U.S. patent application number 11/448663 was filed with the patent office on 2006-12-14 for refrigerating machine having intermediate-pressure receiver.
This patent application is currently assigned to SANYO ELECTRIC CO., LTD.. Invention is credited to Ichiro Kamimura, Hiroshi Mukaiyama, Masahisa Otake, Koji Sato.
Application Number | 20060277932 11/448663 |
Document ID | / |
Family ID | 36997685 |
Filed Date | 2006-12-14 |
United States Patent
Application |
20060277932 |
Kind Code |
A1 |
Otake; Masahisa ; et
al. |
December 14, 2006 |
Refrigerating machine having intermediate-pressure receiver
Abstract
A refrigerating machine including a two-stage compressor, a
high-pressure gas cooler, a first expansion valve, an
intermediate-pressure receiver, a second expansion valve, and an
evaporator is further equipped with an intermediate-pressure
refrigerant bypass circuit for bypassing gas refrigerant in the
intermediate-pressure receiver to an intermediate-pressure portion
of the two-stage compressor, a back flow preventing device provided
to the intermediate-pressure refrigerant bypass circuit for
preventing back flow of refrigerant from the two-stage compressor
to the intermediate-pressure receiver, and a refrigerant-pressure
control unit for controlling the pressure of the refrigerant in the
intermediate-pressure receiver on the basis of the difference
between a specific enthalpy of refrigerant discharged from the
high-pressure gas cooler and a predetermined reference
enthalpy.
Inventors: |
Otake; Masahisa; (Gunma,
JP) ; Sato; Koji; (Gunma, JP) ; Kamimura;
Ichiro; (Gunma, JP) ; Mukaiyama; Hiroshi;
(Gunma, JP) |
Correspondence
Address: |
MCDERMOTT WILL & EMERY LLP
600 13TH STREET, N.W.
WASHINGTON
DC
20005-3096
US
|
Assignee: |
SANYO ELECTRIC CO., LTD.
|
Family ID: |
36997685 |
Appl. No.: |
11/448663 |
Filed: |
June 8, 2006 |
Current U.S.
Class: |
62/196.1 ;
62/510 |
Current CPC
Class: |
F25B 1/10 20130101; F25B
9/008 20130101; F25B 2700/21152 20130101; F25B 2400/13 20130101;
F25B 2341/063 20130101; F25B 2309/061 20130101; F25B 2700/2102
20130101; F25B 2700/2109 20130101; F25B 2400/23 20130101; F25B
13/00 20130101; F25B 2600/2513 20130101; F25B 2700/21151 20130101;
F25B 41/39 20210101; F25B 2500/28 20130101; F25B 2313/02791
20130101 |
Class at
Publication: |
062/196.1 ;
062/510 |
International
Class: |
F25B 41/00 20060101
F25B041/00; F25B 1/10 20060101 F25B001/10 |
Foreign Application Data
Date |
Code |
Application Number |
Jun 8, 2005 |
JP |
P2005-168140 |
Claims
1. A refrigerating machine comprising a two-stage compressor, a
high-pressure gas cooler for cooling high-pressure gas refrigerant
discharged from the two-stage compressor, a first throttling device
for expanding the gas refrigerant from the high-pressure gas
cooler, an intermediate-pressure receiver for adjusting a
refrigerant circulating amount, a second throttling device for
expanding the refrigerant from the intermediate-pressure receiver
and an evaporator that are successively connected to one another to
form a closed refrigerant circuit, further comprising: an
intermediate-pressure refrigerant bypass circuit for bypassing gas
refrigerant in the intermediate-pressure receiver to an
intermediate-pressure portion of the two-stage compressor; a back
flow preventing device that is provided to the
intermediate-pressure refrigerant bypass circuit and prevents back
flow of refrigerant from the two-stage compressor to the
intermediate-pressure receiver; and a refrigerant-pressure control
unit for controlling the pressure of the refrigerant in the
intermediate-pressure receiver on the basis of the difference
between a specific enthalpy of refrigerant discharged from the
high-pressure gas cooler and a predetermined reference
enthalpy.
2. The refrigerating machine according to claim 1, wherein the
predetermined reference enthalpy value is set to the enthalpy of
saturated liquid corresponding to the pressure of the
intermediate-pressure portion of the two-stage compressor.
3. The refrigerating machine according to claim 1, wherein the
first throttling device comprises a first expansion valve and the
second throttling device comprises a second expansion valve.
4. The refrigerating machine according to claim 3, wherein the
refrigerant-pressure control unit controls the valve opening
degrees of the first and second expansion values to thereby adjust
the pressure of the refrigerant in the intermediate-pressure
receiver.
5. The refrigerating machine according to claim 1, wherein the
refrigerant-pressure control unit controls at least one of the
first throttling device and the second throttling device so that
the pressure of the refrigerant in the intermediate-pressure
receiver is lower than the pressure of the refrigerant in the
intermediate-pressure portion of the two-stage compressor when the
specific enthalpy of the refrigerant discharged from the
high-pressure gas cooler is not larger than the predetermined
reference enthalpy, and also so that the pressure of the
refrigerant in the intermediate-pressure receiver is lower than the
pressure of saturated liquid whose enthalpy is substantially equal
to the specific enthalpy of the refrigerant discharged from the
high-pressure gas cooler and also higher than the pressure of the
refrigerant in the intermediate-pressure portion of the two-stage
compressor when the specific enthalpy of the refrigerant discharged
from the high-pressure gas cooler is larger than the predetermined
reference enthalpy.
6. The refrigerating machine according to claim 5, wherein the
reference predetermined enthalpy value is set to the enthalpy of
saturated liquid corresponding to the pressure of the
intermediate-pressure portion of the two-stage compressor.
7. A refrigerating machine comprising: an outdoor unit including a
compressor having an intermediate pressure portion into which
intermediate-pressure refrigerant having intermediate pressure
between refrigerant pressure at a suction port of the compressor
and refrigerant pressure at a discharge port of the compressor can
be introduced, and an outdoor heat exchanger serving as a
heat-source side heat exchanger, a plurality of indoor units each
including an indoor heat exchanger serving as a using side heat
exchanger, the plural indoor units carrying out one of cooling
operation or heating operation at the same time or carrying out a
mixing operation including cooling operation and heating operation
at the same time; an inter-unit pipe for connecting the outdoor
unit and each of the indoor units to each other, the inter-unit
pipe comprising a high-pressure pipe connected to the refrigerant
discharge pipe, a low-pressure pipe connected to the refrigerant
suction pipe and an intermediate pressure pipe connected to the
other end of the outdoor heat exchanger; an intermediate-pressure
receiver that is disposed between the heat-source side heat
exchanger and the using side heat exchanger and separates
gas/liquid mixture refrigerant discharged from any one of the
heat-source heat exchanger and the using side heat-exchanger into
gas refrigerant and liquid refrigerant and supplying the gas
refrigerant to the intermediate-pressure portion of the compressor;
a back flow preventing device that is provided between the
intermediate-pressure receiver and the intermediate-pressure
portion of the compressor and prevents back flow of the gas
refrigerant from the compressor to the intermediate-pressure
receiver; and a refrigerant-pressure control unit for controlling
the pressure of the refrigerant in the intermediate-pressure
receiver on the basis of the difference between a predetermined
reference enthalpy and a specific enthalpy of refrigerant
discharged from any one of the heat-side heat exchanger and the
using side heat exchanger.
8. The refrigerating machine according to claim 7, wherein the
refrigerant-pressure control unit comprises a first expansion valve
disposed between the heat-source side heat exchanger and the
intermediate-pressure receiver and a second expansion valve
disposed between the intermediate-pressure receiver and the using
side heat exchanger, and a controller for controlling at least one
of the valve opening degrees of the first and second expansion
valves on the basis of the difference between the predetermined
reference enthalpy and the specific enthalpy of the refrigerant
discharged from any one of the heat-side heat exchanger and the
using side heat exchanger.
9. The refrigerating machine according to claim 8, wherein when any
one of the heat-side heat exchanger and the using side heat
exchanger operates as a heat-radiation side heat exchanger, the
controller controls at least one of the valve opening degrees of
the first and second expansion valves so that the pressure of the
refrigerant in the intermediate-pressure receiver is lower than the
pressure of the refrigerant in the intermediate-pressure portion of
the compressor when the specific enthalpy of the refrigerant
discharged from the heat-radiation side heat exchanger is not
larger than the predetermined reference enthalpy, and also so that
the pressure of the refrigerant in the intermediate-pressure
receiver is lower than the pressure of saturated liquid whose
enthalpy is substantially equal to the specific enthalpy of the
refrigerant discharged from the heat-radiation side heat exchanger
and also higher than the pressure of the refrigerant in the
intermediate-pressure portion of the compressor when the specific
enthalpy of the refrigerant discharged from the heat-radiation side
heat exchanger is larger than the predetermined reference
enthalpy.
10. The refrigerating machine according to claim 9, wherein the
predetermined reference enthalpy value is set to the enthalpy of
saturated liquid corresponding to the pressure of the
intermediate-pressure portion of the compressor.
Description
BACKGROUND OF THE INVENTION
[0001] 1. Field of the Invention
[0002] The present invention relates to a refrigerating machine
having an intermediate-pressure receiver, gas refrigerant in the
intermediate pressure receiver being supplied to an intermediate
pressure portion of a two-stage compressor.
[0003] 2. Description of the Related Art
[0004] There is known a refrigerating machine in which a two-stage
compressor, a high pressure gas cooler for cooling high-pressure
gas refrigerant, a first throttling device, an intermediate
pressure receiver for adjusting a refrigerant circulating amount, a
second throttling device and an evaporator that are generally
successively connected to one another to thereby form a closed
circuit, an intermediate pressure refrigerant bypass circuit for
bypassing intermediate-pressure refrigerant vapor in the
intermediate pressure receiver to the intermediate pressure portion
of the compressor is equipped, and the high pressure portion is
driven under a supercritical state under normal operation (see
JP-A-2003-106693). In this type of refrigerating machine, gas
refrigerant separated in the intermediate pressure receiver is
introduced into the intermediate pressure portion of the two-stage
compressor while kept under gas state, and thus the refrigerating
machine is set to a so-called two-stage expansion economizer cycle,
so that the refrigerant flow in the evaporator is reduced and the
compression driving force of the first-stage compressor is reduced
and the pressure loss in the evaporator is reduced. Therefore, the
performance of the refrigeration cycle can be enhanced.
[0005] However, in the conventional two-stage expansion economizer
cycle, for example when only liquid-phase refrigerant exists in the
intermediate pressure receiver due to an external temperature, a
load condition or the like, a part of the liquid-phase refrigerant
to be introduced into the evaporator is introduced into the
intermediate pressure portion of the two-stage compressor, so that
the compression efficiency is reduced and also the compressor is
damaged or the like by liquid back.
SUMMARY OF THE INVENTION
[0006] The present invention has been implemented in view of the
foregoing situation, and an object of the present invention is to
provide a refrigerating machine that can keep the optimum
performance in accordance with an external temperature, a load
condition or the like.
[0007] In order to attain the above object, according to an aspect
of the present invention, there is provided a refrigerating machine
comprising a two-stage compressor, a high-pressure gas cooler for
cooling high-pressure gas refrigerant discharged from the two-stage
compressor, a first throttling device for expanding the gas
refrigerant from the high-pressure gas cooler, an
intermediate-pressure receiver for adjusting a refrigerant
circulating amount, a second throttling device for expanding the
refrigerant from the intermediate-pressure receiver and an
evaporator that are successively connected to one another to form a
closed refrigerant circuit, which is further equipped with an
intermediate-pressure refrigerant bypass circuit for bypassing gas
refrigerant in the intermediate-pressure receiver to an
intermediate-pressure portion of the two-stage compressor, a back
flow preventing device that is provided to the
intermediate-pressure refrigerant bypass circuit and prevents back
flow of refrigerant from the two-stage compressor to the
intermediate-pressure receiver, and a refrigerant-pressure control
unit for controlling the pressure of the refrigerant in the
intermediate-pressure receiver on the basis of the difference
between a specific enthalpy of refrigerant discharged from the
high-pressure gas cooler and a predetermined reference
enthalpy.
[0008] In the above-described refrigerating machine, under normal
operation, a high pressure portion of the refrigerating machine
operates under a supercritical state.
[0009] According to another aspect of the present invention, there
is provided a refrigerating machine comprises: an outdoor unit
including a compressor having an intermediate pressure portion into
which intermediate-pressure refrigerant having intermediate
pressure between refrigerant pressure at a suction port of the
compressor and refrigerant pressure at a discharge port of the
compressor can be introduced, and an outdoor heat exchanger serving
as a heat-source side heat exchanger; a plurality of indoor units
each including an indoor heat exchanger serving as a using side
heat exchanger, the plural indoor units carrying out one of cooling
operation or heating operation at the same time or carrying out a
mixing operation including cooling operation and heating operation
at the same time; an inter-unit pipe for connecting the outdoor
unit and each of the indoor units to each other, the inter-unit
pipe comprising a high-pressure pipe connected to the refrigerant
discharge pipe, a low-pressure pipe connected to the refrigerant
suction pipe and an intermediate pressure pipe connected to the
other end of the outdoor heat exchanger; an intermediate-pressure
receiver that is disposed between the heat-source side heat
exchanger and the using side heat exchanger and separates
gas/liquid mixture refrigerant discharged from any one of the
heat-source heat exchanger and the using side heat-exchanger into
gas refrigerant and liquid refrigerant and supplying the gas
refrigerant to the intermediate-pressure portion of the compressor;
a back flow preventing device that is provided between the
intermediate-pressure receiver and the intermediate-pressure
portion of the compressor and prevents back flow of the gas
refrigerant from the compressor to the intermediate-pressure
receiver; and a refrigerant-pressure control unit for controlling
the pressure of the refrigerant in the intermediate-pressure
receiver on the basis of the difference between a predetermined
reference enthalpy and a specific enthalpy of refrigerant
discharged from any one of the heat-side heat exchanger and the
using side heat exchanger.
[0010] In the above-described refrigerating machine, under normal
operation, a high pressure portion of the refrigerating machine
operates under a supercritical state.
[0011] According to the refrigerating machine of the present
invention, when the specific enthalpy of the refrigerant at the
exit of the high-pressure gas cooler (i.e., the refrigerant
discharged from the high-pressure gas cooler) is increased due to
increase of the external temperature, variation of the load or the
like, the two-stage expansion economizer cycle is formed. On the
other hand, when the specific enthalpy of the refrigerant at the
exit of the high-pressure gas cooler is reduced due to reduction of
the external temperature, variation of the load or the like,
one-stage expansion cycle is formed. Therefore, the optimum
performance can be kept with a simple construction.
BRIEF DESCRIPTION OF THE DRAWINGS
[0012] FIG. 1 is a refrigerant circuit diagram showing a
refrigerating machine according to an embodiment of the present
invention;
[0013] FIG. 2 is a pressure-enthalpy (ph) diagram of a
refrigeration cycle;
[0014] FIG. 3 is a diagram showing a control flow; and
[0015] FIG. 4 is a refrigerant circuit diagram of another
embodiment.
DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT
[0016] A preferred embodiment according to the present invention
will be described hereunder with reference to the accompanying
drawings.
[0017] FIG. 1 is a refrigerant circuit diagram showing an
embodiment of the present invention. A refrigerating machine 30 is
equipped with a two-stage compressor 1, a high-pressure gas cooler
3 for cooling high-pressure gas refrigerant, a first throttling
device 5, an intermediate pressure receiver 7 for adjusting a
refrigerant circulating amount, a second throttling device 9 and an
evaporator 11, and these parts are successively connected to one
another to thereby form a closed circuit. The first throttling
device 5 and the second throttling device 9 are designed so that
the opening degree of the restriction thereof is variable. By
varying the degree of the restriction, the pressure can be reduced
and a larger amount of gas refrigerant is generated until it
reaches the intermediate pressure receiver 7, and under this state
the refrigerant is supplied into the intermediate pressure receiver
7, whereby the separation coefficient in the intermediate pressure
receiver 7 can be varied.
[0018] The two-stage compressor 1 comprises a first-stage
compressing portion 1A and a second-stage compressing portion 1B.
The intermediate point (an intermediate pressure portion 1C)
between the first-stage compressing portion 1A and the second-stage
compressing portion 1B is connected to the upper portion of an
intermediate pressure receiver 7 by an intermediate pressure
refrigerant bypass circuit 13 for bypassing the intermediate
pressure refrigerant vapor in the intermediate pressure receiver 7
to the intermediate pressure portion 1C of the compressor 1, and
the intermediate pressure refrigerant bypass circuit 13 is provided
with a check valve (back flow preventing device) 15 having the
function of preventing back flow of refrigerant vapor from the
compressor 1 to the intermediate pressure receiver 7. the back flow
preventing device is not limited to the check valve 15, and an
opening/closing valve or the like may be used, for example.
[0019] Carbon dioxide refrigerant with which the high-pressure side
is set to a supercritical state under normal operation is sealingly
filled in the above-described refrigerant circuit. Ethylene,
diborane, ethane, nitrogen oxide or the like may be used as
refrigerant with which the high-pressure side is operated under
supercritical pressure.
[0020] In this construction, a refrigerant temperature sensor 40 is
secured to the exit of the high-pressure gas cooler 3, an
evaporation temperature sensor 41 is secured to the evaporator 11,
a suction temperature sensor 42 is secured to the suction side of
the two-stage compressor 1, a discharge temperature sensor 43 is
secured to the discharge side of the two-stage compressor 1, and an
intermediate pressure temperature sensor 44 is secured to the
intermediate pressure receiver 7. Furthermore, the respective
sensors 40 to 44, the first throttling device 5 and the second
throttling device 9 are connected to a controller 45.
[0021] In this construction, the controller 45 executes the
following control.
[0022] That is, when the specific enthalpy of refrigerant at the
exit of the high-pressure gas cooler 3 is not larger than the
enthalpy of saturated liquid corresponding to the pressure of the
intermediate pressure portion of the compressor under one-stage
expansion, at least one of the first throttling device 5 and the
second throttling device 9 is controlled so that the pressure of
the intermediate pressure receiver 7 is lower than the pressure of
the intermediate pressure portion of the compressor. For example,
at least one of the first and second throttling devices is
controlled so that the valve opening degree of the first throttling
device 5 is "small" and the valve opening degree of the second
throttling device 9 is "large". Furthermore, when the specific
enthalpy of the refrigerant at the exit of the high-pressure gas
cooler 3 is larger than the enthalpy of saturated liquid
corresponding to the pressure of the intermediate pressure portion
of the compressor, at least one of the first throttling device 5
and the second throttling device 9 is controlled so that the
pressure of the intermediate pressure receiver 7 is lower than the
pressure of saturated liquid whose enthalpy is substantially equal
to the specific enthalpy of the refrigerant at the exit of the
high-pressure gas cooler 3 and also higher than the pressure of the
intermediate pressure portion of the compressor. For example, at
least one of the throttling devices 5 and 9 is controlled so that
the valve opening degree of the first throttling device 5 is
"large" and the valve opening degree of the second throttling
device 9 is "small".
[0023] FIG. 2 is a pressure-enthalpy (ph) diagram of a
refrigeration cycle containing the two-stage compressor, and the
high-pressure side is operated under the supercritical state.
[0024] In FIG. 2, the pressure "P1" corresponds to the pressure of
the intermediate pressure portion of the compressor under one-stage
expansion is carried out, and the enthalpy "h1" corresponds to the
enthalpy of saturated liquid corresponding to the pressure "P1".
Here, when the external temperature increases or the like, the
specific enthalpy "h2" at the exit "E" of the high-pressure gas
cooler 3 is larger than the enthalpy "h1" of saturated liquid
corresponding to the pressure "P1" of the intermediate pressure
portion of the compressor. In this case, at least one of the first
throttling device 5 and the second throttling device 9 is
controlled so that the pressure "P2" in the intermediate pressure
receiver 7 ("F" in FIG. 2) is lower than the pressure "P3" of
saturated liquid whose enthalpy is substantially equal to the
specific enthalpy "h2" of the refrigerant at exit of the
high-pressure gas cooler 3, and also higher than the pressure "P1"
of the intermediate pressure portion of the compressor.
[0025] Specifically, at least one of the throttling devices 5 and 7
is controlled so that the valve opening degree of the first
throttling device 5 is "large" and the valve opening degree of the
second throttling device 9 is "small", for example.
[0026] Here, "A" represents the suction state of the first-stage
compressing portion 1A, "B" represents the discharge state of the
first-stage compressing portion 1A, "C" represents the suction
state of the second-stage compressing portion 1B and "D" represents
the discharge state of the second-stage compressing portion 1B. The
refrigerant discharged from the compressor 1 is passed through the
high-pressure gas cooler 3 and circulated to be cooled. "E"
represents the exit of the high-pressure gas cooler 3, that is, the
entrance of the first restricting deice 5, and "F" represents the
exit of the first throttling device 5. Under this state, the
refrigerant is two-phase mixture of gas/liquid. The ratio between
gas and liquid corresponds to the ratio between the length of a
line segment (gas) from "F" to "G" and the length of a line segment
(liquid) from "F" to "I".
[0027] This refrigerant enters the intermediate pressure receiver 7
under the two-phase mixture state. The gas refrigerant separated in
the intermediate pressure receiver 7 is controlled so that the
pressure "P2" of the intermediate pressure receiver 7 is higher
than the pressure "P1" of the intermediate pressure portion of the
compressor, and thus the refrigerant is passed through the check
valve 15 and then introduced into the intermediate pressure portion
1C of the compressor 1, that is, the intermediate point between the
first-stage compressing portion 1A and the second-stage compressing
portion 1B. "I" represents the exit state of the intermediate
pressure receiver 7, and the refrigerant passed through this exit
reaches the suction port of the second-stage compressing portion 1B
of "C" to be compressed in the second-stage compressing portion
1B.
[0028] The liquid refrigerant separated in the intermediate
pressure receiver 7 reaches the second throttling device 9. "G"
represents the exit of the intermediate pressure receiver 7, that
is, the entrance of the second throttling device 9, "H" represents
the exit of the second throttling device 9, and "A" represents the
exit of the evaporator 11, and also the suction port of the
first-stage compressing portion 1A. The liquid refrigerant entering
the evaporator 11 is evaporated while absorbing heat, and the
gas-phase refrigerant is returned to the suction port of the
first-stage compressing portion 1A.
[0029] In the above construction, even when the gas refrigerant
separated in the intermediate pressure receiver 7 is circulated to
the evaporator 11, it cannot be used for cooling. Accordingly, when
the gas refrigerant concerned is returned to the suction port of
the first-stage compressing portion 1A, the compression efficiency
is lowered.
[0030] This construction is a so-called two-stage expansion
economizer cycle. The gas refrigerant separated in the intermediate
pressure receiver 7 is introduced into the intermediate pressure
portion 1C of the two-stage compressor 1, so that the refrigerant
flow amount in the evaporator 11 is reduced and the compression
driving force of the first-stage compressing portion 1A is reduced.
Therefore, the pressure loss in the evaporator 11 is reduced, and
thus the performance of the refrigeration cycle can be enhanced. In
this construction, particularly, carbon dioxide refrigerant is
sealingly filled in the refrigerant circuit. Therefore, in the
ratio between the gas refrigerant and the liquid refrigerant
separated in the intermediate pressure receiver 7, an amount of the
gas component (the line segment between "F" and "G") is higher than
Freon (chlorofluorocarbon) refrigerant, and a larger amount of the
gas component is introduced into the intermediate pressure portion
1C of the compressor 1, thereby achieving higher performance.
[0031] When the external temperature is reduced or the like, the
exit state of the high-pressure gas cooler 3 moves to "E1". The
specific enthalpy "h3" of "E1" is smaller than the enthalpy "h1" of
the saturated liquid corresponding to the pressure "P1" of the
intermediate pressure portion of the compressor, and under this
state, only liquid-phase refrigerant exists in the intermediate
pressure receiver 7 ("F1") while no gas refrigerant exists in the
intermediate pressure receiver 7.
[0032] In this case, at least one of the first throttling device 5
and the second throttling device 9 is controlled so that the
pressure "P4" of the intermediate pressure receiver 7 is lower than
the pressure "P1" of the intermediate pressure portion of the
compressor. For example, at least one of the throttling devices 5
and 9 is controlled so that the valve opening degree of the first
throttling device 5 is "small" and the valve opening degree of the
second throttling device 9 is "large". When the pressure "P4" of
the intermediate pressure receiver 7 is lower than the pressure
"P1", the check valve 15 of FIG. 1 functions, and the
intercommunication between the intermediate pressure receiver 7 and
the intermediate pressure portion 1C of the compressor 1 is cut
off. All the liquid-phase refrigerant in the intermediate pressure
receiver 7 is passed through the evaporator 11 and then introduced
into the first-stage compressing portion 1A of the two-stage
compressor 1.
[0033] When the above operation is considered in FIG. 2, "A"
represents the suction of the first-stage compressing portion 1A,
"B1" represents the discharge of the first-stage compressing
portion 1A and "D1" represents the discharge of the second-stage
compressing portion 1B. The refrigerant discharged from the
compressor 1 is passed through the high-pressure gas cooler 3 and
circulated to be cooled. As described above, "E1" represents the
exit of the high-pressure gas cooler 3, that is, the entrance of
the first throttling device 5, and "F1" represents the exit of the
first throttling device 5. Under this state, only liquid-phase
refrigerant exists at "F1".
[0034] All the liquid refrigerant concerned reaches the second
throttling device 9. "H1" represents the exit of the second
throttling device 9, and "A" represents the exit of the evaporator
11 and also the suction port of the first-stage compressing portion
1A as described above. The liquid refrigerant entering the
evaporator 11 is evaporated while absorbing heat, and the gas-phase
refrigerant is returned to the suction port of the first-stage
compressing portion 1A.
[0035] In this embodiment, when the specific enthalpy "h2" of the
refrigerant at the exit of the high-pressure gas cooler 3 is larger
than the enthalpy "h1" of saturated liquid corresponding to the
pressure "P1" of the intermediate pressure portion of the
compressor due to increase of the external temperature, variation
of load or the like, the two-stage expansion economizer cycle is
formed. Conversely, when the specific enthalpy "h3" of the
refrigerant at the exit of the high-pressure gas cooler 3 is
reduced to be equal to or smaller than the enthalpy "h1" of
saturated liquid corresponding to the pressure "P1" of the
intermediate pressure portion of the compressor due to reduction of
the external temperature, the load variation or the like, one-stage
expansion cycle is formed. Therefore, the optimum performance
matched with the external temperature, the load variation or the
like can be kept with a simple construction.
[0036] FIG. 3 shows the control flow of the refrigerating machine.
During operation, the evaporation temperature Teva is detected by
the evaporation temperature senor 41 (S1) and the suction
temperature Tsuc is detected by the suction temperature sensor 42
(S2). Furthermore, the discharge temperature Tdis is detected by
the discharge temperature sensor 43 (S3), and the refrigerant
temperature Tm in the intermediate pressure receiver 7 is detected
by the intermediate pressure temperature sensor 44 (S4). The
refrigerant temperature Tout at the exit of the high-pressure gas
cooler 3 is detected by the refrigerant temperature sensor 40 (S5).
Then, the suction pressure Psuc is calculated from the evaporation
temperature Teva (S6), and the high side pressure Ph is calculated
from the discharge temperature Tdis (S7), the actual intermediate
pressure Pm in the intermediate pressure receiver 7 is calculated
from the refrigerant temperature Tm in the intermediate pressure
receiver 7 (S8), and the intermediate pressure (=the pressure of
the intermediate pressure portion of the compressor when one-stage
expansion is carried out) Pm1 as a reference in this control is
calculated from the suction pressure Psuc, the suction temperature
Tsuc and the high side pressure Ph (S9).
[0037] The enthalpy hLiq ("h1") of saturated liquid corresponding
to the intermediate pressure Pm1 is calculated from the
intermediate pressure Pm1 (S10), and the specific enthalpy hout
("h2") at the exit concerned is calculated from the refrigerant
temperature Tout at the exit of the high-pressure gas cooler 3 and
the high side pressure Ph (S1).
[0038] Subsequently, it is judged whether the specific enthalpy
hout is larger than the enthalpy hLiq (S12). When the specific
enthalpy hout is larger than the enthalpy hLiq (S12:Yes), at least
one of the first throttling device 5 and the second throttling
device 9 is controlled so as to satisfy the following condition:
the intermediate pressure Pm (i.e., the actual pressure in the
intermediate-pressure receiver)>the intermediate pressure Pm1
(i.e., the pressure of the intermediate-pressure portion of the
compressor) (S13). Specifically, the throttling devices 5 and 9 are
controlled so that the valve opening degree of the first throttling
device 5 is "large" and the valve opening degree of the second
throttling device 9 is "small", whereby the two-stage expansion
economizer cycle is formed.
[0039] On the other hand, when the specific enthalpy hout is not
larger than the enthalpy hLiq (S12: No), at least one of the first
throttling device 5 and the second throttling device 9 is
controlled so as to satisfy the following condition: the
intermediate pressure Pm (i.e., the actual pressure in the
intermediate-pressure receiver)<the intermediate pressure Pm1
(i.e., the pressure of the intermediate-pressure portion of the
compressor) (S14). Specifically, the throttling devices are
controlled so that the valve opening degree of the first throttling
device 5 is "small" and the valve opening degree of the second
throttling device 9 is "large", whereby the one-stage expansion
cycle is formed.
[0040] The suction pressure Psuc and the high side pressure Ph may
be determined by pressure sensors. With respect to the intermediate
pressure Pm1, a preset value may be stored in a memory.
[0041] FIG. 4 shows another embodiment of the refrigerating machine
of the present invention.
[0042] A refrigerating machine (air conditioner) 130 can perform
both the cooling and heating operation at the same time.
[0043] The refrigerating machine 130 comprises an outdoor unit 101
including a two-stage compressor 102, outdoor heat exchangers 103a
and 103b and outdoor expansion valves 127a and 127b, an indoor unit
105a including an indoor heat exchanger 106a and an indoor
expansion valve 118a, an indoor unit 105b including an indoor heat
exchanger 106b and an indoor expansion valve 118b, and a hot water
supply (stocking) unit 150 including a hot water stocking heat
exchanger 141, a hot water stocking tank 143, a circulating pump
145 and an expansion valve 147.
[0044] The outdoor unit 101, the indoor units 105a and 105b and the
hot water supply unit 150 are connected to one another through
inter-unit pipes 110. The refrigerating machine 130 can carry out
the cooling operation or the heating operation on the indoor units
105a and 105b at the same time or the mixing operation of the
cooling operation and the heating operation on the indoor units
105a and 105b while operating the hot water supply unit 150.
[0045] In the outdoor unit 101, one end of the outdoor heat
exchanger 103a is exclusively connected to the discharge pipe 107
or suction pipe 108 of the compressor 102 through a change-over
valve 109a or change-over valve 109b. Likewise, one end of the
outdoor heat exchanger 103b is exclusively connected to the
discharge pipe 107 or suction pipe 108 of the compressor 102
through a change-over valve 119a or 119b. Furthermore, an
accumulator 104 is disposed in the suction pipe 108.
[0046] The outdoor unit 101 is equipped with an outdoor control
device (not shown), and the outdoor control device controls the
compressor 102, the outdoor expansion valves 127a and 127b, the
change-over valves 109a, 119a, 109b and 119b and the overall
refrigerating machine 130. Furthermore, the refrigerating machine
130 is equipped with a temperature sensor S1 for detecting the
refrigerant temperature at the entrance of the accumulator 104,
temperature sensors S2 for detecting the refrigerant temperature of
the indoor heat exchangers 106a and 106b, temperature sensors S3
for detecting the refrigerant temperature of the outdoor heat
exchangers 103a and 103b, and a temperature sensor S4 for detecting
the refrigerant temperature at the exit of the compressor 102.
[0047] The compressor 102 is a two-stage compressor, and it has a
first-stage compressing portion 102A for compressing refrigerant at
the low-pressure suction side, and a second-stage compressing
portion 102B for compressing refrigerant at the high-pressure
discharge side. The compressor 102 is further equipped with an
intermediate pressure portion 102M that can introduce refrigerant
from the external into the intermediate portion between the
first-stage compressing portion 102A and the second-stage
compressing portion 102B.
[0048] The inter-unit pipe 110 is equipped with a high-pressure
pipe (high-pressure gas pipe) 111, a low-pressure pipe
(low-pressure gas pipe) 112 and an intermediate-pressure pipe
(liquid pipe) 113. The high-pressure pipe 111 is connected to the
discharge pipe 107, and the low-pressure pipe 112 is connected to
the suction pipe 108. The intermediate pressure pipe 113 is
connected to the other ends of the outdoor heat exchangers 103a and
103b through the outdoor expansion valves 127a and 127b.
[0049] An intermediate-pressure receiver (gas/liquid separator) 128
is connected between the intermediate pressure pipe 113 and each of
the outdoor expansion valves 127a and 127b. The intermediate
pressure receiver 128 is roughly equipped with a receiver main body
128A, a vapor exit pipe (serving as an intermediate-pressure
refrigerant bypass circuit) 128B, a first entrance/exit pipe 128C
and a second entrance/exit pipe 128D, and the vapor exit pipe 128B
of the intermediate pressure receiver 128 is connected to the
intermediate pressure portion 102M of the compressor 102 so that
gas-phase refrigerant is introduced from the vapor exit pipe 128B
into the compressor 102. The intermediate pressure receiver 128 is
designed as a bidirectional type gas/liquid separator into which
refrigerant can be introduced from any side of the outdoor heat
exchangers 103a and 103b and the indoor heat exchangers 106a and
106b.
[0050] One ends of the indoor heat exchangers 106a and 106b of the
indoor units 105a and 105b are connected to the high-pressure pipe
111 through discharge side valves 116a and 116b, and also connected
to the low-pressure pipe 112 through suction side valves 117a and
117b. Furthermore, the other ends of the indoor heat exchangers
106a and 106b of the indoor units 105a and 105b are connected to
the intermediate-pressure pipe 113 through the indoor expansion
valves 118a and 118b. The discharge side valve 116a and the suction
side valve 117a are controlled so that when one of these valves is
opened, the other valve is closed. Likewise, the discharge side
valve 116b and the suction side valve 117b are also controlled so
that when one of these valves is opened, the other valve is closed.
Accordingly, one end of each of the indoor heat exchangers 106a and
106b is selectively connected to one of the high-pressure pipe 111
and the low-pressure pipe 112 of the inter-unit pipe 110.
[0051] The indoor unit 105a (105b) is further equipped with an
indoor fan 123a (123b), a remote controller and an indoor control
device. Each indoor fan 123a (123b) is disposed in proximity to
each indoor heat exchanger 106a (106b), and blows air to the indoor
heat exchanger 106a (106b). Each remote controller is connected to
the indoor unit 105a (105b), and each indoor unit 105a (105b)
outputs a cooling or heating operation instruction, a stop
instruction, etc. to each indoor control device.
[0052] In the hot water supply unit 150, one end of the hot water
stocking heat exchanger 141 is connected to the high-pressure pipe
111 through a change-over valve 148, and the other end of the hot
water stocking heat exchanger 141 is connected to the intermediate
pressure pipe 113 through an expansion valve 147. A water pipe 146
is connected to the hot water stocking heat exchanger 141, and a
hot water stocking tank 143 is connected to the water pipe 146
through a circulating pump 145.
[0053] In this embodiment, carbon dioxide refrigerant is sealingly
filled in the outdoor unit 101, the indoor units 105a and 105b and
the pipes of the hot water supply unit 150 and the inter-unit pipe
110.
[0054] Furthermore, a check valve (back flow preventing device) 151
having the function of preventing back flow of refrigerant vapor
from the compressor 102 to the intermediate pressure receiver 128
is provided to the vapor exit pipe 128B of the intermediate
pressure receiver 128. The back flow preventing device is not
limited to the check valve 151, and for example, an opening/closing
valve or the like may be used.
[0055] When the refrigerating machine 130 carries out cooling
operation or heating operation on the indoor units 105a and 105b at
the same time or carries out both the cooling operation and the
heating operation (i.e., the mixing operation) on the indoor units
105a and 105b while operating the hot water supply unit 150, some
of the heat exchangers 103, 106, 141 function as a heat-radiation
side heat exchanger(s). The gas-phase or liquid-phase component in
the refrigerant before the refrigerant enters the
intermediate-pressure receiver 128 is varied in accordance with the
exit temperature of the heat-radiation side heat exchanger
(corresponding to the high-pressure gas cooler 3 of FIG. 1) as
described above. When the exit temperature of the heat-radiation
side heat exchanger increases or the like, the amount of the
gas-phase component of the refrigerant before the refrigerant
enters the intermediate-pressure receiver 128 increases, and thus
the amount of the gas-phase refrigerant introduced into the
intermediate pressure portion 102M of the compressor 102 is
increased, so that the efficiency of the refrigeration cycle can be
enhanced by the degree corresponding the amount of the gas-phase
component which does not contribute to cooling and is not
circulated to the low-pressure circuit subsequent to the
intermediate-pressure pipe 113.
[0056] Particularly, according to the refrigerating machine having
the above construction, since carbon dioxide refrigerant is
sealingly filled in the refrigerant circuit, in the ratio of the
gas-phase component and the liquid-phase component separated in the
intermediate-pressure receiver 128, the amount of the gas-phase
component separated in the intermediate-pressure receiver 128 is
larger as compared with Freon (chlorofluorocarbon) refrigerant, and
a large amount of gas-phase component can be introduced into the
intermediate pressure portion 102M of the compressor 102, thereby
achieving a higher efficiency.
[0057] On the other hand, for example when the exit temperature of
the heat-radiation side heat exchanger is reduced and thus most of
refrigerant in the intermediate-pressure receiver 128 is
liquid-phase component, the efficiency of the refrigeration cycle
would be rather reduced if the liquid-phase component concerned is
introduced to the intermediate-pressure portion 102M of the
compressor 102.
[0058] In this case, referring to FIG. 2, at least one of the
outdoor expansion values 127a, 127b, the expansion valve 147 and
the indoor expansion valves 118a, 118b is controlled so that the
pressure "P4" in the intermediate-pressure receiver 128 is lower
than the pressure "P1" of the intermediate-pressure portion of the
compressor. At this time, the check valve 151 of FIG. 4 functions,
and the intercommunication between the intermediate-pressure
receiver 128 and the intermediate pressure portion 102M of the
compressor 102 is cut off. Therefore, all the liquid-phase
refrigerant in the intermediate-pressure receiver 128 is passed
through the evaporator, and introduced into the first-stage
compressing portion 102A of the two-stage compressor 102.
[0059] In other words, when the specific enthalpy "h2" of the
refrigerant at the exit of the heat-radiation side heat exchanger
increases to be larger than the enthalpy "h1" of saturated liquid
corresponding to the pressure "P1" of the intermediate-pressure
portion of the compressor due to increase of the external
temperature, variation of the load or the like, the two-stage
expansion economizer cycle is formed. Conversely, when the specific
enthalpy "h3" of the refrigerant at the exit of the heat-radiation
side heat exchanger is reduced to be equal to or smaller than the
enthalpy "h1" of saturated liquid corresponding to the pressure
"P1" of the intermediate-pressure portion of the compressor due to
reduction in external temperature or the like, one-stage expansion
cycle is formed. Accordingly, the optimum performance which is
matched with the external temperature, the load condition or the
like can be kept with a simple construction. Here, the "external
temperature" means the temperature of a medium to be heat-exchanged
with refrigerant in the heat-radiation side heat exchanger.
Specifically, the external temperature is the indoor temperature
when the heating operation is carried out, the outdoor air
temperature when the outdoor heat exchanger functions as a radiator
or the water temperature at the entrance of the water supply
(stocking) heat exchanger when the water stocking operation is
carried out.
[0060] The present invention is not limited to the above-described
embodiment, and various modifications may be made without departing
from the subject matter of the present invention.
[0061] In the foregoing description, at least one of the first and
second throttling devices 5 and 9 is controlled so that the valve
opening degree of one of the throttling devices is merely set to be
"large" while the valve opening degree of the other throttling
device is set to be "small" in accordance with whether the enthalpy
of the refrigerant at the exit of the high-pressure gas cooler 3 is
larger or not larger than the enthalpy of saturated liquid
corresponding to the pressure of the intermediate-pressure portion
of the compressor under one-stage expansion. In order to more
finely control the amount of gas refrigerant to be supplied from
the intermediate-pressure receiver to the intermediate-pressure
portion of the compressor, the valve opening degrees of the first
and second throttling devices 5 and 9 may be more finely controlled
on the basis of the difference value between the enthalpy of the
refrigerant at the exit of the high-pressure gas cooler 3 and the
enthalpy of saturated liquid corresponding to the pressure of the
intermediate-pressure portion of the compressor under one-stage
expansion.
* * * * *