CN113821997A - Method for calculating flow coefficient of regulating valve of refrigeration compressor performance testing device - Google Patents

Method for calculating flow coefficient of regulating valve of refrigeration compressor performance testing device Download PDF

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CN113821997A
CN113821997A CN202110899968.2A CN202110899968A CN113821997A CN 113821997 A CN113821997 A CN 113821997A CN 202110899968 A CN202110899968 A CN 202110899968A CN 113821997 A CN113821997 A CN 113821997A
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regulating valve
pressure
refrigerant
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CN113821997B (en
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袁旭东
贾磊
昝世超
王汝金
张秀平
吴俊峰
赵盼盼
周到
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HEFEI GENERAL ENVIRONMENT CONTROL TECHNOLOGY CO LTD
Hefei General Machinery Research Institute Co Ltd
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HEFEI GENERAL ENVIRONMENT CONTROL TECHNOLOGY CO LTD
Hefei General Machinery Research Institute Co Ltd
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Abstract

A method for calculating the flow coefficient of a regulating valve of a refrigeration compressor performance testing device comprises the following steps: aiming at performance design parameters of a refrigeration compressor, establishing a working condition running chart of the compressor in a two-dimensional coordinate system taking evaporation pressure as an abscissa and condensation pressure as an ordinate, and obtaining a working condition inflection point on the working condition running chart; obtaining variable parameters and fixed parameters corresponding to inflection points of various working conditions; and obtaining the maximum value of the initial upper limit values of the flow coefficients of the regulating valve to be calculated under different working condition inflection points as the upper limit value of the flow coefficient of the regulating valve to be calculated, and obtaining the minimum value of the initial lower limit values of the flow coefficients of the regulating valve to be calculated under different working condition inflection points as the lower limit value of the flow coefficient of the regulating valve to be calculated. According to the invention, the finally obtained flow coefficient adjusting range of the adjusting valve to be calculated can meet the requirements of any working condition inflection point, and support is provided for parameter selection of the adjusting valve of the performance testing device of the refrigeration compressor.

Description

Method for calculating flow coefficient of regulating valve of refrigeration compressor performance testing device
Technical Field
The invention relates to the field of compressor testing, in particular to a method for calculating a flow coefficient of an adjusting valve of a performance testing device of a refrigeration compressor.
Background
After the refrigeration compressor is designed to obtain design (or specification) parameters, a matched testing device needs to be developed in order to test and verify whether the performance of the compressor meets the design requirements. The air-supplementing type centrifugal screw-type refrigerant compressor is a main form of compressor product, and is characterized by that the compressor is equipped with air-supplementing port to implement quasi-two-stage compression of refrigerant in the compressor. Because single-mechanism refrigeration capacity represented by a centrifugal type and screw type refrigerant compressor is large, in order to realize high economical efficiency of the test process, a refrigerant gas cooling method flow specified by standards GB/T5773 and JB/T12843 and 2016 is generally adopted as a test loop flow of the performance test device of the centrifugal type and screw type refrigerant compressor. The regulating valve is an indispensable or deficient part in the testing loop of the refrigeration compressor, and the regulating precision and the regulating range of the regulating valve directly influence the applicable range and the testing precision of the testing loop of the compressor. Because the design parameters (such as refrigerating capacity, operation conditions and the like) of the refrigeration compressor are different, and the range of the design parameters is also different, accurate regulating valve parameters need to be provided for meeting the test requirements of all working condition ranges of the compressor, however, no regulating valve parameter calculation method suitable for a refrigeration compressor test loop based on a gas cooling method process exists in the industry at present, so that the selection of a regulating valve in the compressor test loop has no clear index, and the cost and the test requirements of the compressor test loop cannot be balanced.
Disclosure of Invention
In order to overcome the defect that no regulating valve parameter calculation method suitable for a refrigeration compressor test circuit exists in the prior art, the invention provides a regulating valve flow coefficient calculation method of a refrigeration compressor performance test device.
The invention adopts the following technical scheme:
a method for calculating the flow coefficient of a regulating valve of a refrigeration compressor performance testing device comprises the following steps:
s1, aiming at performance design parameters of the refrigeration compressor, establishing a working condition running chart of the compressor in a two-dimensional coordinate system which takes evaporation pressure as an abscissa and condensation pressure as an ordinate, and obtaining a working condition inflection point on the working condition running chart, wherein the working condition inflection point is a point where any two line segments on the working condition running chart intersect; the condensation pressure and the evaporation pressure corresponding to the working condition inflection point are performance design parameters of the refrigeration compressor corresponding to the working condition inflection point;
s2, obtaining variable parameters and fixed parameters corresponding to the inflection points of the working conditions; the variable parameters are compressor design parameters with a variable range, and the fixed parameters are compressor design parameters with fixed values; exhaustively exhausting a variable parameter extreme value set, wherein each variable parameter extreme value set comprises the maximum value or the minimum value of each variable parameter; setting parameter sets corresponding to the variable parameter extreme value sets one by one, wherein the parameter sets are combined into a union set of the corresponding variable parameter extreme value sets and all fixed parameters; calculating the flow coefficient of each regulating valve to be calculated corresponding to the working condition inflection point by combining each parameter set; aiming at each regulating valve to be calculated, acquiring the maximum value and the minimum value in the flow coefficient of the regulating valve to be calculated corresponding to each parameter set, and respectively taking the maximum value and the minimum value as the initial upper limit value and the initial lower limit value of the flow coefficient corresponding to the regulating valve to be calculated;
s3, obtaining the maximum value of the initial upper limit values of the flow coefficient of the regulating valve to be calculated under different working condition inflection points as the upper limit value of the flow coefficient of the regulating valve to be calculated, and obtaining the minimum value of the initial lower limit values of the flow coefficient of the regulating valve to be calculated under different working condition inflection points as the lower limit value of the flow coefficient of the regulating valve to be calculated.
The invention has the advantages that:
(1) firstly, calculating the flow coefficient of the regulating valve to be calculated under each group of parameter set by adopting an exhaustive parameter set mode aiming at the working condition inflection point, namely obtaining the regulating range of the flow coefficient of each regulating valve to be calculated under the determined working condition inflection point; and then, combining the adjusting range of the flow coefficient of the adjusting valve to be calculated under each working condition inflection point to obtain the total adjusting range of the flow coefficient of the adjusting valve to be calculated. Therefore, the finally obtained flow coefficient adjusting range of the adjusting valve to be calculated can meet the requirements of any working condition inflection point, and support is provided for parameter selection of the adjusting valve of the performance testing device of the refrigeration compressor.
(2) The invention provides a specific calculation method of the flow coefficient of the regulating valve, which is beneficial to quickly calculating the flow coefficient of the regulating valve, so that the implementation of the method is more convenient, quicker and more effective.
(3) In the invention, part of compressor design parameters are listed, and the flow coefficient of the regulating valve is calculated according to the design parameters, so that the calculation result is more accurate and reliable.
(4) The method for calculating the flow coefficient of the regulating valve is suitable for various regulating valves such as gas coolers, flash evaporators and the like, and is wide in application range.
Drawings
FIG. 1 is a working condition operation diagram of an embodiment;
FIG. 2 is a flow chart of a method for calculating a flow coefficient of a regulating valve of a performance testing device of a refrigeration compressor, which is provided by the invention;
FIG. 2(a) is a method for calculating the initial upper limit value and the initial lower limit value of the flow coefficient corresponding to the regulating valve;
FIG. 2(b) is a flow chart of a calculation method of flow coefficients of regulating valves to be calculated in a compressor test loop under any operating condition inflection point state;
FIG. 3 is a system diagram of a refrigerant compressor performance testing device based on a refrigerant gas cooling process;
figure 4 is a pressure enthalpy diagram for the test apparatus.
The figure is as follows: a condenser 1, a gas cooler 2, a flash tank 3, an oil separator 4, a liquid regulating valve 5 and a gas regulating valve 6; the system comprises a gas cooler liquid branch a, a gas cooler gas branch b, a suction gas overheating reducing branch c, a flash tank liquid branch d, a flash tank gas branch e, a motor cooling branch f and a measured compressor middle gas supplementing branch g.
Noun interpretation
The testing device comprises: the test system is used for measuring the thermodynamic properties of the refrigeration compressor, such as the refrigerating capacity, the heating capacity and the like;
testing a loop: the refrigerant in the test unit is circulated through the complete thermodynamic cycle from the compressor discharge to the compressor suction.
Mass flow rate: mass of refrigerant passing per unit time.
First-stage compression: the process of compressing the refrigerant at the compressor suction to the output of the low pressure stage impeller.
Air supplement in the middle of the compressor: and a gas supplementing port is arranged at a position in the middle of a refrigerant gas flowing area in the compressor, and the refrigerant gas is supplemented into the compressor through the gas supplementing port.
Mixing the compressor intermediate air make-up refrigerant with the first-stage compressed refrigerant: the refrigerant from the air supplement port in the middle of the compressor to be tested is mixed with the refrigerant from the output port of the low-pressure stage impeller in the volute with fixed volume, and the mixed refrigerant enters the input port of the high-pressure stage impeller.
Mixing the compressor intermediate air make-up refrigerant and the first-stage compressed refrigerant to obtain a first specific volume: and after the compressor intermediate air-supplementing refrigerant is mixed with the first-stage compressed refrigerant, solving the specific volume of the mixed refrigerant based on the pressure and specific enthalpy of the mixed refrigerant of the compressor intermediate air-supplementing refrigerant and the first-stage compressed refrigerant.
The second specific volume of the mixed refrigerant after the middle air-supplementing refrigerant of the compressor and the refrigerant after the first-stage compression is: and after the refrigerant for supplementing air in the middle of the compressor is mixed with the refrigerant after the first-stage compression, the specific volume of the mixed refrigerant is solved based on the specific volume of the refrigerant after the first-stage compression of the compressor and the relative air supplementing rate.
Condensing pressure: the pressure in the condenser in the testing device and the condensing pressure belong to the test condition parameters to be regulated, and the values of the condensing pressure and the pressure of the exhaust gas of the compressor are regarded as equal in the testing process of the compressor according to the GB/T5773-one 2016 specification.
Evaporation pressure: the pressure in the evaporator of the refrigeration system is measured in the actual application of the compressor, the evaporation pressure belongs to the test working condition parameter to be regulated, and according to the regulation of GB/T5773 and 2016, the numerical values of the evaporation pressure and the suction pressure of the compressor are regarded as equal in the testing process of the compressor.
Condensation temperature: the condensing pressure corresponds to the refrigerant saturation temperature.
Evaporation temperature: the evaporation pressure corresponds to the refrigerant saturation temperature.
Detailed Description
The compressor in this case refers to the compressor under test in the performance test. In the scheme, all the isentropic specific enthalpies refer to the isentropic specific enthalpies in the corresponding state of the refrigerant in the tested compressor, and all the specific enthalpies refer to the specific enthalpies in the corresponding state of the refrigerant in the tested compressor; for convenience of representation, the refrigerant isentropic specific enthalpy is simplified into isentropic specific enthalpy, and the refrigerant specific enthalpy is simplified into specific enthalpy.
The method for calculating the flow coefficient of the regulating valve of the refrigeration compressor performance testing device comprises the following steps.
S1, aiming at performance design parameters of the refrigeration compressor, establishing a working condition running chart of the compressor in a two-dimensional coordinate system which takes evaporation pressure as an abscissa and condensation pressure as an ordinate, and obtaining a working condition inflection point on the working condition running chart, wherein the working condition inflection point is a point where any two line segments on the working condition running chart intersect. And the condensing pressure and the evaporating pressure corresponding to the working condition inflection point are performance design parameters of the refrigeration compressor corresponding to the working condition inflection point.
S2, obtaining variable parameters and fixed parameters corresponding to the inflection points of the working conditions; the variable parameters are compressor design parameters with a variable range, and the fixed parameters are compressor design parameters with fixed values; exhaustively exhausting a variable parameter extreme value set, wherein each variable parameter extreme value set comprises the maximum value or the minimum value of each variable parameter; setting parameter sets corresponding to the variable parameter extreme value sets one by one, wherein the parameter sets are combined into a union set of the corresponding variable parameter extreme value sets and all fixed parameters; calculating the flow coefficient of each regulating valve to be calculated corresponding to the working condition inflection point by combining each parameter set; and aiming at each regulating valve to be calculated, acquiring the maximum value and the minimum value in the flow coefficient of the regulating valve to be calculated corresponding to each parameter set as the initial upper limit value and the initial lower limit value of the flow coefficient corresponding to the regulating valve to be calculated respectively.
S3, obtaining the maximum value of the initial upper limit values of the flow coefficient of the regulating valve to be calculated under different working condition inflection points as the upper limit value of the flow coefficient of the regulating valve to be calculated, and obtaining the minimum value of the initial lower limit values of the flow coefficient of the regulating valve to be calculated under different working condition inflection points as the lower limit value of the flow coefficient of the regulating valve to be calculated.
In this way, in this embodiment, first, an exhaustive parameter set is adopted for the operating condition inflection point, and the flow coefficient of the regulating valve to be calculated under each set of parameter sets is calculated, that is, the regulating range of the flow coefficient of each regulating valve to be calculated under the determined operating condition inflection point is obtained; and then, combining the adjusting range of the flow coefficient of the adjusting valve to be calculated under each working condition inflection point to obtain the total adjusting range of the flow coefficient of the adjusting valve to be calculated. Therefore, the finally obtained flow coefficient adjusting range of the adjusting valve to be calculated can meet the requirements of any working condition inflection point, and support is provided for parameter selection of the adjusting valve of the performance testing device of the refrigeration compressor.
In this embodiment, step S2 specifically includes the following steps:
s21, obtaining the condensing pressure p corresponding to the working condition inflection point rcAnd the evaporation pressure peDividing the compressor design parameters corresponding to the working condition inflection points r into variable parameters and fixed parameters, wherein the variable parameters comprise:
Figure BDA0003199387840000061
Figure BDA0003199387840000062
i-th of the operating condition inflection point r staterIndividual variable parameter, 1 ≦ ir≤mr，mrFor the number of the variable parameters to be,
Figure BDA0003199387840000063
Figure BDA0003199387840000064
and
Figure BDA0003199387840000065
respectively represent
Figure BDA0003199387840000066
The upper limit value and the lower limit value of the value range; the fixed parameters include:
Figure BDA0003199387840000067
Figure BDA0003199387840000068
j represents the operating point r staterA fixed parameter, j is more than or equal to 1r≤nr；nrIs the number of fixed parameters.
S22 set of extreme values of exhaustive variable parameters
Figure BDA0003199387840000069
At least one parameter in any two variable parameter extreme value sets is different;
order to
Figure BDA00031993878400000610
Or
Figure BDA00031993878400000611
Figure BDA00031993878400000612
To represent
Figure BDA00031993878400000613
Upper limit of value range
Figure BDA00031993878400000614
Or a lower limit value
Figure BDA00031993878400000615
Any set of variable parameter extrema is written as:
Figure BDA00031993878400000616
s23, setting parameter sets corresponding to the variable parameter extreme value sets one by one
Figure BDA00031993878400000617
Figure BDA00031993878400000618
Wherein the content of the first and second substances,
Figure BDA00031993878400000619
krthe sequence number is shown to indicate that,
Figure BDA00031993878400000620
variable parameter corresponding to inflection point r of working conditionAnd the number of the extreme value sets of the variable parameters corresponding to the inflection points r of the working conditions is equal to the number of the parameter sets.
And S24, calculating the flow coefficient of each regulating valve to be calculated in the compressor test loop under the condition of the working condition inflection point r according to each group of parameter set, and respectively recording the maximum value and the minimum value of the flow coefficient of each regulating valve to be calculated as the initial upper limit value and the initial lower limit value of the flow coefficient corresponding to the regulating valve.
And S25, repeating the steps S21 to S24, and calculating the initial upper limit value and the initial lower limit value of the flow coefficient corresponding to each regulating valve to be calculated under each working condition inflection point.
Specifically, in the present embodiment, the compressor design parameters include: compressor intermediate air supply pressure pmidCompressor middle air supply superheat degree delta TmidCompressor suction superheat degree delta Tcom,iDesign value of cooling capacity QeIsentropic efficiency ηsRelative air supplement rate a and superheat degree delta T of outlet of gas coolergl,oAnd pressure loss coefficient xi in air supply processp(ii) a If the motor cooling branch exists, the method further comprises the following steps: compressor motor cooling heat load QcoolSuperheat degree delta T of motor cooling branch outletf,o. That is, the variable parameters are some of the above compressor design parameters, and the remaining compressor design parameters are fixed parameters. It is worth noting that in the process of testing the compressor, due to different testing purposes and indexes, types of variable parameters adopted by inflection points of different working conditions in the test may be different, and specific values of compressor design parameters corresponding to the inflection points of different working conditions may also be different.
In addition, it is worth noting that in the state of determined working condition inflection point, the corresponding condensation pressure p iscAnd the evaporation pressure peIs determined so that the condensing pressure pcAnd the evaporation pressure peOnly as fixed parameters.
In this embodiment, the control valve to be calculated includes: a gas cooler gas regulating valve, a gas cooler liquid regulating valve, a reduced suction superheat regulating valve, a flash tank gas regulating valve and a flash tank liquid regulating valve; if the motor cooling branch exists, the regulating valve to be calculated further comprises a motor cooling branch regulating valve.
In step S24, the initial upper limit value and the initial lower limit value of the flow coefficient corresponding to the regulating valve to be calculated for the operating condition inflection point r may be represented by combining the following table 1.
Table 1: corresponding relation between each parameter set and flow coefficient of regulating valve to be calculated under working condition inflection point r state
Figure BDA0003199387840000071
Figure BDA0003199387840000081
The flow coefficient of the liquid regulating valve of the gas cooler in the inflection point r state of the working condition is recorded as
Figure BDA0003199387840000082
The upper limit value of the initial flow coefficient of the liquid regulating valve of the gas cooler is recorded as
Figure BDA0003199387840000083
Initial selection lower limit value of flow coefficient of liquid regulating valve of gas cooler
Figure BDA0003199387840000084
The flow coefficient of the gas regulating valve of the gas cooler is recorded as
Figure BDA0003199387840000085
The upper limit value of the initial flow coefficient of the gas regulating valve of the gas cooler is recorded as
Figure BDA0003199387840000086
Initial selection lower limit value of flow coefficient of gas regulating valve of gas cooler
Figure BDA0003199387840000087
Reducing the flow of a suction superheat regulating valveThe coefficient is recorded as
Figure BDA0003199387840000088
The upper limit value of the initial selection of the flow coefficient for reducing the suction superheat regulating valve is recorded as
Figure BDA0003199387840000089
Lower limit value for initially selecting flow coefficient for reducing air suction overheating regulating valve
Figure BDA00031993878400000810
The flow coefficient of the liquid regulating valve of the flash tank is recorded as
Figure BDA00031993878400000811
The initial upper limit value of the flow coefficient of the liquid regulating valve of the flash tank is recorded as
Figure BDA00031993878400000812
Initial selection lower limit value of flow coefficient of flash tank liquid regulating valve
Figure BDA00031993878400000813
The flow coefficient of the gas regulating valve of the flash device is recorded as
Figure BDA00031993878400000814
The initial upper limit value of the flow coefficient of the gas regulating valve of the flash tank is recorded as
Figure BDA00031993878400000815
Initial selection lower limit value of flow coefficient of flash tank liquid regulating valve
Figure BDA00031993878400000816
The flow coefficient of the motor cooling branch regulating valve is recorded as
Figure BDA00031993878400000817
The initial upper limit value of the flow coefficient of the motor cooling branch regulating valve is recorded as
Figure BDA00031993878400000818
Motor cooling branch regulationLower limit value for initial selection of flow coefficient of valve
Figure BDA00031993878400000819
Figure BDA00031993878400000820
Wherein max { } denotes a maximum function, and min { } denotes a minimum function.
The following formula (4) is obtained by the same principle:
Figure BDA00031993878400000821
Figure BDA00031993878400000822
Figure BDA00031993878400000823
Figure BDA00031993878400000824
Figure BDA00031993878400000825
specifically, in this embodiment, in step S24, the method for calculating the flow coefficient of each regulating valve to be calculated in the compressor test loop under the condition inflection point state according to any one set of parameter sets includes the following steps:
s241, calculating the pressure p after the first-stage compression of the compressor according to the parameter set1st，
p1st＝(p1st,upper+p1st,low)/2 (5)；
Wherein p is1st,upperIs the upper limit of the pressure after one-stage compression of the compressorValue, p1st,lowThe lower limit value of the pressure after the first-stage compression of the compressor; p is a radical of1st,upperIs the compressor intermediate make-up air pressure pmid，p1st,lowHas an initial value of 0.
S242, calculating a first specific volume v after the compressor intermediate air supply refrigerant and the first-stage compressed refrigerant are mixedmid,mAnd the second specific volume v of the mixture of the compressor intermediate air make-up refrigerant and the refrigerant after the first-stage compressionmid,m'。
S243, when | vmid,m-vmid,m' | > ε if vmid,m＞vmid,m', then set p1st,upper＝p1stThen, the step S241 is returned to calculate the pressure p after the first-stage compression of the compressor1st(ii) a If v ismid,m＜vmid,m', then set p1st,low＝p1stAnd then returns to step S241 to recalculate the compressor one-stage compression pressure p1st(ii) a ε is the convergence condition parameter.
S244, when | vmid,m-vmid,mWhen | ≦ epsilon, calculating the specific enthalpy h of the compressor exhaustcom,o；
S245, calculating the mass flow m of the liquid inlet of the gas cooleraGas cooler gas inlet mass flow mbAnd the mass flow m of the suction overheating branch is reducedc；
When the performance testing device of the refrigeration compressor does not have a motor cooling branch, a calculation model based on matrix operation is as follows:
Figure BDA0003199387840000091
when the refrigerating compressor performance testing device is provided with a motor cooling branch, a calculation model based on matrix operation is as follows:
Figure BDA0003199387840000101
wherein m iscom,iFor compressor suctionFlow rate, hcom,iIs the compressor specific enthalpy of suction, hc,lSpecific enthalpy, h, of saturated liquid of refrigerant corresponding to condensing pressuregl,oIs the specific enthalpy of the outlet of the gas cooler, hgl,oAccording to the evaporation pressure peAnd gas cooler outlet superheat degree delta Tgl,oDetermination of hcom,oSpecific enthalpy of compressor discharge, hf,oOutlet specific enthalpy, m, of the motor cooling branchfMass flow of the motor cooling branch.
In particular, the above equations (6) and (7) are particularly suitable for the air make-up type refrigeration compressor.
S246, the specific enthalpy h of the saturated liquid of the refrigerant corresponding to the condensing pressurec,lCompressor middle air supply inlet specific enthalpy hmid,iSpecific enthalpy h of compressor dischargecom,oCompressor suction mass flow mcom,iCalculating the mass flow m of the liquid inlet of the flash tankdAnd the mass flow m of the gas inlet of the flash tankeThe calculation model based on matrix operation is as follows:
Figure BDA0003199387840000102
wherein a represents the relative air supplement rate of the compressor.
S247, according to the condensing pressure pcEvaporation pressure peAnd gas cooler liquid inlet mass flow maCalculating the flow coefficient of the liquid regulating valve of the gas cooler; according to the condensing pressure pcEvaporation pressure peAnd gas cooler gas inlet mass flow mbCalculating the flow coefficient of the gas regulating valve of the gas cooler; according to the condensing pressure pcEvaporation pressure peAnd reducing mass flow m of the suction superheat branchcCalculating a flow coefficient for reducing the suction superheat branch regulating valve; according to the condensing pressure pcEvaporation pressure peAnd the liquid inlet mass flow m of the flash tankdCalculating the flow coefficient of the liquid regulating valve of the flash tank; according to the condensing pressure pcEvaporation pressure peAnd the mass flow m of the gas inlet of the flash tankeComputing flashThe flow coefficient of the gas regulating valve of the device; according to the condensing pressure pcThe evaporation pressure peAnd mass flow m of motor cooling branchfAnd calculating the flow coefficient of the motor cooling branch regulating valve.
Specifically, in the present embodiment, the calculation of each parameter refers to the following equation:
vmid,m＝f(pmid,m,hmid,m)
Figure BDA0003199387840000111
pmid,m＝p1st+(pmid-p1st)ξp
hmid,m＝(h1st,o+ahmid,i)/(1+a) (9)；
wherein v ismid,mShowing the first specific volume, v, of the mixture of the intermediate air-supplementing refrigerant and the refrigerant after the first-stage compressionmid,m' represents a second specific volume after the compressor intermediate air make-up refrigerant and the refrigerant after the first-stage compression are mixed; f () represents a functional relationship of physical properties of the refrigerant, pmid,mThe pressure h after the compressor intermediate air make-up refrigerant and the refrigerant after the first-stage compression are mixed is shownmid,mThe specific enthalpy of the mixed intermediate air-replenishing refrigerant and the first-stage compressed refrigerant of the compressor is represented; v. of1st,oIndicating the specific volume, v, of the refrigerant after one-stage compression in the compressor1st,oBased on the physical function relationship of the refrigerant and according to the pressure p after the first-stage compression of the compressor1stAnd specific enthalpy h after first-stage compression of compressor1st,oDetermining; a represents the relative air supplement rate of the compressor; p is a radical ofmidIndicating the intermediate supply pressure, ξ of the compressorpRepresenting the pressure loss coefficient of the air replenishing process; h is1st,oRepresents the specific enthalpy, h, of the compressor after the first stage of compressionmid,iRepresents the specific enthalpy, h, of the intermediate charge air inlet of the compressormid,iAccording to the intermediate air supply pressure p of the compressormidAnd compressor middle air supply superheat degree delta TmidAnd (4) determining.
Figure BDA0003199387840000112
Wherein h iscom,oSpecific enthalpy of compressor discharge, hcom,o,sRepresents the isentropic specific enthalpy, h, of the compressor discharge gascom,o,sAccording to the condensing pressure p based on the physical function relationship of the refrigerantcAnd the specific entropy s of the mixture of the intermediate air-supplementing refrigerant of the compressor and the refrigerant after the first-stage compressionmid,mDetermination of smid,mAccording to the pressure p after the intermediate air-supplementing refrigerant of the compressor is mixed with the refrigerant after the first-stage compressionmid,mAnd the specific enthalpy h of the mixed intermediate air-replenishing refrigerant and first-stage compressed refrigerant of the compressormid,mDetermining; etasIndicating compressor isentropic efficiency.
Figure BDA0003199387840000121
Wherein h is1st,oRepresents the specific enthalpy, h, of the compressor after the first stage of compression1st,o,sRepresents the isentropic specific enthalpy, h, of the compressor after the first-stage compression1st,o,sBased on the physical function relation of the refrigerant, the pressure p after the first-stage compression of the compressor is obtained1stAnd compressor specific suction entropy scom,iDetermining; h iscom,iExpressing the specific enthalpy, eta, of the compressor suctionsRepresenting compressor isentropic efficiency, scom,iAnd hcom,iAre all according to the evaporation pressure peAnd degree of superheat delta T of suction gas of compressorcom,iAnd (4) determining.
mcom,i＝Qe/(hcom,i-hmid,l) (12)；
Wherein m iscom,iRepresenting compressor suction mass flow, hcom,iRepresents compressor suction specific enthalpy; h ismid,lIndicating the compressor intermediate charge pressure pmidCorresponding specific enthalpy of saturated liquid, hmid,lAccording to the compressor intermediate air supplement pressure p based on the physical function relation of the refrigerantmidDetermining; qeIndicating the refrigerating capacity of the compressor, QeIs a design value.
mf＝Qcool/(hf,o-hc,l) (13)；
Wherein m isfRepresenting the mass flow, Q, of the cooling branch of the motorcoolRepresents the cooling heat load of the motor, hf,oAccording to the evaporation pressure p based on the physical function relationship of the refrigeranteAnd superheat degree delta T of motor cooling branch outletf,oDetermination of hc,lIndicating the specific enthalpy of the saturated liquid of the refrigerant corresponding to the condensing pressure.
In order to facilitate understanding of those skilled in the art, the following describes details of the derivation process of the formula and the calculation model in the embodiment of the present invention with reference to fig. 3 and 4.
First, the main loop of the testing device circulates
As shown in fig. 3, the branch of the main circuit cycle of the test apparatus comprises: the system comprises a gas cooler liquid branch a, a gas cooler gas branch b, a suction gas reducing overheating branch c, a motor cooling branch f, a condenser 1, a gas cooler 2, a flash tank 3 and an oil separator 4.
The centrifugal compressor is taken as an example for explanation, the compressor comprises a low-pressure-stage impeller and a high-pressure-stage impeller, an input port of the low-pressure-stage impeller is an air suction port of the compressor, an output port of the low-pressure-stage impeller is connected with an input port of the high-pressure-stage impeller through a volute based on a fixed volume, and an output port of the high-pressure-stage impeller is an air exhaust port of the compressor. For screw compressors, the derivation of the calculation model is the same as for centrifugal compressors.
The exhaust port of the compressor is connected with the input port of the oil separator 4, the output port of the oil separator 4 is connected with the input port of the condenser 1, meanwhile, the output port of the oil separator 4 is also communicated with the input port of the flash evaporator 3 through a flash evaporator gas branch e, the output port of the oil separator 4 is also communicated with the gas input port of the gas cooler 2 through a gas cooler gas branch b, and the output port of the gas cooler 2 is communicated with the input port of the compressor. And the gas regulating valve 6 is connected in series on the flash tank gas branch e and the gas cooler gas branch b. The output port of the condenser 1 is communicated with the input port of the flash tank 3 through a flash tank liquid branch d, and the output port of the flash tank 3 is communicated with the input port of a high-pressure-stage impeller of the compressor through a tested compressor middle air supplement branch g, so that air supplement to a high-pressure end of the compressor is realized. The output port of the condenser 1 is communicated with the cooling liquid inlet of the gas cooler 2 through a gas cooler liquid branch a, and the output port of the condenser 1 is also communicated with the input port of the compressor through a suction gas reducing overheating branch c. The output port of the condenser 1 is also communicated with the inlet of the compressor by a motor cooling branch f for cooling the compressor. And the gas cooler liquid branch a, the suction reducing overheating branch c, the flash tank liquid branch d and the motor cooling branch f are all connected in series with a liquid regulating valve 5. The liquid regulating valve 5 connected in series on the gas cooler liquid branch a is a gas cooler liquid regulating valve, the liquid regulating valve 5 connected in series on the reduced air suction overheating branch c is a reduced air suction overheating branch regulating valve, the liquid regulating valve 5 connected in series on the flash evaporator liquid branch d is a flash evaporator liquid regulating valve, the liquid regulating valve 5 connected in series on the motor cooling branch f is a motor cooling branch regulating valve, the gas regulating valve 6 connected in series on the gas cooler gas branch b is a gas cooler gas regulating valve, and the gas regulating valve 6 connected in series on the flash evaporator gas branch e is a flash evaporator gas regulating valve.
In another embodiment of the solution, the testing device does not comprise a motor cooling branch f.
According to the related intermediate air supply-based compressor refrigerating capacity expression of the JB/T12843 & 2016 centrifugal refrigerant compressor, the mass flow of the refrigerant at the evaporation side, namely the suction mass flow m of the compressor is obtainedcom,iSpecifically, the above formula (12) is referred to.
The state point after the first-stage compression of the compressor is point 2' of fig. 4, and the specific enthalpy h after the first-stage compression of the compressor1st,oCalculated according to the above equation (11), wherein:
isentropic specific enthalpy h after first-stage compression of compressor1st,o,sComprises the following steps: h is1st,o,s＝f(p1st,scom,i)；
The compressor suction state point is point 1' of FIG. 4, and the compressor suction specific entropy scom,iComprises the following steps: scom,i＝f(pe,Te+ΔTcom,i)；
Compressor specific suction enthalpy hcom,iComprises the following steps: h iscom,i＝f(pe,Te+ΔTcom,i)；
Specific volume v of refrigerant after one-stage compression of compressor1st,oComprises the following steps: v. of1st,o＝f(p1st,h1st,o)；
Wherein p is1stThe unit is kPa of the pressure after the first-stage compression of the compressor; t iseTo the evaporation pressure peThe corresponding saturation temperature, i.e. the evaporation temperature; h is1st,o,s、scom,i、hcom,iCan be solved according to a refrigerant thermophysical property calculation program, and f () represents a refrigerant thermophysical property function.
The f () appearing in this embodiment is a refrigerant property function, and can be calculated from a look-up table, a Refprop software database, or a fitting equation.
Among the above parameters, hmid,lIndicating the compressor intermediate charge pressure pmidCorresponding saturated liquid specific enthalpy and intermediate air-supply pressure p of compressormidCompressor isentropic efficiency etasDesign value Q of refrigerating capacity of compressoreTesting working condition parameter evaporation pressure peAnd degree of superheat delta T of suction gas of compressorcom,iAll are input values, namely the design parameters of the tested compressor.
The state point of the middle air supplement port of the compressor and the state point of the outlet of the flash tank are the same as the point 10 'in the figure 4, the refrigerant from the middle air supplement is mixed with the refrigerant from the first-stage compression in the compressor to be tested, the mixed state point in the compressor to be tested is the point 3' in the figure 4, and the following steps are carried out according to the energy conservation:
(1+a)hmid,m＝h1st,o+ahmid,i (14)
specific enthalpy h after intermediate gas supplymid,mComprises the following steps:
hmid,m＝(h1st,o+ahmid,i)/(1+a) (15)
specific enthalpy h of the tested compressor middle air supplement portmid,iComprises the following steps: h ismid,i＝f(pmid,Tmid+ΔTmid)；
Wherein the relative air supplement rate a is an input value, TmidIs a compressor intermediatePressure p of air supplymidCorresponding saturation temperature, Δ TmidAnd supplying gas with superheat degree to the middle of the compressor. Parameters a, pmid、ΔTmidAll are input values, namely the design parameters of the tested compressor.
The first specific volume v of the mixture of the compressor intermediate air make-up refrigerant and the first-stage compressed refrigerantmid,mCan be obtained according to the above formula (9), wherein the pressure loss coefficient xi in the air supply processpIs an input value, namely the design parameter of the tested compressor.
For a centrifugal compressor, an air supplement port is generally arranged on a volute between two stages of impellers, and the specific volume of mixed refrigerant (i.e. the second specific volume after the middle air supplement refrigerant and the first-stage compressed refrigerant of the compressor are mixed) vmid,m' can be obtained according to the above formula (9).
In this embodiment, to obtain p1st、pmid,mAnd hmid,mFirst, a post-compression pressure p is given to the compressor1stBased on equation (11) to solve for h1st,oAnd further solving for v based on equation (9)mid,mAnd vmid,m', continuously modifying p in conjunction with the above steps S241-S24331stAnd solve for vmid,mAnd vmid,m', up to vmid,mAnd vmid,mThe absolute value of the difference of' is less than or equal to the convergence condition parameter epsilon. Where ε is the input value.
The compressor discharge state point is point 4' of FIG. 4, the compressor discharge specific enthalpy hcom,oCan be obtained by calculation according to the formula (10), and the isentropic specific enthalpy h of the exhaust gas of the compressorcom,o,s＝f(pc,smid,m) (ii) a Condensing pressure pcIs an input value, namely the design parameter of the tested compressor.
Specific entropy s of mixed intermediate air-supplementing refrigerant and first-stage compressed refrigerant of compressormid,m＝f(pmid,m,hmid,m)。
The motor of the tested compressor has two conditions of a motor cooling branch and a non-motor cooling branch in the test, and the calculation derivation of the two conditions is discussed in detail respectively as follows:
(1) when there is no motor cooling branch, the calculation is derived as follows:
the exit state point of the gas cooler liquid branch a is point 7 ' in fig. 4, the exit state point of the gas cooler gas branch b is point 9 ' in fig. 4, and in fig. 3, the refrigerants from the gas cooler liquid branch a and the gas cooler gas branch b are mixed in the gas cooler to form superheated vapor, and are discharged from the gas cooler exit to reach the compressor suction port, and at this time, the gas cooler exit state point is the compressor suction state point (point 1 ' in fig. 4), and according to the energy conservation:
mahc,l+mbhcom,o＝(ma+mb)hgl,o (16)
converting the above formula into the following expression form:
ma(hc,l-hgl,o)+mb(hcom,o-hgl,o)＝0 (17)
the state point of the saturated liquid of the refrigerant in the condenser is point 5' of fig. 4, and the specific enthalpy h of the saturated liquid of the refrigerant in the condenserc,lComprises the following steps: h isc,l＝f(pc) (18)
Specific enthalpy h at the outlet of the gas coolergl,oIs h isgl,o＝f(pe,Te+ΔTgl,o) (19)
Wherein the condensing pressure pcEvaporation pressure peEvaporation temperature TeSuperheat degree delta T of gas cooler outletgl,oAre all input values, namely the design parameters of the tested compressor.
The liquid injected into the suction pipe by the suction superheat reducing branch c is mixed with the refrigerant gas from the gas cooler, while the specific enthalpy of the refrigerant from the suction superheat reducing branch c is increased from hc,lIs raised to hcom,iOn the other hand, the specific enthalpy of the refrigerant from the gas cooler is hgl,oIs reduced to hcom,iTherefore, there is conservation of energy:
mc(hcom,i-hc,l)＝(ma+mb)(hgl,o-hcom,i) (20)
converting the above formula into the following expression form:
(ma+mb)hgl,o+mchc,l＝(ma+mb+mc)hcom,i (21)
in terms of mass conservation, the sum of the refrigerant mass flows from the gas cooler liquid branch a, the gas cooler gas branch b, and the reduced suction superheat branch c in fig. 3 is equal to the compressor suction mass flow, and therefore there is:
ma+mb+mc＝mcom,i (22)
in conjunction with equation (22), equation (21) can be transformed into:
mahgl,o+mbhgl,o+mchc,l＝mcom,ihcom,i (23)
the simultaneous equations (17), (22), (23) form a system of equations, which can be expressed as:
Figure BDA0003199387840000171
in equation (24), the gas cooler liquid inlet mass flow maMass flow m of gas inlet of gas coolerbMass flow m of the suction overheating branch is reducedcAre all unknowns to be solved.
(2) When the motor cooling branch exists, the calculation formula is derived as follows:
in fig. 3, the liquid in the reduced suction superheat branch c is mixed with the refrigerant superheated vapor from the gas cooler, and the corresponding energy conservation equation is:
mc(hsuc-hc,l)＝(ma+mb)(hgl,o-hsuc) (25)
in the formula, hsucTo reduce the specific enthalpy of the refrigerant in the suction tube before the suction superheat branch c.
Equation (25) can be transformed into:
mahgl,o+mbhgl,o+mchc,l＝(ma+mb+mc)hsuc (26)
in fig. 3, the refrigerant at the outlet of the motor cooling branch f is mixed with the refrigerant in the suction line and finally reaches the suction port of the compressor, and the corresponding energy conservation equation is as follows:
mfhf,o+(ma+mb+mc)hsuc＝mcom,ihcom,i (27)
the liquid refrigerant of the motor cooling branch f is used to cool the motor, so there is an energy balance:
mf(hf,o-hc,l)＝Qcool (28)
specific enthalpy h of outlet of motor cooling branchf,oComprises the following steps:
hf,o＝f(pe,Te+ΔTf,o) (29)
wherein the evaporation pressure peEvaporation temperature TeAnd the superheat degree delta T of the outlet of the motor cooling branchf,oAll are input values, namely the design parameters of the tested compressor.
According to equation (28), the mass flow rate of the suction superheat branch f is reduced as:
mf＝Qcool/(hf,o-hc,l) (30)
wherein the motor cools the heat load QcoolIs an input value, namely the design parameter of the tested compressor.
Combining equation (26) and equation (27) can derive:
mahgl,o+mbhgl,o+mchc,l＝mcom,ihcom,i-mfhf,o (31)
in terms of mass conservation, the sum of the refrigerant flows from the gas cooler liquid branch a, the gas cooler gas branch b, the reduced suction superheat branch c, and the motor cooling branch f in fig. 3 is equal to the compressor suction mass flow, and therefore there is:
ma+mb+mc＝mcom,i-mf (32)
for a gas cooler, equation (17) is still true, and the set of equations for simultaneous equations (17), (31), and (32) can be:
Figure BDA0003199387840000181
in equation (33), gas cooler liquid inlet mass flow maMass flow m of gas inlet of gas coolerbMass flow m of the suction overheating branch is reducedcAre all unknowns to be solved.
(3) The mass flow calculation model of each branch of the main loop circulation of the testing device is derived as follows:
by comparing the equation set (24) with the equation set (33), it is found that the following matrix form can be uniformly expressed:
MH＝A (34)
wherein the content of the first and second substances,
Figure BDA0003199387840000182
when there is no cooling of the electric machine,
Figure BDA0003199387840000183
when the cooling of the electric machine is taken into account,
Figure BDA0003199387840000184
based on equation (34), the mass flow rates of the gas cooler liquid branch a, the gas cooler gas branch b, and the reduced suction superheat branch c can be calculated according to the following equations:
M＝AH-1 (35)
second, the air supply loop circulation of the testing device
As shown in fig. 3, the branch of the test unit air make-up circuit cycle comprises: a flash tank liquid branch d and a flash tank gas branch e.
The outlet state point of the liquid branch of the flash tank is shown as a figure4, the flash tank gas branch exit state point is point 8 'of fig. 4, the refrigerants from flash tank liquid branch d and flash tank gas branch e of fig. 3 are mixed in the flash tank, the flash tank exit state point after mixing is point 10' of fig. 4, the flash tank liquid inlet mass flow rate mdMass flow m of gas inlet of flash tankeThe formula of (c) is derived as follows:
like the gas cooler, the energy conservation of the flash tank is expressed as:
mdhc,l+mehcom,o＝(md+me)hmid,i (36)
formula (36) may be expressed as: m isd(hc,l-hmid,i)+me(hcom,o-hmid,i)＝0 (37)
According to the conservation of mass, the sum of the refrigerant flow rates from the flash tank liquid branch d and the flash tank gas branch e is equal to the compressor air supply mass flow rate mmidThus, there are: m isd+me＝mmid (39)
Air supply mass flow of the compressor: m ismid＝amcom,i (40)
Wherein a represents the relative air supplement rate of the compressor.
The matrix form is available based on equation (37), equation (39), and equation (40):
MmidHmid＝Amid (41)
wherein the content of the first and second substances,
Figure BDA0003199387840000191
mass flow m of liquid inlet of flash tankdMass flow m of gas inlet of flash tankeIt can be calculated according to the following formula: mmid＝AmidHmid -1 (42)
In order to make the technical solutions provided by the present invention more clearly understood, the following detailed description is given with reference to specific embodiments.
Example 1
In this embodiment, the working medium of the tested compressor is the refrigerant R134 a.
The first step is as follows: and obtaining a working condition running chart of the compressor.
The operating condition operation diagram is shown in fig. 2, and includes operating condition inflection points a1, a2, A3 and a4, and the condensing temperature, the condensing pressure, the evaporating temperature and the evaporating pressure corresponding to each operating condition inflection point are shown in table 2 below.
Table 2: the condensation temperature, condensation pressure, evaporation temperature and evaporation pressure corresponding to the inflection point of each working condition in FIG. 1
Figure BDA0003199387840000201
Note that the condensing pressure is the saturation pressure corresponding to the condensing temperature, and the evaporating pressure is the saturation pressure corresponding to the evaporating temperature.
The second step is that: and calculating the preliminary flow coefficient of each working condition inflection point, wherein the working condition inflection point A1 is specifically used as an explanatory object in the embodiment, and introduction of the calculation process is carried out.
The compressor performance testing device in this embodiment has a motor cooling branch, so the compressor design parameters include: compressor intermediate air supply pressure pmidCompressor middle air supply superheat degree delta TmidCompressor suction superheat degree delta Tcom,iDesign value of cooling capacity QeIsentropic efficiency ηsRelative air supplement rate a and superheat degree delta T of outlet of gas coolergl,oCompressor motor cooling heat load QcoolSuperheat degree delta T of motor cooling branch outletf,o。
In the embodiment, the refrigerating capacity design value Q is used at the working condition inflection point A1eAnd isentropic efficiency ηsAs the variable parameters, the value ranges of the variable parameters are shown in table 3 below; the fixed parameters include: compressor intermediate air supplement pressure pmidCompressor middle air supply superheat degree delta TmidCompressor suction superheat degree delta Tcom,iRelative air supplement rate a and superheat degree delta T of outlet of gas coolergl,oCompressor motor cooling heat load QcoolMotor cooling branch outletDegree of superheat Δ Tf,oThe values of the fixed parameters are shown in table 4 below.
Table 3: value of variable parameter
Figure BDA0003199387840000211
In the above table, the first and second sheets,
Figure BDA0003199387840000212
represents the upper limit value of the g variable parameter under the k working condition inflection point state,
Figure BDA0003199387840000213
and the lower limit value of the g variable parameter in the k working condition inflection point state is shown.
It can be seen that, in this embodiment, for the operating condition inflection point a1, there are 4 combinations of the extreme values of the variable parameters, that is, the extreme value sets of the variable parameters are exhaustive as follows:
Figure BDA0003199387840000214
table 4: fixed parameter value
Figure BDA0003199387840000215
Figure BDA0003199387840000221
In this embodiment, for the operating condition inflection point a1, there are the following 4 parameter sets:
Figure BDA0003199387840000222
Figure BDA0003199387840000223
Figure BDA0003199387840000224
Figure BDA0003199387840000225
in this embodiment, p is the inflection point A1 of the operating conditionc＝0.77MPa，pe＝0.35MPa。
For any one of the above sets of parameters, in combination with equations (5) to (29) above, the corresponding gas cooler liquid inlet mass flow m can be calculatedaGas cooler gas inlet mass flow mbAnd the mass flow m of the suction overheating branch is reducedcMass flow m of liquid inlet of flash tankdMass flow m of gas inlet of flash tankeAnd mass flow m of motor cooling branchf。
M can then be combined according to the prior arta、mb、mc、md、meAnd mfCalculating the flow coefficient of the liquid regulating valve of the gas cooler, the flow coefficient of the gas regulating valve of the gas cooler, the flow coefficient of the suction superheat reducing regulating valve, the flow coefficient of the liquid regulating valve of the flash tank, the flow coefficient of the gas regulating valve of the flash tank and the flow coefficient of the cooling branch regulating valve of the motor.
The flow coefficient calculation formula of the regulating valve adopted in the embodiment is as follows:
Figure BDA0003199387840000226
wherein, CvalRepresenting the flow coefficient of the regulating valve, m representing the mass flow of the branch in which the flow valve is located, pinThe density of the refrigerant at the inlet of the valve is adjusted, and delta p is the pressure difference between the inlet and the outlet at the front valve and the rear valve.
In this embodiment, the jth parameter set pair in the ith working condition inflection point state is usedMass flow m of liquid inlet of gas cooleraGas cooler gas inlet mass flow mbAnd the mass flow m of the suction overheating branch is reducedcMass flow m of liquid inlet of flash tankdMass flow m of gas inlet of flash tankeMass flow m of motor cooling branchfRespectively recording as:
Figure BDA0003199387840000227
the flow coefficient of the gas cooler liquid regulating valve, the flow coefficient of the gas cooler gas regulating valve, the flow coefficient of the reduced suction superheat regulating valve, the flow coefficient of the flash tank liquid regulating valve, the flow coefficient of the flash tank gas regulating valve and the flow coefficient of the motor cooling branch regulating valve corresponding to the jth parameter set in the ith working condition inflection point state are respectively recorded as follows:
Figure BDA0003199387840000231
thus, there are
Figure BDA0003199387840000232
1≤r≤6。
For the present embodiment, i is 1. ltoreq. 4, and j is 1. ltoreq. 4.
For gas cooler gas regulating valve, gas cooler liquid regulating valve, reduce and breathe in overheated regulating valve, motor cooling branch road regulating valve: Δ p ═ pc-pe；pcDenotes the condensation pressure, peRepresents the evaporation pressure;
for flash tank gas regulating valves, flash tank liquid regulating valves: Δ p ═ pc-pmid，pmidRepresenting the compressor intermediate make-up air pressure.
ρinFor gas cooler gas regulating valves, flash tank gas regulating valves, to regulate the valve inlet refrigerant density, the valve inlet refrigerant state is superheated steam, ρinAccording to the condensing pressure pcAnd compressor specific enthalpy determination hcom,oI.e. pin＝f(pc,hcom,o) (ii) a For gas cooler liquid regulating valve, suction superheat reducing regulating valve, flash tank liquid regulating valve and motor cooling branch regulating valve, the inlet state of the valve is refrigerant saturated liquid, rhoinTo the condensation pressure pcCorresponding saturated liquid density, i.e. pin＝f(pc)。
The valve flow coefficient is generally a function of the mass flow through the valve, the differential pressure before and after the valve, and the density of the refrigerant at the inlet of the valve. It should be noted that the calculation formula of the flow coefficient of the regulating valve does not necessarily need to be calculated according to the formula (43), and the specific formula form is different according to the type of the refrigerant, the structural type of the valve, and the working condition range, and the related flow coefficient calculation formula belongs to the known knowledge in the valve field, and is not described herein again.
In this embodiment, for each set of parameter set in the state of the operating condition inflection point a1, the finally obtained flow coefficient of each regulating valve to be calculated is as shown in the following table 5:
table 5: corresponding relation system of parameter set and flow coefficient of each regulating valve to be calculated under working condition inflection point A1 state
Figure BDA0003199387840000241
So, the flow coefficient of the gas cooler liquid regulating valve, the flow coefficient of the gas cooler gas regulating valve, the flow coefficient of the reducing air suction superheat regulating valve, the flow coefficient of the flash tank liquid regulating valve, the flow coefficient of the flash tank gas regulating valve and the flow coefficient of the motor cooling branch regulating valve obtained to operating condition inflection point A1 are respectively recorded as:
Figure BDA0003199387840000242
and
Figure BDA0003199387840000243
then:
Figure BDA0003199387840000244
Figure BDA0003199387840000245
Figure BDA0003199387840000246
Figure BDA0003199387840000247
Figure BDA0003199387840000248
Figure BDA0003199387840000249
wherein the content of the first and second substances,
Figure BDA00031993878400002410
respectively representing the maximum value and the minimum value of the flow coefficient of the gas cooler liquid regulating valve corresponding to the working condition inflection point A1;
Figure BDA00031993878400002411
respectively representing the maximum value and the minimum value of the flow coefficient of the gas cooler gas regulating valve corresponding to the working condition inflection point A1;
Figure BDA00031993878400002412
respectively representing the maximum value and the minimum value of the flow coefficient of the reduced suction superheat regulating valve corresponding to the working condition inflection point A1;
Figure BDA00031993878400002413
respectively representing the maximum value and the minimum value of the flow coefficient of the flash tank liquid regulating valve corresponding to the inflection point A1 of the working condition;
Figure BDA00031993878400002414
respectively representing the maximum value and the minimum value of the flow coefficient of the flash tank gas regulating valve corresponding to the working condition inflection point A1;
Figure BDA0003199387840000251
respectively representing the maximum value and the minimum value of the flow coefficient of the motor cooling branch regulating valve corresponding to the working condition inflection point A1.
By referring to the calculation method of the maximum value and the minimum value of the flow coefficient of each regulating valve to be calculated corresponding to the working condition inflection point A1, the maximum value and the minimum value of the flow coefficient of each regulating valve to be calculated corresponding to the working condition inflection point A2, the working condition inflection point A3 and the working condition inflection point A4 can be obtained.
In this embodiment, the maximum value and the minimum value of the flow coefficient of each regulating valve to be calculated, which correspond to each operating condition inflection point, are shown in table 6 below.
Table 6: the maximum value and the minimum value of the flow coefficient of each regulating valve to be calculated, which correspond to the working condition inflection points A1-A4
Figure BDA0003199387840000252
Thus, in this embodiment, the value range C of the finally obtained flow coefficient of the liquid regulating valve of the gas cooler1Flow coefficient C of gas regulating valve of gas cooler2Reducing the flow coefficient C of the suction superheat regulating valve3Flow coefficient C of flash tank liquid regulating valve4Flow coefficient C of gas regulating valve of flash evaporator5Flow coefficient C of motor cooling branch regulating valve6Respectively as follows:
C1＝[Cmin，1,Cmax，1]；C2＝[Cmin，2,Cmax,2]；C3＝[Cmin,3,Cmax,3]
C4＝[Cmin,4,Cmax,4]；C5＝[Cmin,5,Cmax,5]；C6＝[Cmin,6,Cmax,6]
Figure BDA0003199387840000261
Figure BDA0003199387840000262
Figure BDA0003199387840000263
Figure BDA0003199387840000268
Figure BDA0003199387840000264
Figure BDA0003199387840000265
wherein, Cmin,rLower limit value C representing the flow coefficient of the r-th regulating valve to be calculatedmax,rRepresents the upper limit value of the flow coefficient of the regulating valve to be calculated,
Figure BDA0003199387840000266
the lower limit value of the flow coefficient of the regulation valve to be calculated in the r-th working condition inflection point i state is shown,
Figure BDA0003199387840000267
and the upper limit value of the flow coefficient of the r-th regulating valve to be calculated in the state of the working condition inflection point i is represented.
The invention is not to be considered as limited to the specific embodiments shown and described, but is to be understood to cover all modifications, equivalents, and alternatives falling within the spirit and scope of the invention as defined by the appended claims.

Claims (10)

1. A method for calculating the flow coefficient of a regulating valve of a refrigeration compressor performance testing device is characterized by comprising the following steps:
s1, aiming at performance design parameters of the refrigeration compressor, establishing a working condition running chart of the compressor in a two-dimensional coordinate system which takes evaporation pressure as an abscissa and condensation pressure as an ordinate, and obtaining a working condition inflection point on the working condition running chart, wherein the working condition inflection point is a point where any two line segments on the working condition running chart intersect; the condensation pressure and the evaporation pressure corresponding to the working condition inflection point are performance design parameters of the refrigeration compressor corresponding to the working condition inflection point;
s2, obtaining variable parameters and fixed parameters corresponding to the inflection points of the working conditions; the variable parameters are compressor design parameters with a variable range, and the fixed parameters are compressor design parameters with fixed values; exhaustively exhausting a variable parameter extreme value set, wherein each variable parameter extreme value set comprises the maximum value or the minimum value of each variable parameter; setting parameter sets corresponding to the variable parameter extreme value sets one by one, wherein the parameter sets are combined into a corresponding variable parameter extreme value set and a union set of all fixed parameters; calculating the flow coefficient of each regulating valve to be calculated corresponding to the working condition inflection point by combining each parameter set; aiming at each regulating valve to be calculated, acquiring the maximum value and the minimum value in the flow coefficient of the regulating valve to be calculated corresponding to each parameter set, and respectively taking the maximum value and the minimum value as the initial upper limit value and the initial lower limit value of the flow coefficient corresponding to the regulating valve to be calculated;
s3, obtaining the maximum value of the initial upper limit values of the flow coefficient of the regulating valve to be calculated under different working condition inflection points as the upper limit value of the flow coefficient of the regulating valve to be calculated, and obtaining the minimum value of the initial lower limit values of the flow coefficient of the regulating valve to be calculated under different working condition inflection points as the lower limit value of the flow coefficient of the regulating valve to be calculated.
2. The method for calculating the flow coefficient of the regulating valve of the performance testing device of the refrigeration compressor as claimed in claim 1, wherein the step S2 specifically comprises the following steps:
s21, obtaining the correspondence of the working condition inflection point rCondensing pressure p ofcAnd the evaporation pressure peDividing the compressor design parameters corresponding to the working condition inflection points r into variable parameters and fixed parameters, wherein the variable parameters comprise:
Figure FDA0003199387830000011
Figure FDA0003199387830000012
i-th of the operating condition inflection point r staterIndividual variable parameter, 1 ≦ ir≤mr，mrFor the number of the variable parameters to be,
Figure FDA0003199387830000013
Figure FDA0003199387830000014
and
Figure FDA0003199387830000015
respectively represent
Figure FDA0003199387830000016
The upper limit value and the lower limit value of the value range; the fixed parameters include:
Figure FDA0003199387830000017
Figure FDA0003199387830000018
j represents the operating point r staterA fixed parameter, j is more than or equal to 1r≤nr；nrIs the number of fixed parameters;
s22 set of extreme values of exhaustive variable parameters
Figure FDA0003199387830000021
At least one parameter of any two extreme value sets of variable parameters is different;
order to
Figure FDA0003199387830000022
Or
Figure FDA0003199387830000023
Figure FDA0003199387830000024
To represent
Figure FDA0003199387830000025
Upper limit of value range
Figure FDA0003199387830000026
Or a lower limit value
Figure FDA0003199387830000027
Any set of variable parameter extrema is written as:
Figure FDA0003199387830000028
s23, setting parameter sets corresponding to the variable parameter extreme value sets one by one
Figure FDA0003199387830000029
Figure FDA00031993878300000210
krThe sequence number is shown to indicate that,
Figure FDA00031993878300000211
and the quantity of the variable parameter extreme value sets corresponding to the working condition inflection points r is represented, and is equal to the quantity of the parameter sets.
S24, calculating the flow coefficient of each regulating valve to be calculated in the compressor test loop under the condition of a working condition inflection point r according to each group of parameter set, and respectively recording the maximum value and the minimum value of the flow coefficient of each regulating valve to be calculated as the initial upper limit value and the initial lower limit value of the flow coefficient corresponding to the regulating valve;
and S25, repeating the steps S21 to S24, and calculating the initial upper limit value and the initial lower limit value of the flow coefficient corresponding to each regulating valve to be calculated under each working condition inflection point.
3. The method for calculating the flow coefficient of the regulating valve of the performance testing device of the refrigeration compressor as recited in claim 2, wherein the compressor design parameters further comprise: compressor intermediate air supply pressure pmidCompressor middle air supply superheat degree delta TmidCompressor suction superheat degree delta Tcom,iDesign value of cooling capacity QeIsentropic efficiency ηsAnd pressure loss coefficient xi in air supply processpRelative air supplement rate a and superheat degree delta T of outlet of gas coolergl,o(ii) a If the motor cooling branch exists, the method further comprises the following steps: compressor motor cooling heat load QcoolSuperheat degree delta T of motor cooling branch outletf,o。
4. The method for calculating the flow coefficient of a control valve of a performance testing apparatus of a refrigerating compressor according to claim 3, wherein the control valves to be calculated include a gas cooler gas control valve, a gas cooler liquid control valve, a suction superheat reduction control valve, a flash tank gas control valve and a flash tank liquid control valve; if the motor cooling branch exists, the regulating valve to be calculated further comprises a motor cooling branch regulating valve.
5. The method for calculating the flow coefficient of the regulating valve of the performance testing device of the refrigeration compressor as claimed in claim 1, wherein in step S24, the method for calculating the flow coefficient of each regulating valve to be calculated in the compressor testing loop under the working condition inflection point state according to any set of parameter sets comprises the following steps:
s241, calculating the pressure p after the first-stage compression of the compressor according to the parameter set1st，p1st＝(p1st,upper+p1st,low)/2；
Wherein p is1st,upperIs the upper limit value, p, of the pressure after one-stage compression of the compressor1st,lowThe lower limit value of the pressure after the first-stage compression of the compressor; p is a radical of1st,upperIs the compressor intermediate make-up air pressure pmid，p1st,lowIs 0;
s242, calculating a first specific volume v after the compressor intermediate air supply refrigerant and the first-stage compressed refrigerant are mixedmid,mAnd the second specific volume v of the mixture of the compressor intermediate air make-up refrigerant and the refrigerant after the first-stage compressionmid,m'；
vmid,m＝f(pmid,m,hmid,m)；
Figure FDA0003199387830000031
Wherein f () represents a functional relationship of physical properties of the refrigerant, pmid,mThe pressure h after the intermediate air make-up refrigerant of the compressor is mixed with the refrigerant after the first-stage compressionmid,mThe specific enthalpy of the mixed intermediate air-replenishing refrigerant and the first-stage compressed refrigerant of the compressor is represented; v. of1st,oThe specific volume of the refrigerant after the first-stage compression of the compressor is shown, and a shows the relative air supplement rate of the compressor;
s243, when | vmid,m-vmid,m' | > ε if vmid,m＞vmid,m', then set p1st,upper＝p1stAnd then returns to step S241 to recalculate the compressor one-stage compression pressure p1st(ii) a If v ismid,m＜vmid,m', then set p1st,low＝p1stAnd then returns to step S241 to recalculate the compressor one-stage compression pressure p1st(ii) a Epsilon is a convergence condition parameter;
s244, when | vmid,m-vmid,mWhen | ≦ epsilon, calculating the specific enthalpy h of the compressor exhaustcom,o；
S245, calculating the mass flow m of the liquid inlet of the gas cooleraGas cooler gas inlet mass flow mbReduce the branch quality of the suction overheatingFlow rate mc；
When the performance testing device of the refrigeration compressor does not have a motor cooling branch, a calculation model based on matrix operation is as follows:
Figure FDA0003199387830000041
when the refrigerating compressor performance testing device is provided with a motor cooling branch, a calculation model based on matrix operation is as follows:
Figure FDA0003199387830000042
wherein m iscom,iIs the compressor suction mass flow, hcom,iIs the compressor specific enthalpy of suction, hc,lSpecific enthalpy, h, of saturated liquid of refrigerant corresponding to condensing pressuregl,oIs the specific enthalpy of the outlet of the gas cooler, hgl,oAccording to the evaporation pressure peAnd gas cooler outlet superheat degree delta Tgl,oDetermination of hcom,oSpecific enthalpy of compressor discharge, hf,oIs the specific enthalpy of the outlet of the cooling branch of the motor, mfMass flow rate of the motor cooling branch;
s246, the specific enthalpy h of the saturated liquid of the refrigerant corresponding to the condensing pressurec,lCompressor middle air supply inlet specific enthalpy hmid,iSpecific enthalpy h of compressor dischargecom,oCompressor suction mass flow mcom,iCalculating the mass flow m of the liquid inlet of the flash tankdAnd the mass flow m of the gas inlet of the flash tankeThe calculation model based on matrix operation is as follows:
Figure FDA0003199387830000043
wherein a represents the relative air supplement rate of the compressor;
s247, according to the condensing pressure pcEvaporation pressure peAnd gas coolingMass flow m of liquid inlet of deviceaCalculating the flow coefficient of the liquid regulating valve of the gas cooler; according to the condensing pressure pcEvaporation pressure peAnd gas cooler gas inlet mass flow mbCalculating the flow coefficient of the gas regulating valve of the gas cooler; according to the condensing pressure pcEvaporation pressure peAnd reducing mass flow m of the suction superheat branchcCalculating a flow coefficient for reducing the suction overheating branch regulating valve; according to the condensing pressure pcEvaporation pressure peAnd the liquid inlet mass flow m of the flash tankdCalculating the flow coefficient of the liquid regulating valve of the flash tank; according to the condensing pressure pcEvaporation pressure peAnd the mass flow m of the gas inlet of the flash tankeCalculating the flow coefficient of the gas regulating valve of the flash tank; according to the condensing pressure pcEvaporation pressure peAnd mass flow m of motor cooling branchfAnd calculating the flow coefficient of the motor cooling branch regulating valve.
6. The method for calculating the flow coefficient of the control valve of the performance testing device of the refrigerating compressor as claimed in claim 5, wherein in step S242:
pmid,m＝p1st+(pmid-p1st)ξp；
hmid,m＝(h1st,o+ahmid,i)/(1+a)；
wherein p ismid,mThe pressure h after the intermediate air make-up refrigerant of the compressor is mixed with the refrigerant after the first-stage compressionmid,mThe specific enthalpy of the mixed intermediate air-replenishing refrigerant and the first-stage compressed refrigerant of the compressor is represented; p is a radical of1stThe pressure after the first stage of compression of the compressor is represented; a represents the relative air supplement rate of the compressor; p is a radical ofmidIndicating the intermediate supply pressure, ξ of the compressorpRepresenting the pressure loss coefficient of the air replenishing process; h is1st,oRepresents the specific enthalpy, h, of the compressor after the first stage of compressionmid,iRepresents the specific enthalpy, h, of the intermediate charge air inlet of the compressormid,iAccording to the intermediate air supply pressure p of the compressormidAnd compressor middle air supply superheat degree delta TmidAnd (4) determining.
7. The method for calculating the flow coefficient of the regulating valve of the performance testing device of the refrigeration compressor as recited in claim 6, wherein:
Figure FDA0003199387830000061
wherein h iscom,oSpecific enthalpy of compressor discharge, hcom,o,sRepresenting the isentropic specific enthalpy of compressor discharge; compressor exhaust isentropic specific enthalpy hcom,o,sAccording to the condensing pressure p based on the physical function relationship of the refrigerantcAnd the specific entropy s of the mixture of the intermediate air-supplementing refrigerant of the compressor and the refrigerant after the first-stage compressionmid,mDetermination of smid,mAccording to the pressure p after the intermediate air-supplementing refrigerant of the compressor is mixed with the refrigerant after the first-stage compressionmid,mAnd the specific enthalpy h of the mixed intermediate air-replenishing refrigerant and first-stage compressed refrigerant of the compressormid,mDetermining; etasIndicating compressor isentropic efficiency.
8. The method for calculating the flow coefficient of the regulating valve of the performance testing device of the refrigeration compressor as recited in claim 6, wherein:
Figure FDA0003199387830000062
wherein h is1st,oRepresents the specific enthalpy, h, of the compressor after the first stage of compression1st,o,sRepresents the isentropic specific enthalpy, h, of the compressor after the first-stage compression1st,o,sBased on the physical function relation of the refrigerant, the pressure p after the first-stage compression of the compressor is obtained1stAnd compressor specific suction entropy scom,iDetermining; h iscom,iExpressing the specific enthalpy, eta, of the compressor suctionsRepresenting compressor isentropic efficiency, scom,iAnd hcom,iAre all according to the evaporation pressure peAnd degree of superheat delta T of suction gas of compressorcom,iAnd (4) determining.
9. The method for calculating the flow coefficient of the control valve of the performance testing device of the refrigerant compressor as set forth in claim 5The method is characterized in that: m iscom,i＝Qe/(hcom,i-hmid,l)；
Wherein m iscom,iRepresenting compressor suction mass flow, hcom,iRepresents compressor suction specific enthalpy; h ismid,lIndicating the compressor intermediate charge pressure pmidCorresponding specific enthalpy of saturated liquid, hmid,lAccording to the compressor intermediate air supplement pressure p based on the physical function relation of the refrigerantmidDetermining; qeIndicating the refrigerating capacity of the compressor, QeIs a design value.
10. The method of branch mass flow calculation for a cold compressor performance testing apparatus of claim 5, wherein: m isf＝Qcool/(hf,o-hc,l)；
Wherein m isfRepresenting the mass flow, Q, of the cooling branch of the motorcoolRepresents the cooling heat load of the motor, hf,oAccording to the evaporation pressure p based on the physical function relationship of the refrigeranteAnd superheat degree delta T of motor cooling branch outletf,oDetermination of hc,lIndicating the specific enthalpy of the saturated liquid of the refrigerant corresponding to the condensing pressure.
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