at
both engine cooling and waste heat utlli7ation. Perrine, U.S.
Patent No. 2,327,451, discloses a system for recovering some of
the engine cooling jacket waste heat from the conventional cooling
water leaving the engine, by interposing a heat exchanger/boiler
for refrigerant motive fluid in a jet ejector which drives a vapor
compression cycle for air conditioning. Because of thP intermediat~
heat transfer used, involving a separate refrigeranb working fluid,
boiler and condenser, Perrine's system can utilize only a fraction
of the jacket heat available and then only at reduced availability
(temperature level). Ophir et al, U.S. Patent No. 3,922,877, dis-
close a system similar to Perrine wherein engine exhaust gas in-
stead of jacket water is contacted in a separate refrigerant
boiler/heat exchanger. Since both Perrine and Ophir et al use
separate refrigerant condensers placed in the same air stream as
the engine radiator, both impair engine cooling. Newcombe, U.S.
Patent No. 2,353,966 shows a system for cooling engine cylinder,
oil, intake, charge or other medium directly with evaporating
refrigerant wherein a combination of engine exhaust and separate
combustion gases are used to motivate an ejector vapor compressor.
Newcombe's system places the engine cooling jacket at the lowest
temperature level in the system and employs vapor compression
heat pumping for the purpose o~ raising the temperature level
of heat rejection to ambient to reduce the size and frontal area
in air craft cooling applications. Extra fuel must be burned to
produce the motive fluid to drive the system. In Newcombe, U.S.
Patent No. 2,360,969, a similar system is disclosed, wherein the
ejector is used to augment the pressure and temperature of the
vapor produced in the engine jacket to drive a vapor turbine coupled
to a supercharger. In this case the heat recovered is used for
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charge compression instead of heat pumping. Allison, U.S. Patent
No. 2,40~,778, uses the engine jacket not as a boiler or evaporator
but as a preheater with the exhaust gas providing the heat for
evaporation to vaporize the xefrigerant motive fluid in an ej~ctor.
Like Ophir et al, the Allison system requires out-sized heat trans-
fer surfaces for the exhaust heated boiler and subjects the refrigerant
material to decomposition temperatures.
SUMMARY OF THE INVENTI_ .
In the present invention, the motive power for heat
pumping is derived from engine jacket heat recovery.
Specifically, in a cooling mode, vaporized refrigerant
fluid is conveyed via a jet ejector through a condenser/radiator
while a portion of the fluid is diverted from the condenser to a
heat exchanger which serves as a refrigerant evaporator. The
diverted portion is evaporated in the heat exchanger to produce a
refrigeration effect and is then returned to the condenser by
aspiration via the jet ejector. A high degree of efficiency can
be accomplished (when required) by heat exchange through crossing
vapor and liquid lines (recuperation) and by the use of the thermo-
static controls, as will be described. ~ high pressure vapor driven
rotary expander (e.g. turbine or piston engine) shaft connected to
a rotary compressor is the rotary equivalent of the static jet
ejector referred to above and shown hereinafter as the preferred
embodiment.
In a heating mode, the high pressure, hish temperature
vapor is conveyed in a reverse direction directly from the engine
to the heat exchanger and then to the condenser. In this mode,
the heat exchanger (conditioner) serves as a condenser to provide
a heating effect.
Cooling or heating is selected by appropriate valving.
When both modes are closed out of the circuit, the pump can circu-
late the refrigerant in a closed loop between the engine andcondenser/radiator via a pressure relief by-pass valve or other means,
e.g. the variable area con-:roi ejector shown in Fig. 11, limit
the maximum refrigerant pressure, and thus engine jacket temperature,
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to the saturation value corresponding to the pressure relief setti~.g.
Additionally, by the use of a normally open thermostatically con-
trolled valve, co~lant by-pass of the engine can accelerate warm-~?
on starting. When neither heating nor cc,oling modes dissipates
sufficient heat to maintain proper engine jacket temperatures, th~
excess refrigerant vapor may be automatically by-passed around the
ejector and/or heat exchanger directly to the condenser/radiator
which has sufficient capacity to reject all engine jacket waste
heat to ambient. The engine jacket temperature may also be main-
tained by use of the variable area nozzle ejector shown in Fig. 11.
Also, a thermostatic expansion valve is used to by-pass small amo~nts
of liquid to limit the maximum superheat leaving the engine/exhaust
jacket in applications where the refrigerant has a substantial
exposure to exhaust gases.
By means of the present invention, a charge cooling
effectiveness exceeding 80% can be attained without utilizing
additional shaft power, vehicle frontal area, excess fuel or an
extra disposable evaporant. At 60 in. Hg. boost, the compressor
discharge temperature would be typically about 360~F and without
any charge cooling, a little over 100% increase in engine torque
can be obtained provided detonation can be avoided and engine
cooling can be maintained. If 80% of the heat of compression is
removed by intercooling (80~ effectivenes$), the torque gain from
turbocharging could be almost doubled without suffering impractical
increases in octane and cooling requirements.
Another feature of the invention is that in addition to
charge cooling, a multiplicity of other refrigeration and heating
functions can be provided simultaneously by the addition of parallel
vapor and liquid circuits.
The present invention obtains a basic thermal advantage
over conventional direct and indirect intercooling methods using
an ambient air heat sink. The thermal advantage of the heat pumping
method disclosed is due to three basic novel features:
1. The latent heat transport mechanism minimizes
flow rates, heat transfer surface and volume,
,
l~Z~i523
and overall temperature differences required bet-
ween the coolant and the intake charge.
2. The heat pumping effect obtains a low coolant tem-
pera~ure in the intercooler anc a high coolant
temperature in the radiator.
3. Heat transfer in the engine jacket occurs at nearly
optimum conditions for nucleate boiling which gives
the highest possible heat transmission at the least
overall temperature variation providing significant
benefits to engine integrity in high performance,
high output applications.
In particular, the present invention combines an improved
automotive engine cooling s~stem with a Rankine cycle heat pump
system which utilizes waste heat from the engine to vaporize
refrigerant fluid which travels through a condenser and is pumped
back to the engine~ While the present system uses a smaller than
conventional coolant pump which may be either driven by the fan
belt or electrically driven, there are no additional high-speed
rotary or reciprocating compressor parts such as bearings, belts,
dynamic seals, lubricants, clutches, or other mechanical features
as presently required for air conditioning. Accordingly, wear
and leakage are eliminated as factors limiting the service life
of the equipment. Furthermore, as the engine coolant is caused
to undergo a liquid-to-vapor phase change in the cooling jackets,
high heat transfer rates to the refrigerant fluid will result in
greater uniformity of temperature of engine parts, thereby reducing
thermal distortions, corrosion, hot spots, scaling, lubricant
break-down and, consequently, engine wear. The high-rate heat
transfer characteristic should be particularly valuable in high
performance and rotary (Wankel) engines, especially when aluminum
construction is used. Further the accelerated warm-up characteristic
is valuable in reducing the exhau$t emissions of passenger automobiles.
Finally, the waste heat drive feature eliminates the shaft power
penalty of refrigeration for air conditionins or charge cooling and
thereby improves vehicle fuel economy.
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Broadly, my invention is a compound engine including
internal combustion within the cylinder-combustion chambers, and
external combustion represented by the cooling jacket space func-
tioning as a boiler t~ provide motive yapor for a Rankine cycle
power system to drive a reverse Rankine cycle heat pump. This
system is unique in that the engine coolant, the power cycle
working fluid and the heat pump working fluid are the same fluid.
.. , .. . . . . . . _
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BRIEF DESCRIPTION OF DRAWINGS
Figure 1 is a schematic diagram, partially in section,
of a charge cooling system for a high performance engine application
constructed in accordance with the present invention showing typi-
cal flows, temperatures and pressures occurring at various points
in the system;
Figure 2 is a schematic illustration of an air conditioning
system constructed in accordance with the present invention;
Figure 3 is a thermodynamic diagram of the processes
taking place when utilizing the present system, the cooling mode
being represented by a solid line and the heating mode by a dashed
line;
Figure 4 is a diagrammatic representation of the flow of
refrigerant fluid when both cooling and heating modes are closed
out of the circuit and only engine cooling is required;
Figure S is a diagrammatic representation of the flow
of refrigerant fluid during an air conditioning cooling mode;
Figure 6 is a diagrammatic representation of the flow
of refrigerant fluid during an air conditioning heating mode;
Figure 7 is a graph of relative torque and temperature
versus boost pressure showing the value of turbocharging and
charge cooling to-engine performance and indicating the importance
of intercooler effectiveness which can be obtained with the
present invention;
Figure 8 is a graph of turbocharged engine torque im-
provement versus intercooler effectiveness at various levels
of boost pressure showing more specifically the value of inter-
cooler effectiveness;
Figure 9 is a thermal diagram comparing the temperature
distributions in the components of the system constructed in
accordance with the present invention and those prevailing in
conventional intexcooler systems; and
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Figure 10 is a graph of the film coefficient of heat trans-
fer versus wall-to-~luid temperature difference for various fluids
in nucleate boiling at various conditions showing peak heat flux
and "burnout" temperature differences. Fig. 10 also shows the
regime of conventional water jacket heat transfer performance.
Figure 11 shows a modified jet ejector with variable area
nozzle, which may be used in one embodiment of my invention.
DESCRIPTION OF PREFERRED EMBODIMENTS
Referring now more specifically to the drawings, Fig. 1
shows a simplified schematic diagram of my present system as applied
to a turbocharged vehicle and Figs. 2-6 show a system combining
engine cooling with cabin air conditioning, as will be described
in detail hereinafter. The general concept of my invention is
shown in Fig. 1 with typical temperatures, pressures and flows in-
dicated for Refrigerant 11, a typical commercial refrigerant.
Fig. 1 shows how the present system is used to recover engine waste
heat to power the jet driven heat pump for cooling the compression-
heated charge as well as for auxiliary cooling functions such as
fuel, cabin air, etc. In order to efficiently use the heat pumping
capacity available at various temperatures, it is found preerable
to use a multiplicity of jet pumps, each designed to provide the
required lift for each evaporator, intercooler, auxiliary cooler,
etc.
In the system shown in ~ig. 1, a major part of a moderately
pressurized refrigerant liquid from the radiator (condenser) 100 is
pumped into an engine cooling jacket manifold 102 through line 104
incorporating an orifice with pump 106. The pump effluent passes
through a fuel orifice 105 which allows the pump to operate at a
suficiently higher pressure than the pressure in the jacket to
pérmit pump 106 to operate free of pressure fluctuation caused by
refrigerant boiling in the cooling jacket. The cooling jacket 102
functions as a boiler for producing a high pressure refrigerant vapor
and may be of conventional design or configured- to produce super
heated vapor. A minor part of the medium pressure refrigerant liquid
l~Z~;52~
from the condenser 100, flows throush line 108 and is expanded in
throttle valve 110 to a low pressure region via line 112 in the
heat exchanger (evaporator) 114. In this low pressure region, the
refrigerant liquid is evaporated in heat exchanger 114 to create
a cooling effect by which the charge compressor discharge fed
through line 116 in heat exchange relationship with the low pressure
refrigerant liquid through heat exchanger 114, is cooled.
The low pressure refrigerant vapor effluent from the
evaporator 114 is aspirated from the evaporator through line 118
by a jet ejector 120 which is driven by the super heated high
pressure vapor discharged from the engine cooling jacket 102 through
line 122. A portion of the superheated cooling jacket effluent may
be used to drive an auxiliary jet pump 124 which may be utilized in
connection with an auxiliary cooler, for example, an air conditioner,
fuel cooler, etc., shown in Fig. 1 as cooler 126. The jet pump 120
may be of conventional design though a variable area jet ejector as
shown in Fig. 11 is preferred to convert the maximum available
driving energy into kinetic energy for entrainment and exchange with
the secondary fluid over the widest range of operating conditions.
A variable area primary nozzle ejector shown in Fig. 11
contains a cone-shape plug or spike nozzle body 12 e which is posi-
tioned within the nozzle throat 130 and is axially moveable and
positioned with a servo actuator 132. Operation of the jet pump
can thus be automatically controlled by a simple direct-acting,
upstream-pressure and/or temperature position servo 132 operating
on the nozzle body 128 to match throat area to available flow.
Control of the nozzle plug 128 position is obtained by balancing
forces due to downstream pressure distribution, spring compression,
diaphragm resistance and upstream static pressure. The equilibrium
position of the nozzle plug 128 is adjustable by a spring pre-load
adjustment screw 134. The primary fluid inlet 136 and the secondary
fluid inlet 138 are modified in the variable nozzle jet pump shown
in Fig. 11 but serve essentially the same purpose as the conventional
pump 120 shown in Fig. 1. Although one form of differential pressure
- 13 -
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control is illustrated in Fig. 11, it is obvious that various othe~
system parameters could be used to acco~plish nozzle control.
The admixture of the hot, high pressure vapor from the
cooling jacket (boiler) 102 and the cooler low pressure vap3r from
the heat exchanger (evaporator) 114 in the jet pump results in a h~t
medium pressure vapor effluent in line 140 which is cooled and
condensed in radiator 100, which is basically an air cooled condenser.
From the condenser 100, the condensed liquid is stored in receiver
144 and from there recirculated through the engine cooling jacket 102
and evaporator 114 through line 146. The heat exchanger 114 may be
used to heat the fluid 115 passing in heat exchange relationship with
the refrigerant, by opening heating valve 117 in line 119 to pass
hot refrigerant vapors into and through heat exchanger 114 in which
case heat exchanger 114 would serve as a condenser. Such heating
may for example be useful for preheating an engine charge mixture in
extreme ambient or cold start conditions. A pressure relief by-pass
valve 148 may be positioned in line 150 to permit direct circulation
of the refrigerant in a closed loop between the engine cooling jacket
102 and radiator 100 to limit maximum engine jacket pressure and
consequently temperature.
Although my present invention is described herein in re-
spect to charge cooling (Fig. 1) and hereinafter in detail in respect
to air conditioning, both cooling functions may be combined in a
single system by provision of an auxiliary jet pump 124 and line
152 for providing hot, high pressure motive vapor from the engine
cooling ~acket 102. Medium pressure liquid from the condçnser (radi-
ator) may be ~hrottled through line 154 to auxiliary cooler 126 with
throttle valve 156 for evaporation and consequent cooling therein.
The cool low pressure vapor generated in the auxiliary cooler
(evaporator) is educted therefrom through line lS8 by jet eductor
124 and the admixed vapors recycled to the radiator 100 through
line 160. Typical fluid pressures and temperatures are shown in
Fig. 1 to further demonstrate the effectivenss of the present sys-
tem as used for engine charge cooling. The temperatures and pressureS
~lZ~;S23
shown on the drawing do not account for llne temperarture and
pressure losses.
Referring to Figure 2, the present sys~em incorporates tne
engine cooling system of an automobile, whereby the engine 14 and
condenser/radiator 16 of an automobile respectively vaporize and
condense refrigerant fluid carried in a closed circuit. The head
and block cooling jacket 20 of the engine 14 consists of cores
(cavities) forming conduits therethrough in the conventional manner
and tubing 18 is co~nocted to the cooling jacket inlet to supply
refrigerant fluid thereto. The engine's exhaust manifold 22 may be
jac~eted, as is commonly done with marine engines. When the exhaust
manifold jacket is also cored to provide conduits for conveyance
of refrigerant fluid which is supplied from the head and block
cooling jacket 20 by tubing 24, exhaust heat is recovered only to
a small extent but this is valuable in obtaining additional vapor
super heat which can enhance ejector performance.
The system shown in Fig. 2 includes a circulation mode
wherein engine waste heat is dumped to ambient via the condensor
16. In the circulation mode, the system is neither cooling nor
heating the cabin, in which case waste heat i s insufficient to
remove engine waste heat at the rate required for proper engine
temperature control and must be removed by other means as shown
in Fig. 2.
Flows of refrigerant fluid in the circulation, cooling
and heating modes are shown schematically in Figures 4, 5 and 6,
respectively, and the schematic figures will be referred to in con-
junction with Figure 2 in describing each mode of the system's
operation. Additionally, the thermodynamic diagram of Figure 3
will be referred to in describing the cooling and heating modes.
Referring to Figures 2 and 4, in a circulation mode of
operation, solenoid valves 30 and 32 are closed so that only the
left side of the system, as depicted in Figure 2, is operative.
Initially, with a cool engine, refrigerant fluid driven by a pump
34 is pumped through tubing 38 (~ia a recuperator functional in
65z3
the cooling mode to be described) and bypass tubing 40 through a
normally open thermostatic valve 42. Bypass through the therm~sta'ic
valve 42 limits engine cooling so as to accelerate engine warm-up.
The thermostatic valve 42 senses the engine temperature, as indi-
cated by the dashed line 44, and closes after engine warm-up so
that the refrigerant fluid then passes through the engine cooling
jacket 20 and jacketed exhaust manifold 22 via the tubing 18 and
24 thereto.
The refrigerant fluid is vaporized by the engine cooling
jacket 20 and jacketed exhaust manifold 22 and, as vapor, travels
through tubing 46 to a pressure by-pass valve 48. The pressure by-
pass valve 48 is responsive, as indicated at 50, to pressure in
the tubing 46 to open and permit refrigerant vapor to be conducted
through tubing 52 to and through the condenser 16 via a four-way
tubing intersection 54. As a result of closure of the valves 30
and 32, the only pathway through the intersection 54 is to and
through t,he condenser 16. As the refrigerant fluid passes through
the condenser 16, it is condensed and accumulated at a receiver 56
and then travels from there through tubing 58 to the pump 34 to
repeat the cycle, while heat from engine cooling is rejected to
ambient air 89, drawn through the condenser/radiator 16 by the
engine driven fan 91. The refrigerant fluid vapor may be superheated
in the exhaust manifold jacket 22, as indicated at point 1 in
Figure 3. Control of the degree of superheat can be obtained by
a modulating thermostatic expansion valve 93 responsive to the
vapor temperature, as indicated at 94, by a liquid charged capillary
sensor to admit refrigerant liquid directly into the exhaust mani-
fold jacket as required to maintain a preset increment of superheat
regardless of the proportions of heat input received in the engine
jacket and the exhaust manifold jacket, respectively.
A conditioner 26 is provided, including an electrically
driven (by means not shown) fan 28 which blows air to be comfort
conditioned across coils ~not shown) constituting the conditioner
26. The conditioner 26 serves as an evaporator during a cooling
., .
l~Z65Z3
mode of operation and as a condenser during a heating mode, as
will be described in more detail hereinafter. By appropriate
operation of solenoid valves 30 and 32, as hereinafter described,
vaporized refrigerant fluid, e.g., trichloroCluoromethane
(Refrigerant 11), is conveyed either from the condenser 16 to the
cunditioner 26, in a cooling mode, or from the engine 20 to the
conditioner 26, in a heating mode. The pump 34 is either mechani-
cally driven by the engine 20, as schematically indicated at 36,
or is electrically driven by means not shown, and delivers con-
densed refrigerant fluid from the condenser/radiator 16 to the
engine 20.
~he cooling de of operation can best be described by
referring to Figures 2 and 5 together with Figure 3. In Figure 3,
the pressure and enthalpy at various point locations of the appa-
ratus are plotted with respect to the vapor/liquid equilibrium
line 60 for the refrigerant fluid. Points to tne left of the line,
e.g., points 6, 7, 11 and 12 (and point c for the heating mode) re-
present locations in apparatus at which the refrigerant fluid is in
a liquid state. The regions within the curve, e.g., point 8 (point
d for the heating mode), represent points in the apparatus at which
the refrigerant fluid is a mixture of vapor and liquid. Points to
the right of the curve, e.g., points 1-5, 9 and 10 (point a in the
heating mode) represent points in the apparatus at which the re-
frigerant fluid is superheated vapor.
For operation in the cooling mode, the system includes
an aspirator in the form of a jet ejector 62 connected to the
valve 30 by tubing 64 and conveying refrigerant vapor to a heat
exchanger or recuperator 66 through tubing 68. Jet ejectors are
known; see ior example the article entitled "Experimental Research
on a Fluorinated Hydrocarbon Jet Refrigerant Plant" by Cavallini
et al, appearing in "Progress in Refrigeration Science
and Technology", Vol. 2, (1967), pages 1225-1238. The
heat-exchanger 66 is referred to in the drawing as a high
(pressure) side recuperator and permits recoverv, from the
1126523
.
high (pressure) side of the system, of the residual heat of com-
pression by transfer to the engine coolant (refrigerant) feed
liquid. The refrigerant vapor is then conveyed past the inter-
section 54 through the condenser 16, to the receiver, and, for the
most part, back to the pump 34.
A minor portion of the refrigerant liquid from the con-
denser 16 is diverted through the ~onditioner 26 by means of tubing
constituting a parallel circuit across the ejector 62 and condenser
16. The parallel circuit includes a low (pressure) side recuperator
70 through which a portion of the condensed refrigerant fluid is
led via tubing 72 and then via tubing 74 through a thermostatic
expansion or throttle valve 76 and tubing 78 to and through the
conditioner 26. The thermostatic expansion valve 76 is controlled
to increase delivery in accordance with refrigerant superheat tem-
perature increases sensed at the refrigerant exit of the conditioner
26 as indicated at 80. The conditioner 26 functions as an evaporator
to vaporize the refrigerant fluid,:thereby cooling the conditioned
air 81 circulated therethrough by means of the fan 28.
Tubing 82 and 84 connects the conditioner 26 via the low
side recuperator 70 to the ejector 62 so that flow of the primary
stream of refrigerant through the ejector 62 induces flow from the
system's low pressure side to aspirate refrigerant vapor from`the
conditioner 26. The three-way solenoid valve 32 is in a disposition
rotated 45~ clockwise from that shown in Figure 2 so as to direct
the refrigerant vapor through the tubing 82. The refrigerant vapor
is drawn through the low side recuperator 70, for heat exchange with
and subcooling of the condensed refrigerant liquid from the receiver
56, and then through the tubing 84 into the ejector 62, to complete
the circuit.
- In fluid flow terms, and referring more closely to the
points 1-12 on the thermodynamic diagram of Figure 3, refrigerant
vapor heated by the engine cooling jacket 20 and jacketed exhaust
manifold 22 enters the ejector 62 primary nozzle via the solenoid
valve 30 at 1 and is supersonically expanded to the system low
6SZ3
pressure level at 2, inducing flow from the system low pressure
side 9 to evacuate the evaporator of refrigerant vla the low side
recuperator 70 at 7. The two streams combine a~d mix in the ejector
62 at 3 and diffuse to the intermediate pressure level of the
system at 4 through the high side recuperator 66. Some of the
residual heat of compression from the high side is recovered by
transfer to the engine feed liquid in the high side recuperator
66 at 11-12 prior to the refrigerant fluid entering the condenser
16 at 5 where the remaining superheat, latent heat and perhaps
some additional sensible heat (subcooling) is extracted by ambient
air to condense the refrigerant at 6 which is then held up in the
receiver 56. The major portion of this subcooled refrigerant
li~uid is returned to the engine by the pump 34 at the system
high pressure level 11-12 via the recuperator 66. The remainder
passes through the low side recuperator 70 for further subcooling
at 7 by heat exchange with the vapor leaving the conditioner 26
and is throttled to the conditioner 26 at the system 1QW pressure
level at 8 to make available its latent heat of evaporation at a
low temperature to produce a refrigeration effect.
Referring now to Figure 6, in conjunction with Figures 2
and 3, with respect to the heating mode of operation, the solenoid
valve 30 is closed and the three-way center closed solenoid valve
32 is rotated 45 counterclockwise from the position shown in
Figure 2 so that the ejector 62 is closed out of the system. In
this mode, refrigerant fluid, vaporized by the engine cooling jacket
20 and jacketed exhaust manifold 22, is conveyed through tubing 86
to and through the conditioner 26, now in a reversed direction.
The conditioner 26 now serves as a condenser with the refrigerant
producing a heating effect in giving up its latent heat of con-
densation to the air passing therethrough by force of the fan 28.
The partially condensed refrigerant is then conducted through the
tubing 78 but by-passes the expansion valve 76 via tubing 88 and
a check valve 90. The refrigerant is ducted by tubing 92 to the
~ubing intersection 54 and from there is conveyed to and through
llZ~5Z3
the condenser 16. The refrigerant is returned to the high pressure
- side via the receiver 56 and pump 34 in the same manner as described
above with respect to the other modes of operation.
In fluid f70w terms, and referring more c].osely to the
points a, b and c in the thermodynamic diagram of Flgure 3, super-
heated refrigerant vapor at point a, adjacent the three-way solenoid
valve 32, is conducted through the conditioner 26 where it partially
condenses, travels to point b, through the check valve 90, through
the condenser to poi~t c then via the pump 34 back to the engine.
It will be appreciated that one can simply switch from a
cooling mode to a heating mode, or to neither cooling nor heating,
by simple actuation of solenoids controlling the valves 30 and 32.
Modulation of the cooling capacity can be effected by cycling
valve 30 on demand from an air temperature thermostat (not shown).
Modulation of heating capacity can be similarly effected by cycling
valve 32 with valve 30 closed. When neither heating nor cooling
of conditioned air is called for, both valves 30 and 32 are closed
to all ports and the high side pressure (and temperature~ is per-
mitted to float upward to the pressure relief setting of the pressure
by-pass valve 48. Enough refrigerant is circulated directly to the
condenser/radiator 16 to dump all the engine 14 waste heat to
ambient via the condenser/radiator 16 and the ambient air circula~ed
therethrough by force of the fan 91. The refrigerant pump 34
operates continuously while the engine is operating and together
with the pressure bypass valve 48 serves to limit the maximum engine
jacket temperature to an acceptable level~
The adiabatic efficiency of the ejector 62 and the overall
coefficient of performance (COP) of the system can be calculate~
utilizing hypothetical cycle conditions and refrigerant ~DuPont
Freon 11), as listed in Table 1.
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112~523
TABLE 1
Fig. 3 Temp., Pressure, Enthalpy,
Location ~ psia BTU/l~.
1 280 160 126
2 63 7 100
3 68 7 101
4 199 33 118
152 33 111
6 120 33 33
7 101 33 29
8 40 7 29
9 50 7 98
7 102
11 140 160 3~
12 173 160 45
The sross thermal ratio of the ejector 62 is:
qross compression energy
ne~ available motion energy (1)
lh4 - hlo) (Gl + Glo)
(hl - h2) G~ .
where subscripts (here and hereinafter) refer to locations in
Figure 2 and Figure 3 and where h - enthalpy and G - mass flow.
From an energy balance in the ejector 62,
G1ohlo ~ Glhl - (Gl ~ GlO)h4 (2)
Assuming a unit mass of low side vapor, Glo = 1,
from Eq. 2:
h4 ~ hlO
Gl
hl _ h4
;5Z3
vsing properties listed in Table 1:
&1 = 126 - 118 = 2;
then from Eq. ~1)
~ 102 3 = 0. 92
ne~ ~ 126 - loo X 2
and the adiabatic efficiency of the ejector 62 is:
fh4 ~ hlo . Gv~ 16
~hl - h2 Gs) 26 2
The coolina effect under these conditions is
Qref = (h9 - hg)Glo
= 69 X 1 = 69 BTU.
The net heat input of the engine to the refrigeration system
would be
Qin = (hl - h12)Gl
~ 81 X 2 = 162 BTU.
The overall coefficient of performance of the system would
be
COP ~ QQief = 6962 = 0.426.
Typical cooling capacity and a comparison with mechanical
vapor compression refrigeration can be determined by assuming
typical cruise characteristics of an automotive installation, as
listed in Table 2.
~126~23
TA3LE 2
.
Road horsepower (RHP) 30
Specific fuel consumption (SFC) 0.5
lb/RHP-hr.
Speed, mph 45
Fuel consumption, mpg 18
Fuel heating value (HV) BTU/lb 20,000
Fuel density, lb/gal 6
Exhaust temperature, D F1,O0O
Cylinder overall temperature F 250
Determining first the potential cooling capacity available
from the foregoing automotive installation, the total heat input
to the engine is
QF - R}IP X SFC X HV
= 30 % 0.5 X 20,000 = 300,000 BTU/hr.
of which about half is passed out via the exhaust system.
Assuming no more than 25% of the engine waste heat is
economically recovered by 173 to 280F refrigerant from 1000F ex-
haust gas and 250F cylinder walls, about 75,000 BTU/hr is added
to the refrigerant in the boiler (engine coolins jackets).
The net refrigeration capacity at the 40F evaporator (conditioner
20) is then
Qref ~ COP X Qin
= 0.426 X 75,000 = 31,950 BTU/hr
= 2.66 tons.
The net heating capacity at a 180F (80 psia) conditioner temper-
ature would be 75,000 BTU/hr.
~ _
112~i523
Conventional mechanical vapor compression refrigeration,
which has a COP, measured from the shaft, of 3 would require a
compressor shaft power input ~HPC) of approximately
HPc = Q54es COP = 2545 X 3 ~ 4.185 hp-
This additional load on the engine results in an increasein power requirements of
30 + 4.185 X 100 12.2%
and a comparable increase in fuel consumption.
Table 3 sets forth the typical physical characteristic
of the system of Figure 2.
TABLE 3
Pump pressure rise, psi 127
Pump shaft efficiency 0.5
Pump shaft horsepower (SHP) 0.178
Pump flow rate, gpm 1.2
Compressor-induced flow rate, cfm 45.2
High side recuperator effectiveness 0.6
High side recuperator number of
transfer units (Counterflow) 1.5
Low side recuperator effectiveness 0.57
Low side recuperator number of
transfer units (Counterflow) 1.3
Condenser/radiator capacity BTU/hr.
at 120F refrigerant 150,000
at 250F refrigerant 3,000,000
Evaporator/conditioner capacity BTU/hr.
at 40F refrigerant 32,000
at 150F refrigerant 85,000
~l'Z65Z3
The load of the pump on the engine results in a power re-
quirement of
30 - 0 178 X 100 = 0.59~
which approximately offsets the existing pumping requirement for
cooling only so that changes in parasitic sha~t power are
ne~ligible.
Accordingly, it will be appreciated that an effective air
cooling and heating system has been provided with a negligible shaft
power re~uirement Further, a sealed engine cooling system incor-
porating an evaporating fluid has been provided which virtually
eliminates scaling, corrosion and freezing while reducing weight
and bulk and increasing engine performance and life by improving
temperature control and uniformity.
While a preferred embodiment has been shown with a parti-
cular halocarbon refrigerant fluid, it is obvious that a variety
of other materials could be used. The refrigerant fluids useful
with the present system are chemically stable, constant-boiling
dielectric fluids, e.g. water, ammonia, inert organic materials,
for example, the commercial halocarbons, alcohols, etc., and certain
azeotropes of the foregoing.
Referring to Figure 7, the effectiveness of turbocharging
and intercooling in improving internal combustion engine performance
is shown. Also shown is the typical turbocompressor discharge
temperature characteristic which gives rise to the striking bene-
fits of intercooling. Removing all the heat of compression from
the compressed charge (perfect intercooling) more than doubles the
maximum torque available at 60 in. Hg. boost. As shown more parti-
cularly in Figure 8, doubling intercooler effectiveness (from 50%
to 100%) can increase the torque of an engine turbocharged to
60 in. ~g. gage almost three-fold.
11265Z3
If all other heat transfer factors are considered equal,
the effectiveness will depend on the temperature of the heat ab-
sorbing medium used in the intercooler according to the equation
E Th1 ~ Th2
Thl ~ T
where Thl and Th2 are the charge air temperatures in and out of the
intercooler and TC1 is the cooling medium temperature at inter-
cooler i~let. If TC1 is the ambient or initial charge air tem-
perature (before compression~, the effectiveness relation expresses
the degree of charge compression energy removed by the intercooler.
Coolant inlet temperatures lower than the ambient air value could
result in an intercooling effectiveness exceeding 100~ of the
charge compression energy removed. The heat pumping characteristic
of the present system is capable of producing such below-ambient
coolant temperatures which greatly increases its intercooling
potential compared with ordinary ambient air heat sink methods
(see Fig. 9).
Figure 9 depicts the temperature distribution in the
components of the present system applied to charge cooling as
compared with that of two conventional methods using ambient air
heat rejection. The value of the latent heat transport character~
istic of the present system is clearly indicated as having a sub-
stantial thermal advantage in achieving high overall temperature
differences in the intercooler which improve effectiveness and/or
reduce size. The ability of the heat pumping effect of the
present system to reduce coolant temperatures below ambient in-
creases the thermal advantage still further. Figure 9 shows some
other characteristics of the present system such as the nearly
isothermal temperature distribution of the coolant in the engine
jacket and condenser-radiator. This isothermal characteristic is
due to the latent heat capacity of the working fluid which is
changing phase ~evaporating in the engine jacket and condensing
in the radiator). As a result, higher average temperature levels
l~Z6S~:3
are obtained for heat rejection to ambient which are valuable for
reducing radiator size to a minimum. The isothermal c~ndition in
the engine jacket is valuable for obtaining uniform temperatures
in the engine parts which, in turn, reduces stresses and de-
formations and allows closer tolerances and greater structural
margins to be used and at higher material temperatures.
Figure 10 shows some typical heat transfer characteristics
obtaining in the engine iacket as a result of the nuclPate boiling
produced. By comparison, the regime of conventional convective
water cooling is also shown. The value of nucleate b~iling is
clearly indicated, showing more than 10 times the heat transfer
rate at comparable temperature differences. This characteristic
leads to further improvement in engine temperature uniformity and
structural integrity.
The use of boiling organic working fluids as engine
coolants has been found to offer a number of advantages. First,
they have somewhat lower critical pressures and at required
engine coolant temperatures, the boiling pressures are an appre-
ciable fraction of the critical pressure. This substantially
reduces the ratio of vapor to liquid volume which reduces the
volume fraction of vapor due to evaporation and allows a higher
level of surface wetting to be maintained as the boiling pro-
gresses. Purther, the organics possess low inter-facial tension
characteristics which produce low bubble contact angles which,
when combined with the small vapor volume changes, makes for
æmall bubble size and maintains the regime of nucleate boiling
over a greater range of conditions. This means that high peak
heat fluxes can be accommodated at lower temperature differences
over the wides possible range of operating conditions.
While my invention has been described herein with a
certain degree of particularity in reference to certain specific
embodiments, my invention is not to be limited to the details
set forth herein, but should be afforded the full scope and
equivalents of the appended claims.
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