Note: Descriptions are shown in the official language in which they were submitted.
CA 02610421 2007-11-22
1
DESCRIPTION
OIL-FREE LIQUID CHILLER
Background of the Invention
The present invention relates to liquid chillers. More particularly, the
present invention
relates to relatively large tonnage centrifugal chillers in which so-called
hybrid bearings are
employed and in which the lubrication of such bearings is by the refrigerant
which comprises the
chiller's working fluid. With still more particularity, the present invention
relates to oil-free, direct
drive centrifugal water chillers capable of achieving optimized part load
performance and in which
the cooling of the chiller's compressor drive motor is enhanced.
Refrigeration chillers are machines that use a refrigerant fluid to
temperature condition a
liquid, such as water, most often for purposes of using such liquid as a
cooling medium in an
industrial process or to comfort condition the air in a building.
Refrigeration chillers of larger
CA 02610421 2007-11-22
2
capacity (from two hundred or so to thousands of tons of
refrigeration) are typically driven by large centrifugal
compressors. At lower capacities, compressors of the screw,
scroll or reciprocating type are most often used in water
chiller applications.
Centrifugal compressors are compressors which, by
the rotation of one or more impellers in a volute housing,
compress a refrigerant gas for use in the chiller's
refrigeration circuit. The impeller or impellers of a
centrifugal compressor, the shaft on which they are mounted
and, in the case of so-called direct drive compressors, the
rotor of the compressor drive motor, weigh hundreds if not
thousands of pounds. The high speed rotation of such
physically large and heavy chiller components at several
thousand RPM results in unique and challenging bearing
lubrication issues, particularly at start-up when these
components are at rest, but also during chiller shutdown when
these components coast to a stop.
Centrifugal compressors are of the direct drive or
gear drive type. Hence, the chillers in which such compressors
are used are generally referred to as direct drive chillers or
gear drive chillers.
In direct drive chillers, the rotor of the
compressor's drive motor is mounted directly to the shaft on
which the compressor's one or more impellers are mounted. That
shaft, in turn, is typically mounted for rotation in one or
more bearings which are in need of lubrication when the chiller
is in operation.
CA 02610421 2007-11-22
3
In gear drive centrifugal chillers the shaft on
which the one or more impellers are mounted is driven through a
series of gears rather than by the direct mounting of the rotor
of the compressor drive motor to the shaft on which the
impellers are mounted. The gears of a gear drive chiller act
to increase the speed of rotation of the impeller beyond that
of the motor which drives the impeller and in so doing increase
the refrigeration effect or.capacity of the chiller. in the
case of a gear drive chiller, both the drive gears and the
bearings in which the impeller shaft rotates require
lubrication, heretofore by oil, and both direct drive and gear
drive chillers have most typically employed induction motors,
the speeds of which are typically limited to 3600 RPM.
It can generally be stated that chillers of the
direct drive type are quieter and more efficient than chillers
of the gear drive type. Further, chillers of the direct drive
type are viewed as being more reliable than present day
chillers of the gear drive type for the reason that chillers of
the gear drive type make use of multiple gears, more bearings
and other rotating parts, not found in a direct drive chiller,
which are susceptible to breakage and/or wear. Gear drive
chillers do, however, offer certain advantages in some
applications, including, in some instances, a cost advantage
over direct drive chillers.
In the cases of both direct drive and gear drive
large tonnage centrifugal chillers, lubrication of their
rotating components has historically proven both challenging
and expensive and has been exclusively or at least
fundamentally accomplished by the use of oil as the lubricant.
CA 02610421 2007-11-22
4
The need for such lubrication systems has vastly complicated
the design, manufacture, operation, maintenance and control of
centrifugal.chillers of both the direct drive and gear drive
type and has added great initial and operational cost to them.
Elimination of oil as a lubricant in a large
tonnage centrifugal refrigeration chiller system and the use of
the refrigerant which comprises the chiller's working fluid for
that purpose offers potentially tremendous advantages. Among
those advantages are: elimination of many chiller failure
modes associated with oil-based chiller lubrication systems;
elimination of so-called oil migration problems associated with
the mixing of oil and refrigerant in such chiller systems;
enhancement of overall system efficiency by eliminating the
oil-coating of heat exchange surfaces that results from the
entrainment of oil in system refrigerant and the carrying of
that entrained oil into a chiller's heat exchangers;
elimination of what is viewed as an environmentally unfriendly
material (oil) from the chiller system as well as the problems
and costs associated with the handling and disposal thereof;
and, elimination of a great number of expensive and relatively
complex components associated with chiller lubrication systems
as well as the control and maintenance costs associated
therewith.
Further, the elimination of oil as a lubricant in a
centrifugal chiller system suggests the possibility of a
centrifugal chiller that offers the advantages of direct drive
machines yet which, by virtue of variable speed operation, is
fully the equal of or superior to gear drive machines.
Heretofore, particularly good part load efficiencies have been
achieved in gear drive machines by the use of specially
CA 02610421 2007-11-22
configured gear sets capable of driving a chiller's impeller at
relatively very high and/or optimal speeds. As was noted
earlier, however, gear drive machines do not offer many of the
advantages of direct drive machines and their use brings
5 several distinct disadvantages, the need for an oil-based
lubrication system for the purposeof ensuring the adequate
lubrication of the gear train being one of them.
There have been and continue to be efforts to
eliminate the need for oil-based lubrication systems in
centrifugal chiller applications. Such efforts have, however,
heretofore focused primarily on specialized small capacity
refrigeration machines in which the bearing-mounted shaft and
impeller are relatively very small and lightweight and on the
use of hydrostatic, hydrodynamic and magnetic bearings in
applications where bearing loads are relatively very light. In
that regard, hydrostatic and hydrodynamic bearings are journal-
type bearings which, while relatively low cost, simple and
technically well understood, are intolerant of the momentary
loss or reduction of lubricant flow. The intolerance of such
bearings to the loss or reduction of lubricant available to
them is exacerbated in a refrigerant environment. Further,
such bearings detract from the efficiency of the compressor's
in which they are used as a result of the frictional losses
that are inherent in such bearings as compared to the
frictional loses associated with rolling element bearings.
While hydrodynamic and hydrostatic bearings
lubricated by refrigerant may have been at least prospectively
employed in specialized, relatively physically small capacity
compressors, the use of such bearings in large tonnage
centrifugal chillers poses significant difficulties due, among
CA 02610421 2007-11-22
6
other things, to the masses and weights of the chiller
impellers and shafts that must be rotationally started and
supported in that application. The sizes and weights of such
components are such as to present significant design
difficulties, particularly at chiller start-up and shutdown and
during momentary loss of lubricant flow, which are yet to be
overcome in the industry.
Further, even if such design difficulties are
capable of being overcome with respect to the use of
refrigerant-lubricated hydrostatic or hydrodynamic bearings in
large tonnage refrigeration chillers, the efficiency penalties
incurred in the use of such bearings due to the inherent
frictional losses associated with them is disadvantageous.
That disadvantage becomes larger and larger as real world
issues, such as global warming, drive the need for energy
consuming equipment to operate more efficiently.
Still further, the employment of hydrostatic
bearings is additionally disadvantageous as a result of the
need in such systems for a pump by which to deliver relatively
very high pressure liquid refrigerant to such bearings in the
absence of oil, the bearings of such pumps themselves requiring
lubrication in operation. Such high pressure pumps are seen to
be subject to breakdown and, potentially, pose an issue of
chiller reliability where hydrostatic bearing arrangements are
attempted to be used.
Even further and more generally speaking, the
employment of liquid refrigerant to lubricate bearings of any
type in the absence of oil in a chiller system presumes the
reliable availability of a supply of refrigerant in the liquid
state whenever t:~.e compressor is operating and the ability to
CA 02610421 2007-11-22
7
deliver such refrigerant to the bearings. However, there is
essentially no single location within a chiller that contains
liquid refrigerant that is capable of being delivered to such
bearings under all prospective chiller operating conditions in
a form or state that is appropriate for bearing lubrication.
In that regard, when a chiller is shutdown and even at very low
load conditions, liquid refrigerant will tend to be most
reliably available from the evaporator. When the chiller is
operating at load, the condenser is the most reliable source of
liquid refrigerant. Therefore, the prospective lubrication of
bearings by liquid refrigerant requires that an assured source
of liquid refrigerant be provided for whether the chiller is
shutdown, starting up, under very low load, operating at load
or is coasting to a stop after it is shutdown.
An exciting opportunity exists, (1.) to achieve all
of the advantages offered by direct drive centrifugal chillers,
(2.) to simultaneously achieve enhanced part load chiller
efficiencies, (3.) to eliminate the use of oil-based
lubrication systems and (4.) to increase overall chiller
efficiency, in the prospective use in refrigeration chillers of
rolling element, as opposed to journal-type bearings, where the
rolling element bearings are lubricated only by the refrigerant
which comprises the chiller's working fluid. The possibility
of eliminating oil as a lubricant in centrifugal chiller
systems has become a reality with the recent advent of so-
called hybrid rolling element bearings in which at least the
rolling elements thereof (which are significantly less
expensive than the bearing races to fabricate), are fabricated
from a ceramic material. Although such bearings have been
commercially available for a few years and although there has
CA 02610421 2007-11-22
8
been speculation with respect to the possibility of their use in relatively
very small refrigeration
chillers, their actual use has primarily been in machine tool applications and
in such applications,
lubrication of such bearings has been and is recommended by the bearing
manufacturer to be by
the use of grease or, preferably, oil.
Certain of the characteristics of such bearings have, however, suggested to
applicants the
possibility of a large capacity centrifugal refrigeration chiller which
eliminates the use of oil as a
lubricant and the substitution of the chiller's working fluid therefor, even
with respect to bearing
lubrication. Further, such bearings are particularly well suited for high and
variable speed
operation as a result of the relatively lower mass of ceramic rolling elements
as compared to their
steel counterparts, such reduced mass resulting in reduced centrifugal forces
within hybrid bearings
at high speeds which, in turn, results in a reduction in the forces the
bearing races must withstand
during high speed operation. The use of the chiller's working fluid as the
lubricant for such
bearings and the need to ensure the availability of such liquid for that
purpose from one source or
another under all chiller operating conditions does, however, present many new
and unique
challenges that must be overcome.
Summary of the Invention
It is therefore desirable to eliminate the need for oil as a lubricant in a
centrifugal
refrigeration chiller.
It is also desirable to provide a centrifugal refrigeration chiller in which
the bearings
thereof are lubricated, in a manner which adequately removes heat from the
bearing location, by
the refrigerant which comprises the working fluid of the chiller system.
It is also desirable to provide a centrifugal chiller in which the bearings
thereof are
lubricated by the liquid refrigerant which comprises the working fluid of the
chiller system and
CA 02610421 2007-11-22
9
wherein a supply of liquid refrigerant from one location or another within the
chiller is assured as
the chiller starts up, when it operates at very low loads, when it operates at
load and when it shuts
down and the compressor apparatus of the chiller coasts to a stop.
It is also desirable to eliminate oil migration problems and the need to
return oil from
chiller system heat exchangers to the chiller's compressor as a result of the
migration of oil to
those heat exchangers during chiller operation.
It is also desirable to, by the elimination of oil migration, increase chiller
system efficiency
by eliminating the oil-coating of heat exchange surfaces in the chiller
system's heat exchangers
and the resulting diminishment of the heat transfer process that results
therefrom.
It is also desirable to provide a centrifugal chiller which, by the use of
rolling element
bearings lubricated by refrigerant rather than oil, is of increased efficiency
as compared to systems
in which bearings of other than the rolling element type are used.
It is also desirable to eliminate an environmentally unfriendly material, that
material being
oil, from refrigeration chillers and to eliminate the need to handle and
dispose of that material.
It is also desirable to eliminate the many expensive and complex components
associated
with the lubrication by oil of centrifugal chiller components as well as the
failure modes and
manufacturing costs associated therewith and the costs imposed thereby in
terms of controlling an
oil-based chiller lubrication system.
It is also desirable to provide a centrifugal chiller which is capable of both
high speed and
variable speed operation so as to enhance system part load efficiency,
preferably using relatively
conventional and inexpensive induction motor technology.
CA 02610421 2007-11-22
It is also desirable to provide a cost competitive multi-stage, direct drive
centrifugal chiller
capable of part load performance equaling that of a gear drive chiller in
which the need for an oil-
based lubrication system is eliminated.
It is also desirable to provide an oil-free centrifugal chiller in which
system refrigerant is
5 available to the chiller's bearings in sufficient quantity, at all times
necessary and in the proper
state, to assure their adequate lubrication.
It is also desirable to provide an oil-free centrifugal chiller in which the
centrifugal forces
to which the bearings of the chiller are exposed, at high operational speeds,
are reduced by the use
of ceramic rolling elements which are of less mass than rolling elements used
in conventional steel
10 bearings.
It is also desirable to provide for enhanced cooling of the compressor drive
motor of a
centrifugal refrigeration chiller.
Accordingly a refrigeration chiller is disclosed wherein the shaft on which
the chiller's
impellers and drive motor rotor can be mounted is itself mounted for rotation
in so-called hybrid
rolling element bearings, such bearings being lubricated and cooled, in the
absence of oil, by the
refrigerant which comprises the chiller's working fluid. Apparatus can be
provided which ensures
that system refrigerant, in the appropriate state and amount, is available to
the bearings for
lubrication and heat removal purposes at chiller start-up, during chiller
operation and for a
sufficient period of time subsequent to chiller shutdown during which the
shaft on which the
chiller's impellers and drive motor rotor are mounted coasts to a stop and to
the compressor drive
motor for motor cooling purposes. Additionally, by the use of an induction
motor and a variable
speed drive capable, superior part load efficiency can be achieved, all in a
refrigeration chiller
having the reliability advantages offered by direct drive but which avoids the
efficiency and
, _ _ . _. .
CA 02610421 2007-11-22
11
reliability disadvantages associated with gear drive
machines and the need for an oil-based lubrication system
associated with the gear set thereof.
According to one aspect of the invention, there is
provided a liquid chiller comprising: a condenser, the
condenser condensing refrigerant gas to the liquid state
when the chiller is in operation; a metering device, the
metering device receiving refrigerant from the condenser; an
evaporator, the evaporator receiving refrigerant from the
metering device; a compressor, the compressor receiving
refrigerant from the evaporator and delivering refrigerant
in the gaseous state to the condenser when the chiller is in
operation; a motor, the motor driving the compressor and
being cooled by refrigerant sourced from the condenser; a
variable speed drive, the variable speed drive being
electrically connected to the compressor drive motor for
varying the speeds thereof, refrigerant being delivered from
the condenser to the variable speed drive so as to cool the
variable speed drive.
According to another aspect of the invention, there is
provided a method of operating the liquid chiller comprising
the steps of: connecting a condenser, a metering device, an
evaporator and a compressor for flow so as to form a
refrigeration circuit; flowing a cooling medium through the
condenser so as to carry heat thereoutof; driving the
compressor by the use of a motor; controlling the speed of
the motor by the use of a variable speed drive which
includes heat generating components; delivering refrigerant
from the condenser to the at least one of the compressor
drive motor and the variable speed drive; rejecting heat
from the at least one of the compressor drive motor and the
variable speed drive to the refrigerant delivered in the
delivering step; returning refrigerant to which heat has
been rejected in the rejecting step to the condenser; and
transferring heat carried back to the condenser in the
CA 02610421 2007-11-22
12
returning step to the cooling medium which flows through the
condenser.
According to another aspect of the invention, there is
provided a liquid chiller comprising: a condenser for
condensing refrigerant gas to a liquid state when the
chiller is in operation; a metering device arranged to
receive refrigerant from the condenser for reducing the
pressure thereof; an evaporator arranged to receive
refrigerant from the metering device; a compressor arranged
to receive refrigerant from the evaporator and deliver
refrigerant in a gaseous state to the condenser when the
chiller is in operation; and a motor for driving the
compressor, the motor being cooled by refrigerant sourced
from the condenser; and characterized in that a variable
speed drive is electrically connected to the compressor
motor for varying the speed thereof, refrigerant being
delivered from the condenser to the variable speed drive so
as to cool the variable speed drive.
According to another aspect of the invention, there is
provided a method of operating a liquid chiller that
comprises: a condenser, a metering device, an evaporator and
a compressor connected to form a refrigeration circuit; and
a motor for driving the compressor and a variable speed
drive, which is for controlling the speed of the motor and
includes heat generating components; the method comprising
the steps of: flowing a cooling medium through the condenser
so as to carry heat therefrom; driving the compressor by the
use of the motor; controlling the speed of the motor by the
use of the variable speed drive; delivering refrigerant from
the condenser to at least one of the motor and the variable
speed drive, which refrigerant receives heat from the at
least one of the compressor drive motor and the variable
speed drive; and returning the refrigerant which has
received the heat to the condenser such that the heat can be
CA 02610421 2007-11-22
13
transferred to the cooling medium which flows through the
condenser.
Description of the Drawing Figures
Figures la and lb are end and top views of the
centrifugal refrigeration chiller of the present invention.
Figure 2 is a cross-sectional view of the compressor
portion of the centrifugal chiller of Figure 1 illustrating
the primary components of the compressor.
Figure 2A is an enlarged view of the back-to-back
bearing arrangement of bearing package 50 of Figure 2.
Figure 3 schematically illustrates the chiller
lubrication system of the present invention.
Figure 4 schematically illustrates an alternative
embodiment of the chiller lubrication system of the present
invention.
Figure 5 schematically illustrates still another
alternate embodiment of the present invention.
Figure 6 schematically illustrates still another
alternate embodiment of the present invention.
Description of the Preferred Embodiment
Referring to Drawing Figures la and lb, a chiller 10,
which in the preferred embodiment is a centrifugal chiller,
and its basic components are illustrated. In that regard,
chiller 10 is comprised of a compressor portion 12, a
condenser 14 and an evaporator 16. A refrigerant gas is
compressed within compressor portion 12. Such refrigerant
gas is directed out of discharge volute 18 into piping 20
which connects the compressor portion 12 to condenser 14.
Condenser 14 will typically be cooled by a liquid
which enters the condenser through inlet 22 and exits
through outlet 24. This liquid, which is typically city
CA 02610421 2007-11-22
13a
water or water that passes to, through and back from a
cooling tower, exits the condenser after having been heated
in a heat exchange relationship with the hot, compressed
system refrigerant which is directed out of the compressor
into the condenser in a gaseous state.
The heat exchange process occurring within condenser
14 causes the relatively hot, compressed refrigerant gas
delivered thereinto to condense and pool as a relatively
much cooler liquid in the bottom of the condenser. The
condensed refrigerant is then directed out of condenser 14,
through discharge piping 26, to a metering device 28 which,
in the preferred embodiment, is a fixed orifice. That
refrigerant, in its passage through metering device 28, is
reduced in pressure and is still further cooled by the
process of expansion and is next delivered, primarily in
liquid form, through piping 30 into evaporator 16.
Refrigerant passing into and through evaporator 16
undergoes a heat exchange relationship with a medium, such
as water, which enters the evaporator through an inlet 32
and exits the evaporator through outlet 34. In the process
of cooling the medium which flows through the evaporator and
being
CA 02610421 2007-11-22
14
heated thereby, system refrigerant vaporizes and is directed,
as a relatively low pressure but relatively warm gas, through
piping 36 back to the compressor. It is there again compressed
and heated in an ongoing and repetitive process.
Referring additionally now to Figures 2 and 2a,
compressor portion 12 of chiller 10 includes a housing 39 in
which chiller drive motor 40 is disposed. Impellers 42 and 44
are disposed in volute housing 45 and are, together with rotor
46 of drive motor 40, mounted for rotation on shaft 48. Shaft
48, in turn, is mounted for rotation in first bearing package
50 and second bearing 52. It is to be noted that although the
present invention is, in its preferred embodiment, a
centrifugal chiller, chillers driven by other than centrifugal
compressors fall within its scope. In such cases, the
compressive element mounted on shaft 48.might be the rotor of a
rotary screw compressor (in which case chiller 10 would be a
screw chiller).
As will be apparent, the centrifugal chiller of the
preferred embodiment is a so-called direct drive chiller,
having the rotor 46 of its drive motor 40 mounted directly to
the shaft 48 on which the compressor's impellers are mounted.
Drive motor 40 of compressor 12 is, in the preferred
embodiment, a somewhat structurally strengthened (as will
further be explained) but essentially conventional induction
motor which is driven by a variable speed drive 54 although
other kinds of variable speed motors are contemplated as
falling within.the scope of the present invention.
By the use of drive 54, chiller 10 and its
compressor can be operated at lower speeds when the load on the
chiller system does not require the operation of the compressor
at maximum capacity and at higher speeds when there is an
.,., I . . . . . .. . . .... :.........
CA 02610421 2007-11-22
increased demand for chiller capacity. By running compressor
12 and its impellers at lower speeds when the load on the
chiller is not high or at its maximum, sufficient refrigeration
effect can be provided to cool the reduced heat load in a
5 manner which saves energy, making the chiller more economical
from a cost-to-run standpoint and making chiller operation
extremely efficient as compared to chillers which are incapable
of such load matching. Additionally, compressor 12 may employ
inlet guide vanes 55 which, in. cooperation with the controlled
10 speed of motor 40, permit very precise control of chiller
capacity so that chiller output closely and responsively
matches the system load, all while using as little energy as
possible and eliminating the need for specially designed drive
gears optimized for a specific chiller application, the need
15 for relatively more exotic and expensive variable speed drives
and/or motors or the need for an oil system to provide for the
lubrication of bearings and/or a gear train.
In the preferred embodiment, compressor 12 is a
two-stage compressor. The two-stage nomenclature indicates
that there are two distinct stages of gas compression within
the chiller's compressoY portion. Such two-stage compression
is accomplished by increasing the pressure of the system
refrigerant a first time by passing it to, through and past
first stage impeller 42 and then by communicating such once-
compressed gas to, through and past second stage impeller 44
which increases the pressure of the refrigerant a second time.
While compressor 12 is a two-stage compressor in the preferred
embodiment, it is to be understood that the present invention
is applicable not only to two-stage compressors/chillers but to
single stage and other multiple stage chiller's as well.
.. . i . .... .. . . ... _ ._.. . ._.:._-. _ _. . . __ _ .. . . . .. . . .. .
. . . .. . . . . .
CA 02610421 2007-11-22
16
Referring particularly now to Figures 2 and 2a, the
bearing arrangement associated with shaft 48 will more
thoroughly be described. Shaft 48, as earlier noted, is
supported for rotation in bearing package 50, which, in the
preferred embodiment, is comprised of first and second rolling
element bearings 50a and 50b and carries both the thrust load
and the majority of the radial load imposed through shaft 48 by
the operation of compressor 12. Bearing 52, which is an
axially floating, single angular-contact bearing having a
rolling element 53, takes up a relatively small portion of the
radial load and a portion of the thrust load. Bearing 52 is,
however, preloaded in a direction which is opposite the thrust
direction of the primary thrust load so as to minimize the net
thrust load on bearing 50b which carries the majority of the
thrust load.
Bearing package 50 is disposed approximately half-
way down the length of shaft 48 and bearings 50a and 50b are
back-to-back, preloaded, angular-contact rolling element
bearings. The rolling elements 51a and 51b of bearings 50a,
50b and the rolling element of bearing 52 will preferably be
balls rather than rollers so as to reduce the cost of the
bearings. Bearings 50a and 50b could, alternatively, be
oriented in a face-to-face manner. In any event, the races of
bearings 50a and 50b are oppositely oriented, as best
illustrated in Figure 2a, so as to take up the thrust loads
imposed through shaft 48 irrespective of the direction of that
thrust load. These bearings also carry the majority of the
radial load imposed through shaft 48.
CA 02610421 2007-11-22
17
Impellers 42 and 44 are mounted on shaft 48 on one
side of bearing package 50 while drive motor rotor 46 is
mounted on the other. Bearing package 50 is located along
shaft 48 such that the weight of the shaft and impellers on one
side of the bearing package essentially balance the weight of
the shaft and motor rotor located on the other side of that
bearing package. The impellers and the portion of shaft 48 on
which they are mounted are, however, cantilevered in the
preferred embodiment and are thus unsupported at distal end 58
of the drive shaft. The other portion of the drive shaft and
itsdistal end 60, as earlier noted, is to some extent radially
supported and carried in bearing 52. It is to be noted that
the mounting of shaft 48 in a single bearing or bearing
package, depending upon the design of such bearing or bearings,
is possible but also that different bearing arrangements and
locations are contemplated as being within the scope.of the
invention.
In the chiller of the preferred embodiment, the
bearings that comprise bearing package 50 are relatively large
bore bearings. Their location between drive motor rotor 46 and
impellers 42 and 44 permits the diameter of shaft 48 to be
large which, together with the bearing radial stiffness that
results therefrom, enhances compressor operation by elevating
critical speeds so that they are higher than the shaft will see
in operation. As such, critical speeds are avoided.
In the past, many chiller manufacturers have been
dissuaded from using rolling element bearings to support the
impeller shaft of a centrifugal compressor for rotation,
particularly where the portion of the shaft on which the
chiller's impellers are mounted is cantilevered from a support
bearing. Rather,' such manufacturers have resorted.to the use
CA 02610421 2007-11-22
18
of journal bearings which, while relatively low cost, are very
intolerant to reduced or poor lubrication (a disadvantage which
is exacerbated in a refrigerant environment) and res.ult in
increased frictional losses that are to the detriment of both
compressor and overall chiller efficiency. While the assignee
of the present invention has long successfully manufactured
centrifugal chillers having compressors the impeller shafts of
which are mounted in rolling element bearings, those rolling
element bearings have heretofore required lubrication by oil.
With the advent of so-called hybrid bearings of the
rolling element type which, as of the filing date hereof, have
only recently come to be commercially available, thought has
turned to the possibility of eliminating oil as a lubricant in
centrifugal chillers by the use of such bearings in direct
drive machines to mount the shaft on which the chiller's motor
rotor and impellers are mounted. Such hybrid bearings can be
characterized as rolling element bearings that, applicants have
found, are capable of being lubricated by refrigerant, in the
absence of oil despite manufacturer's contrary position that
oil is the preferred lubricant of such bearings with grease
being a lesser alternative.
The hybrid bearings, in the preferred embodiment of
the present invention, use non-metallic rolling elements which
are fabricated from a ceramic material. The use of a ceramic
material, such as silicon nitride, results in rolling elements
that are of on the order of 60i less dense, have a modulus of
elasticity up to 50% higher, thermally expand only 30% as much
as steel bearings and have a coefficient of, friction on the
order of 20% of that of rolling elements fabricated from
steel.
CA 02610421 2007-11-22
19
Because of the reduced density of ceramic rolling elements, the bearings in
which they are
used are subject to significantly reduced centrifugal force. The higher
modulus of elasticity reduces
friction in such bearings and makes such bearings stiffer, which reduces
distortion and friction.
Reduced distortion in these bearings increases, in turn, critical speeds in
the machines in which
they are employed. Reduced thermal expansion minimizes bearing preload
variation and likewise
reduces friction and increases bearing life. This is significant in
refrigeration chiller applications
where bearings are exposed to widely varying temperatures. While the races in
such ceramic
rolling elements run are, in the preferred embodiment, fabricated from steel,
making such bearings
"hybrid" bearings, they could likewise be fabricated from a ceramic material.
Applicants have found that the running of such ceramic rolling elements on and
within steel
races results in the creation of a mirror-like finish on the surfaces of the
races due to the hardness
and smoothness of the ceramic rolling elements that run on them. Applicants
have also found that
given this characteristic of such bearings, only a relatively very thin
elastohydrodynamic film is
required to provide adequate lubrication for such bearings.
In that regard, applicants have found that by providing a refrigerant, which
comprises the
working fluid of a centrifugal chiller, primarily and preferably in the liquid
state and at appropriate
times and in appropriate quantities to hybrid bearings, such bearings are
provided adequate
lubrication, are adequately cooled and can function across the operating
envelope of a chiller in the
absence of oil as a lubricant. That possibility does not exist with
conventional
CA 02610421 2007-11-22
bearing technology, where both the rolling elements and races
in which they run are fabricated from steel, for the reason
that the characteristics of refrigerant are not such as to
provide a sufficiently thick film between such conventional
5 rolling elements and races for lubrication purposes.
In the present invention, by the use of hybrid
bearings and liquid refrigerant to lubricate them, a thin but
sufficiently thick elastohydrodynamic film between the ceramic
rolling elements and the races in which they run is created
10 which has been found to be sufficient for bearing lubrication
purposes. With the hybrid bearings used in the present
invention, not only is the film created by system refrigerant
sufficient for lubrication purposes, it has been found that
even if the ceramic rolling elements do momentarily make
15 contact across the refrigerant film with the steel races on
which they run, the rolling elements and races continue to
function and are not susceptible to "welding" together (as
conventional steel bearings are prone to do) due to the
fabrication of the rolling elements and races from
20 significantly dissimilar base materials.
Applicants have also found, in developing the
centrifugal chiller of the present invention, that refrigerant
supplied to such hybrid bearings for lubrication purposes will
preferably be all or essentially all in the liquid state. The
liquid refrigerant delivered to such bearings serves two
purposes, the first being to create the thin elastohydrodynamic
film necessary to lubricate the bearing as between its ceramic
rolling elements and its steel races and the second being to
carry the heat of friction away from the bearing location. As
CA 02610421 2007-11-22
21
such, the liquid refrigerant delivered to the bearings for
lubrication purposes must be in a state such that an excessive
percentage of it does not flash to gas on contact with the
bearings which will be relatively warm in operation.
Applicants have therefore established a design
parameter, with respect to the chiller system of their
invention, to deliver liquid refrigerant at a sufficient rate
of flow to the bearing locations such that the amount of
refrigerant discharged from those locations in the liquid
state, after its use in the bearing lubrication process,
comprises an amount equal'to 80% of the liquid refrigerant
delivered to those locations. By allowing for up to an
approximately 20t rate of refrigerant flashing at the location
of the bearings under fringe chiller operating conditions, it
has been found that an adequate amount of liquid refrigerant
will, under all foreseeable chiller operating conditions, be
available for bearing lubrication and heat removal purposes.
- That rate of flashing, while not necessarily an upper limit, is
one with which applicants are comfortable at this stage of
development.
Despite the many advantages associated with the
elimination of the need for oil in centrifugal chiller systems,
an anomaly associated with the use of refrigerant to lubricate
the hybrid bearings in such systems has, however, been
discovered which creates a difficulty where none existed in
oil-based lubrication systems. In that regard, when oil is
used as a lubricant in a chiller system, a portion of the oil
adheres to and is maintained on the bearing surfaces as a film
for a relatively long period of time after the chiller and its
active oil delivery system is shutdown. As such, when oil is
used as a bearing lubricant, at least some of it will remain on
~ . ... _. _.. _ __...._.._.._... .. .___ .. ....... . . ... ..... .. .. ....
. . . . . . . . . .. ... . ._.. .. . . .... ... . . . . .
CA 02610421 2007-11-22
22
the bearing surfaces to provide for initial bearing lubrication
when the chiller next starts up. Such residual oil can, to at
least some extent, be relied upon to lubricate the bearings
until the chiller's oil delivery system comes to actively
provide oil to the bearing locations.
When refrigerant is used as a bearing lubricant,
little or no residual refrigerant has been found to remain on
the bearing surfaces when.the chiller system shuts down.
Rather, any refrigerant at the bearing locations when the
system is shutdow:z-drains away from or boils off of the bearing
surfaces leaving an essentially dry bearing. As such,
lubrication of the bearings in a centrifugal chiller employing
hybrid bearings lubricated exclusively by refrigerant presents
unique difficulties and challenges both at chiller start-up and
subsequent to chiller shutdown. Those problems have been
successfully addressed by the chiller lubrication system
illustrated schematically in Figure 3 which ensures the
delivery of liquid refrigerant to bearing package 50 and
bearing 52 at.compressor start-up, during normal chiller
operation and for the relatively lengthy period of time after
the chiller shuts down during which shaft 48 coasts to a stop.
Referring additionally now to Figure 3, lubrication
of bearing package 50 and bearing 52 at chiller start-up is
accomplished by providing a source of liquid refrigerant from a
location within the chiller in which liquid refrigerant resides
while the chiller is shutdown. In that regard, when a chiller
start-up signal is received, liquid refrigerant pump 62 pumps
liquid refrigerant from refrigerant sump 64. Pump 62 is
capable of pumping saturated liquid refrigerant without causing
a significant amount of 'the liquid refrigerant to flash to gas
as a result of the pumping process. Sump 64, as will
CA 02610421 2007-11-22
23
subsequently be described, is in selective flow communication,
through line 66, with system evaporator 16. Disposed in line
66 is a fill valve 68 which is open when the chiller is
shutdown and, optionally, a screen 70 for removing any
impurities/debris that might otherwise make its way into sump
64 from the evaporator.
When a chiller shuts down, the internal temperature
and pressure conditions within a chiller are such that the
refrigerant therein will migrate tc the evaporator as
temperatures and pressures within zze chiller system equalize.
Further, because the evaporator is the coldest portion of the
chiller at the time the chiller shuts down, not only will
refrigerant migrate to that location, it will condense there to
liquid form. Therefore, when the chiller next starts up, at
least the majority of the refrigerant in the chiller system can
be expected to reside in the evaporator in the liquid state.
Refrigerant sump 64 is positioned on chiller 10
such that when fill valve 68 is open, liquid refrigerant pooled
in evaporator 16 will drain to and fill refrigerant sump 64.
When the chiller is called upon to start-up, fill valve 68 is
.closed which isolates refrigerant sump 64 from the evaporator.
Absent the closure of valve 68 at this time, pump 62, which
goes into operation when the chiller start-up sequence
commences, would cavitate as the liquid refrigerant in the
evaporator boils to gas due to the pressure drop that occurs
quickly in the evaporator as the chiller starts up. It will be
appreciated that sump 64, while a discrete volume, need not be
a discrete structure but could be incorporated within another
of the many housings/shells (including condenser 14 and
evaporator 16) of which =chiller 10 is comprised.
CA 02610421 2007-11-22
24
Refrigerant pump 62, the motor 63 of which resides
within refrigerant sump 64, pumps liquid refrigerant from sump
64 through refrigerant line 72 to a liquid refrigerant
reservoir 74 which is preferably located above the chiller's
compressor section to facilitate delivery, with the assistance
of gravity, of liquid refrigerant thereoutof to bearing
locations. Sump 64 is sized to ensure that an adequate supply
of liquid refrigerant will be available for bearing lubrication
purposes during chiller start-up. Reservoir 74, as will
further be described, is the source location from which
refrigerant is delivered to bearing.package 50 and bearing 52
for lubrication purposes and is a volume, like sump 64, that is
discrete from condenser 14 and evaporator 16.
It is to be noted that pump 62 need only elevate
the pressure of the liquid refrigerant it pumps a few PSI, so
as to overcome the head against which it is pumping and the
resistance of filter 78, if one is disposed in line 72, to
' ensure that liquid refrigerant is available for bearing
lubrication purposes under all chiller operating conditions and
circumstances. Contrarily, where hydrostatic bearings are
employed, extremely high pressure "lubricant" must be made
available to bearing surfaces under certain conditions such as
at compressor start-up.
It is also to be noted that one problem associated
with pumping saturated liquid refrigerant is maintaining the
refrigerant in the liquid state within the pump. Any pressure
depression in the liquid refrigerant within the pump causes
some flashing which makes the liquid refrigerant difficult or
impossible to pump. Even with the best pump design, this
necessitates that some positive suction head be provided above
the pump inlet. Therefore, the inlet 65 to the housing 67 in
CA 02610421 2007-11-22
which pump impeller 69 is disposed must be below the liquid level of the
liquid source. In the
embodiment of Figure 3, inlet 65 of impeller housing 67 is physically below
the bottom of
condenser 14 and is, additionally, below the level of the liquid refrigerant
that will be found in
sump 64 when the chiller starts up.
5 Disposed within the line 72 is a check valve 80 which prevents backflow out
of reservoir
74 into line 72. As will further be described, pump 62 also pumps liquid
refrigerant through the
line 72 to compressor drive motor housing 39 while the chiller is in
operation. Such refrigerant is
there brought into heat exchange contact with motor 40 in order to cool it.
Liquid refrigerant pumped to reservoir 74 is metered out of reservoir 74 to
both bearing
10 package 50 and bearing 52 through metering devices 82 and 84 respectively.
Shortly after
energization of pump 62, compressor motor 40 is started and shaft 48 begins to
rotate with its
bearings being fed liquid refrigerant as a lubricant which is sourced during
the start-up period from
sump 64.
Once chiller 10 is in operation, condenser 14 becomes the source of the liquid
refrigerant
15 for bearing lubrication purposes. In that regard, once compressor 12 begins
to deliver compressed
refrigerant gas to condenser 14, the process of condensing it to the liquid
state actively commences
within the condenser. Such condensed liquid refrigerant pools at the bottom of
the condensed
liquid refrigerant pools at the bottom of the condenser and is directed
thereoutof through piping 26
to metering device 28.
20 In addition to being in flow communication with refrigerant sump 64 via
line 56, impeller
housing 65 of refrigerant pump 62, through which refrigerant is pumped into
line 72, is in open
flow communication through line 88 with the
CA 02610421 2007-11-22
26
lower portion of condenser 14. Therefore, once chiller 10
starts up and liquid refrigerant comes to be produced in
sufficient quantity condenser 14, refrigerant pump 62 commences
pumping liquid refrigerant out of condenser 14 through line 88.
A constant flow of liquid refrigerant to reservoir 74 for
bearing lubrication purposes and to compressor drive motor 40
for motor cooling purposes is thereby provided during chiller
operation with condenser 14 being the source of the liquid
refrigerant. Like sump 64, it is contemplated that reservoir
74 can be structurally incorporated into one or another of the
housing/shells that comprise chiller 10 and that it need not be
a stand alone structure although it is,.once again, a defined
volume which is discrete from condenser 14 and evaporator 16 in
the sense that it is capable of being isolated under certain
operational circumstances, with respect to flow and/or
pressure, from them.
With respect to compressor drive motor cooling,
compressor drive motor 40, in the chiller of the preferred
embodiment, is cooled by the delivery of liquid refrigerant
into direct or indirect contact with motor 40. As will be
appreciated, the source of liquid refrigerant for motor cooling
purposes is the same as the source of liquid refrigerant for
bearing lubrication purposes.
In that regard, liquid refrigerant line 90, in ~
which valve 92 is disposed, branches off from line 72 in the
embodiment of Figure 3 and liquid refrigerant is delivered
therethrough into the interior of the drive motor housing 39
where it cools drive motor 40. Valve 92 is bypassed by line
94. In this embodiment, a first flow metering device 96 is
disposed in line 90 upstream of the location at which bypass
1 . . .. ... ...... ............I.... . .. .... ...
CA 02610421 2007-11-22
27
line 94 rejoins line 90 and -a second metering device 97 is
disposed in bypass line 94. The amount of liquid permitted to
flow through device 97 is considerably less than the amount
permitted to flow through metering device 96.
Valve 92 is open during chiller operation and
provides liquid refrigerant to compressor 12 through both
metering devices 96 and 97 in a predetermined quantity which is
sufficient to cool the compressor drive motor. However, during
the chiller start-up sequence, during the chiller coast-down
period and while the chiller is shutdown, valve 92 will be
closed. As a result, liquid refrigerant flow out of line 72
into and through branch line 90 for motor cooling purposes is
significantly reduced during the chiller start-up and coast-
down time periods since such flow will only be through metering
device 97. That, in turn, helps to ensure that adequate liquid
refrigerant is available for bearing lubrication purposes
during those periods which are, as it turns out, periods during
which the need for compressor drive motor cooling is reduced.
Also, there are times when the chiller operates at
on the order of 15% or less capacity. In such instances the
condenser may not produce the quantity of liquid refrigerant
necessary to provide for both sufficient liquid refrigerant
flow to the bearings and unthrottled flow to the drive motor
for motor cooling purposes. At such times, however, motor
cooling requirements are reduced and valve 92 can similarly be
closed to ensure that adequate liquid refrigerant is available
for bearing lubrication under such light load conditions.
It is to, be noted that liquid refrigerant delivered
to the compressor's bearings will, in the preferred embodiment,
drain from the bearings, subsequent to being used for
lubrication purposes, into the interior of motor housing 39 and
CA 02610421 2007-11-22
28
will drain thereoutof, together with the refrigerant used for
motor cooling purposes, through a line 98 to condenser 14.
Return of this refrigerant to the condenser is made possible by
the use of pump 62 which, in operation, increases the pressure
of the refrigerant used for bearing lubrication and motor
cooling purposes to a pressure higher than condenser pressure
irrespective of variations in condenser pressure while the
chiller is operating. By returning such "used" refrigerant,
which has been heated in the motor cooling process and in the
process of removing heat from the bearing locations, to the
condenser, the motor and bearing heat is carried out of the
condenser and chiller by transfer to the cooling medium that
flows through the condenser. As a result, the parasitic effect
of this heat on the overall efficiency of the chiller is
eliminated. In typical refrigeration systems, refrigerant used
to cool the compressor drive motor is returned by the use of
differential pressure to the evaporator, which is at
significantly lower pressure than the condenser. In such
systems, the delivery of such additional heat to the evaporator
acts to reduce chiller efficiency and/or results in the need to
provide additional heat transfer surface area within the
evaporator to provide sufficient for both cooling the load on
the chiller system and cooling the compressor drive motor which
is a significant source of heat.
When chiller 10 is called upon to shut down,
compressor motor 40 is de-energized. That, in turn, removes
the driving force that causes shaft 48 of compressor 12 to
rotate. However, because of the large mass of shaft 48 and the
components mounted on it, the relatively very low friction of
hybrid bearings and the high speed at which all of these
components are rotating while in operation, shaft 48 continues
CA 02610421 2007-11-22
29
to rotate for a relatively long period of time, measured on the
order of several or more minutes, after the compressor drive
motor is de-energized. During that coast-down period, liquid
refrigerant must be provided to bearing package 50 and bearing
52 to provide for their lubrication until such time as shaft 48
coasts to a stop.
It will be remembered that so long as compressor 12
operates, the source of liquid refrigerant for bearing
lubrication purposes will 'be the chiller condenser. Upon
chiller shutdown, however, the supply of refrigerant gas to the
condenser stops, pressure in the condenser drops rapidly and
the liquid refrigerant in the condenser starts to boil. As
such, very soon after chiller 10 is shutdown, the then-existing
source of liquid refrigerant for bearing lubrication purposes
comes to be unavailable as it flashes to gaseous form and
another source for liquid refrigerant must be turned to for
bearing lubrication purposes as shaft 48 coasts to a stop.
As an aside, it will be noted that refrigerant sump
64 is vented through line 104 to condenser 14 so that upon
compressor shutdown, not only will the refrigerant in condenser
14 commence to boil to the gaseous state but any liquid
refrigerant in refrigerant sump 64 will do likewise.
Refrigerant pump 62 may be permitted to continue to run for a
short period of time, on the vrder of 20 seconds or so, after
compressor drive motor 40 is de-energized because sufficient
liquid refrigerant will remain in condenser 14 and refrigerant
sump 64 to permit pump 62 to continue pumping liquid
refrigerant for that period of time. After that period of time
pump 62 would commence cavitating as a result of the flashing
CA 02610421 2007-11-22
of the liquid refrigerant to the gaseous state. Once again,
however, the need for liquid refrigerant for bearing
lubrication purposes extends to a matter of several minutes or
more as shaft 48 coasts to a stop, not a matter of seconds.
5 As was earlier noted, a check valve 80 is disposed
in line 72 which prevents flow out of reservoir 74 back through
line 72. When refrigerant pump 62 is de-energized shortly
after chiller shutdown, the pressure in line 72 upstream of
check valve 80 drops and the pressure in reservoir 74 causes
10 check valve 80 to seat. A sufficient amount of pressurized
liquid refrigerant is thus trapped within reservoir 74 between
check valve 80 and metering devices 82 and 84 to ensure that
bearing package 50 and bearing 52 are provided adequate liquid
refrigerant, by gravity feed and residual pressure, during the
15 compressor coast-down period. Reservoir 74 is appropriately
sized for that purpose. It is to be noted that reservoir 74
also ensures that a supply of lubricant in the form of liquid
refrigerant is available to the compressor bearings for a
sufficient period of time should power to the chiller be
20 interrupted (even though pump 62 will not continue to operate
as it would during a normal shutdown sequence where it
continues to operate for a brief period of time subsequent to
chiller shutdown).
After chiller shutdown, whether "normal" or in
25 response to an abnormal condition such as interruption of
power, when pressure has equalized across the chiller, fill
valve 68 is again opened and refrigerant sump 64 fills with
liquid refrigerant from evaporator 16. The system is then
ready, from the bearing lubrication standpoint, to start once
30 again.
CA 02610421 2007-11-22
31
It is to be noted that each time the chiller shuts
down, it will be required to remain shut down for some
relatively small period of time, such as ten minutes, during
which refrigerant sump 64 refills with liquid refrigerant. In
most circumstances, however, once chiller 10 shuts down, it
will not normally be called upon to start-up for at least that
amount of time irrespective of the need to refill reservoir 64.
Therefore, the mandatory shutdown period for purposes of re-
filling reservoir 64 has little or no effect on chiller
operation in real terms.
It has been noted that refrigerant pump 62 is
disposed in refrigerant sump 64 and is bathed within the liquid
refrigerant found therein. Because of its location, pump 62
can likewise make use of hybrid bearings lubricated by liquid
refrigerant, eliminating a still further need for an oil-based
lubrication system found in other refrigeration chillers.
Further, because pump 62 is disposed within refrigerant sump
64, it and its motor are effectively kept cool by the liquid
refrigerant in which they are immersed.
Referring to refrigerant reservoir 74, it is to be
noted that a unique device 100, which is the subject of a co-
pending patent application U.S. Serial Number 08/924,228,
likewise assigned to the assignee of the present invention, is
used to "prove" the presence of liquid in reservoir 74. This
device protects the compressor against failure by its ability
to differentiate between the existence of liquid and gaseous
foam in a flowing fluid.
As has been mentioned, lubrication of bearing
package 50 and bearing 52 depends upon the continuous delivery
to them of liquid refrigerant in sufficient quantity. By the
use of flow proving devi.ce 100 which, if insufficient liquid
CA 02610421 2007-11-22
32
content in the fluid flow passing through reservoir 74 is
detected, causes chiller 10 to shutdown, the chiller is
protected from damage or failure for lack of proper
lubrication. The lubrication scheme of the present invention
is therefore made subject to a safeguard which protects the
chiller and its compressor against catastrophic damage should
reservoir 74, for some reason, come to contain refrigerant
which, to too great an extent, is other than in the liquid
state. As will be appreciated, device 100and the safeguarding
of chiller 10, while important in the context of the commercial
embodiment of chiller 10, is a peripheral feature with respect
to the refrigerant-based lubrication system of the present
invention.
Referring now to Figure 4, an alternate embodiment
of the present invention will be described, individual
different features of which are capable of being employed in
the Figure 3 and other embodiments of the present invention
that are found herein. In this embodiment of Figure 4,
refrigerant sump 64 of the preferred embodiment is eliminated
in circumstances/applications where bearing package 50 and
bearing 52 of compressor 12 can tolerate dry operation during
the period of time, subsequent to chiller start-up, when the
condensation process in condenser 14 is incapable of providing
liquid refrigerant of the quality and in the quantity which
becomes necessary for bearing lubrication purposes while the
chiller is in steady state/normal operation. The embodiment of
Figure 4, while less costly and less complicated than the
preferred embodiment, represents a more risky design philosophy
which is predicated on the ability of hybrid bearings to run
dry or essentially dry for some relatively small but
permissible period of time at chiller start-up.
CA 02610421 2007-11-22
33
In the Figure 4-embodiment, refrigerant pump 200 is
disposed immediately adjacent liquid weir 202 of condenser 14
and is therefore capable of moving liquid refrigerant from that
location to the bearings of the compressor as soon as such
liquid becomes available. In this embodiment, liquid
refrigerant produced in condenser 14 drains out of weir 202
into pump housing 204. Pump housing 204 is such that its motor
206 is bathed in liquid refrigerant which both cools the motor
and provides a source of lubricant for the hybrid bearings used
in pump 200 itself.
A delay in the start-up of pump 200 for a period of
time after chiller start-up until such time as liquid
refrigerant comes to be produced in condenser 14 prevents pump
200 from cavitating as it would otherwise do if it was started
coincident with chiller start-up. During the period of time
during which pump 200 remains de-energized, bearings 50 and 52
are permitted to run dry. As soon as liquid refrigerant comes
to be available in weir 202, however, pump 200 is energized and
liquid refrigerant is provided to those bearings for
lubrication purposes.
Another mechanical modification in the system of
Figure 4 which is applicable to others of the embodiments
herein is the sourcing of refrigerant for motor cooling
purposes from reservoir 74 rather than by diversion from line
72 upstream of check valve 80. In that regard, motor cooling
refrigerant is supplied to motor housing 39 from reservoir 74
through line 208. The size of reservoir 74 in this embodiment
is adjusted accordingly. Line 208 will preferably source
refrigerant from reservoir 74 at a level higher than the level
CA 02610421 2007-11-22
34
at which bearing lubricant is sourced so that should the liquid
level fall, bearing lubrication will.continue even if motor
cooling is interrupted. The motor can be protected in such
circumstances in other ways.
A further mechanical modification in the system of
Figure 4 which is applicable to others of the embodiments
herein involves the use of an economizer 106 the purpose of
which, as is well known with respect to refrigeration chillers,
is to make use of intermediate pressure refrigerant gas
existing within the system to enhance overall system
efficiency. In that regard, economizer 106 is disposed within
the chiller system so that condensed liquid refrigerant passes
from condenser 14 through a first metering device 108 into
economizer 106. Economizer 106, in the preferred embodiment,
defines two discrete volumes 110 and 112. Refrigerant flowing
through metering device 108 flows into volume 110 of economizer
106 and a portion of it flashes to gas at a first pressure.
Such gas is then directed through line 114 to the portion of
volute housing 45 (see Figure 2) in which second stage impeller
44 is housed to increase the pressure of the gas delivered to
the second stage impeller without its being acted upon by the
impeller driven compression process.
A second metering device 116 is disposed between
volumes 110 and 112 which meters refrigerant from volume 110 to
volume 112. That process lowers refrigerant pressure in the
process and causes a still further portion of the refrigerant
to flash to gas at a somewhat lower pressure than the flash gas
generated in volume 110.
CA 02610421 2007-11-22
Gas from volume 112 flows through line 118 to the
portion of volute housing 45 (see Figure 2) in which first
stage impeller 42 is housed and acts to increase the pressure
of the refrigerant gas in that location without its being acted
5 upon by the first stage impeller. By the use of an economizer,
additional efficiencies are added to the compression process
that takes place in chiller 10 and the overall efficiency of
chiller 10 is increased.
Liquid refrigerant exits volume 112 of economizer
10 106, flows through a third metering device 120 and enters
evaporator 16. In the embodiment of Figure 4, like the
embodiment of Figure 3, metering devices 108, 116 and 120 are
fixed orifices. As is shown by the routing of line 98 to the
economizer in the Figure 4 embodiment, the present invention
15 contemplates the possible return of refrigerant used for motor
cooling and/or bearing lubrication purposes to the economizer,
where one is employed, rather than to the condenser. The
condenser does, however, remain a viable return location. In
all other pertinent respects, the lubrication of the hybrid
20 bearings of compressor 12 in the Figure 4 embodiment is the
same as is accomplished in the Figure 3 embodiment, including
with respect to their lubrication after chiller shutdown as
shaft 48 coasts to a stop.
Referring now to Figure 5, still another embodiment
25 of the present invention will be described. In the embodiment
of Figure 5, refrigerant pump 62 of the embodiment of Figures 3
is dispensed with and condenser pressure is used to drive a
controlled amount of liquid refrigerant from weir 300 of
condenser 14 to the bearings 50 and 52 of compressor 12. The
30 embodiment of Figure 5, like the embodiment of Figure 4, is a
system in which the hybrid bearings of compressor 12 are
CA 02610421 2007-11-22
36
permitted to run dry after chiller start-up until such.time as
sufficient liquid refrigerant has been produced and pressure
developed in condenser 14 to drive liquid refrigerant from the
condenser to the compressor for both bearing lubrication and
motor cooling purposes.
Elimination of the pump used to pump liquid
refrigerant to the compressor bearings, the cost associated
with such a pump as well as elimination of the failure modes
associated therewith offer distinct advantages. However, with
the embodiment of Figure 5 it must '--e assured that condenser
pressure will at all times be sufficient during chiller
operation to ensure that liquid refrigerant, in adequate
quantities, is delivered to the reservoir 74 across the entire
operating envelope of the chiller and is likewise sufficiently
high to ensure that there is adequate liquid refrigerant at a
sufficiently high pressure in reservoir 74 to cause delivery of
liquid refrigerant thereoutof to the compressor bearings during
the compressor coastdown period. The availability of such
_ pressure in the condenser can be marginal under some chiller
operating conditions and/or in some chiller applications so it
will be appreciated that the lubrication system of Figure 5
represents a still more risky design philosophy than the
philosophy underlying the Figure 4 embodiment. It is to be
noted that because pump 62 is eliminated in the Figure 5
embodiment, the return of refrigerant used for motor cooling
purposes through line 98 is to the evaporator 16 rather than to
condenser 14.
CA 02610421 2007-11-22
37
Referring now to Figure 6, a still further
alternate to the Figure 3 preferred embodiment of the present
invention will be described. In the embodiment of Figure 6,
valve 68 in line 66 from evaporator 16 is dispensed with and
sump 64 is replaced by pump 400. Pump apparatus 400 is
therefore in free-flow communication with both condenser 14 and
evaporator 16.
Pump 400 is comprised of a housing 402 in which a
motor 404, comprised of a stator 406 and rotor 408, is
disposed. Stator 406 is fixedly mounted in housing 402 while
rotor 408 is mounted for rotation on a drive shaft 410. Drive
shaft 410, in turn, is mounted for rotation in ceramic bearings
412 and 414.
A first impeller 416 is mounted on one end of drive
shaft 410 while a second impeller 418 is similarly mounted on
the other end of the drive shaft. Impellers 416 and 418 are
respectively disposed in impeller housings 420 and 422 and
together, impeller 416 and housing 420 form a first pumping
mechanism 421 while impeller 418 and housing 422 form a second
pumping mechanism 423. As will be appreciated, impellers 416
and 418 are commonly driven by drive shaft 410 which, in turn,
is driven by motor 404:
Impeller housing 420 defines an inlet 425 through
which liquid refrigerant is drawn by pumping mechanism 421 from
condenser 14 through piping 88. Impeller housing 422 similarly
defines an inlet 427 through which liquid refrigerant is drawn
by pumping mechanism 423 through piping 66. Piping 66, in this
embodiment, is in flow communication with evaporator 16.
CA 02610421 2007-11-22
38
In operation, impeller 416 draws liquid refrigerant
from condenser 14, when it is available therefrom, while
impeller 418 drawsliquid refrigerant from evaporator 16 when
liquid refrigerant is available from that source location.
Liquid refrigerant pumped by impeller 416 from condenser 14 is
delivered out of impeller housing.420 into piping 424 while
liquid refrigerant pumped by impeller 418 from system
evaporator 16 is delivered out of impeller housing 422 into
piping 426.
In the embodiment of Figure 6, piping 424 and
piping 426 converge at the location of a valve 428 which is
connected to piping 72 of the preferred Figure 3 and other
alternate embodiments. Valve 428 includes a flapper element
430 which is automatically and without the need for a control
or sensors positioned in accordance with the effect.and
pressure of the respective flow streams that enter that valve
from piping 424 and piping 426. Therefore, if liquid
refrigerant is available in one source location at a first
pressure and in the other source location at a second pressure,
valve 28 will be positioned automatically and under the effect
of such pressures such that the output of the pump apparatus
will be from the one of the two source locations which is at
higher pressure.
As has been mentioned and as applies to all of the
embodiments of the present invention, where liquid refrigerant
is relied upon in a chiller for a purpose other than providing
a refrigerating or cooling effect, the need is to ensure that a
supply of liquid refrigerant is reliably available for such
other purposes under all chiller operating conditions and
circumstances. As has further been mentioned, there is
essentially no location within a chiller that can reliably be
CA 02610421 2007-11-22
39
assumed to contain liquid refrigerant that is capable of being
pumped under all such conditions and circumstances. In
general, when a chiller is shutdown or is operating at
extremely low load conditions, liquid refrigerant will reliably
be found to exist in the system evaporator. When the chiller
is operating at load, the most reliable source of liquid
refrigerant is the system condenser (liquid refrigerant in the
evaporator will be boiling and thus not in a form that is
readily pumped).
As has still further been mentioned, liquid
refrigerant pump development to date has demonstrated that the
amount of head required to permit the successful pumping of
saturated liquid refrigerant is greater as the saturation
temperature decreases. It is therefore more difficult to pump
liquid refrigerant from the relatively more cold evaporator
than from the condenser. As with the other embodiments herein,
the alternate embodiment of Figure 6 uses liquid refrigerant
sourced from the condenser for bearing lubrication and
compressor drive motor cooling purposes under the majority of
chiller operating conditions and uses liquid refrigerant
sourced from the evaporator for such purposes when liquid
refrigerant is not reliably available from the system condenser
(such as at chiller start-up) or is not in a state within the
condenser that facilitates pumping. It can be expected,
however, that under any chiller operating condition or
circumstance, liquid refrigerant that is capable of being
pumped will be available from one and sometimes both of these
source locations.
CA 02610421 2007-11-22
With respect to the Figure 6 embodiment, when pump
apparatus 400 is in operation, both of impellers 416 and 418
rotate and simultaneously attempt to draw liquid refrigerant,
if available, from their respective source locations. Because
5 of the pressure, amount and condition of the refrigerant in
their respective source locations, the refrigerant, if any,
respectively discharged into piping 424 by pumping mechanism
421 and into piping 426 by pumping mechanism 423 will, at any
given moment, most often be at different pressures in
10 accordance with the then-existing conditions in those
respective source locations.
Valve 428 is essentially a simple check valve
arrangement that channels the flow of liquid refrigerant into
piping 72 from the one of the two pumping mechanisms that
15 constitute pump apparatus 400 the output of which is at higher
pressure. That pumping mechanism will be the one which draws
refrigerant from the source location where liquid refrigerant
exists and is at higher pressure at the moment. As internal
chiller conditions change and the other source location comes
20 to contain liquid refrigerant at higher pressure, the position
of flapper element 430 will change and the source of liquid
refrigerant will shift in accordance with such changed
conditions. It will be noted that the assured supply of liquid
refrigerant to piping 72 in the embodiment of Figure 6 is
25 accomplished very simply, in accordance with the laws of
physics, and without the need for sensors or proactive control
of any device to select the appropriate source location.
Rather than using flapper type check valve 428, a
first check valve 440, shown in phantom in Figure 6, could be
30 disposed in line 424 and a similar second check valve 442,
likewise shown in phantom in Figure 6, could be disposed in
CA 02610421 2007-11-22
41
piping 426. Like the aforementioned arrangement in which valve
428 is employed, the purpose of individual check valves 440 and
442 is to permit the flow of liquid refrigerant out of the one
of piping 424 and piping 426 which is the source of higher
pressure liquid refrigerant while blocking the flow of such
higher pressure liquid refrigerant into the other of pipes 424
and 426 and to the impeller which feeds it.
It is to be noted that although the embodiment of
Figure 6 employs two impellers, the costs associated with the
use of a second impeller are minimal. With respect to the
lubrication of ceramic bearings 412 and 414 and the cooling of
pump motor 404, bearing 412, which is adjacent pumping
mechanism 421 that draws liquid refrigerant from condenser 14
(a typically higher pressure location), will preferably be a
shielded bearing that permits the metered leakage of liquid
refrigerant out of impeller housing 420 leakage through it and
into the interior 432 of motor housing 402. Bearing 414,
adjacent pumping mechanism 423, may or may not be shielded.
During normal chiller operation, a metered amount
of liquid refrigerant will pass through shielded bearing 412
from the relatively high pressure condenser location, will
enter the interior 432 of the motor housing. In the process,
it will both lubricate bearings 412 and 414 and cool motor 404.
Under the more infrequent circumstance where evaporator 16 is
the higher pressure source for liquid refrigerant, such
refrigerant will flow through bearing 414 into the interior of
432 and will both lubricate the pump bearings and cool the
motor. The interior of housing 402 in the embodiment of Figure
6 is vented through line 434 to evaporator 16 although the best
vent location has not, as of this writing, been determined.
Bearing 412 must be shielded and refrigerant flow therethrough
CA 02610421 2007-11-22
42
metered or that location would constitute a high-to-low side
leak within the chiller which would be detrimental to chiller
operation and efficiency. That same concern does not exist
when "atypical" systems conditions cause the evaporator to be
the source of higher pressure liquid refrigerant.
It is also to be noted that the pump impeller that
is not active at any one time to pump liquid refrigerant into
line 72 against the pumping action'of the other impeller may
experience refrigerant churning in its attempts to pump a
mixture of gas and liquid refrigerant from its source location.
Such churning should not be problematic since any heat
generated thereby will cause the churned liquid portion of the
refrigerant to flash to gas which, in turn, will provide
cooling in the location of that impeller.
It is still further to be noted that the present
invention also contemplates the use of pump apparatus that is
constituted of two discrete motor/pump combinations,
appropriately piped together. The use of two motors to drive
two pump mechanisms is, of course, less attractive for many
reasons than the use of a single motor to drive two pump
mechanisms.
Finally, and as will be apparent, the pumping
arrangement of Figure 6, while specifically designed in
contemplation of a chiller system using ceramic bearings in
which liquid refrigerant is used to lubricate such bearings, is
applicable for motor cooling purposes in conventional chiller's
where oil is used for compressor bearing lubrication.
CA 02610421 2007-11-22
43
Referring back now to Figure 2, as applicants have
noted, drive motor 40 is, in the preferred embodiment, an
induction motor driven by a variable speed drive. Heretofore,
typical induction motors, which bring with them advantages of
low cost and reliability, have generally not been driven by
variable speed drives in chiller applications at speeds greater
than 3600 RPM.
In chillers of the gear drive design, while the
induction motor which drives the gear train is typically driven
to a maximum speed on the order of 3600 RPM, the impeller of
the machine and the shaft on which the impeller is mounted are
driven at relatively very much higher speed by the speed
increasing effect of the gear train. Such machines, which are
most typically single stage machines, are run over a range of
speeds in order to modulate the capacity of the chiller over a
design capacity range. Relatively very high speeds (on the
order of 15,000 RPM) are often required of such single stage
machines in order for such chillers to deliver their maximum
capacity and, once again, such machines have the disadvantage
of requiring the existence of an oil-based lubrication system.
Applicant's have prospectively determined that
proven, less expensive induction motors can be structurally
strengthened with respect to their construction, so as to
permit such motors to be driven at speeds which are higher than
the 3,600 RPM they are typically driven at in current direct
and gear drive chillers but which are relatively far lower than
the speeds required of high speed gear drive machines to
deliver the same and maximum capacity. In that regard,
applicant's have found that where the compressor's drive motor
is a structurally strengthened induction motor that is reduced
in size but driven at speeds higher than 3600 RPM and where the
CA 02610421 2007-11-22
44
chiller is a multiple stage direct drive chiller, a capacity
modulated chiller is capable of being produced which can
deliver a capacity equal to that of a gear drive machine under
a circumstance in which the impellers are driven at a speed
which is only on the order of one-half of the speeds required
of single stage gear drive chillers in delivering such
capacity. Such a direct drive chiller is capable of delivering
its capacity by the use of an induction motor driven by
conventional variable speed drive technology and without resort
to exotic or expensive emerging motor and/or motor drive
technology, and, by the use of hybrid bearings, offers the
still further advantages of a chiller in which the need for an
oil-based lubrication system is eliminated entirely.
One other aspect of the present invention related
to the use of a variable speed compressor drive motor in
association with the oil-free liquid chiller disclosed herein
is the opportunity to cool variable speed drive 54 with liquid
refrigerant as opposed to air which is the more typical case.
As is illustrated in Figure 6, line 500, shown in phantom,
branches off of line 90 through which liquid refrigerant is
delivered into heat exchange contact with chiller drive motor
40. The liquid refrigerant flowing into drive 54 cools the
heat generating components therein and will preferably be
returned to condenser 14 through line 502. Line 500, through
which liquid refrigerant is sourced for purposes of cooling
drive 54 could alternatively branch directly off of line 72 or
could be fed out of reservoir 74. Alternatively, liquid
refrigerant could sequentially be caused to flow in a series
CA 02610421 2007-11-22
rather than parallel fashion to the compressor drive motor and
controller 54. It will be appreciated that this concept is not
limited in application to the embodiment of Figure 6 but could
likewise be applied to the other embodiments described herein.
5 It will be appreciated that while the present
invention has been described in terms of a preferred and
alternate embodiments, other modifications and features
pertaining thereto fall within the scope of the invention. As
such, the present invention is not limited to such embodiments.
10 but encompasses such still other embodiments and modifications
that will be apparent to those skilled in the art given the
teaching hereof.
What is claimed is:
